OopightW COPYRIGHT DEPOSIT. Steam Power Plant Engineering BY G. F. GEBHARDT PROFESSOR OF MECHANICAL ENGINEERING, ARMOUR INSTITUTE OF TECHNOLOGY CHICAGO, ILL. THIRD EDITION, REVISED AND ENLARGED FIRST THOUSAND NEW YORK JOHN WILEY & SONS London: CHAPMAN & HALL, Limited 1910 r Copyright, 1908, 1910, BY G. F. GEBHARDT d Stanhope lpress P. H. GILSON COMPANY BOSTON. U.S.A. CI.A2732G6 PREFACE This book is the outcome of a series of lectures delivered to the Senior class of the Armour Institute of Technology, Chicago, 111. It is primarily intended as a text-book for engineering students, but, it is hoped, will also be of interest to practicing engineers. The field embraced by the title is a large one and it has been necessary to limit the treatment to essential elements. Much of the matter contained in the author's original notes, including that relat- ing to steam engine design, valve gears, steam boiler design, and the like, has therefore been omitted. The numerous references appear- ing throughout the text and the appended bibliographies, which have been carefully compiled, are depended upon to extend the scope of the work. The standard codes of the American Society of Mechanical Engineers for conducting engine and boiler trials are in frequent demand by engineers and have therefore been included as an appendix. Authorities have been freely consulted and extensive use made of current engineering literature, due acknowledgment being made by footnote or reference whenever possible. The matter included is representative of American practice and no effort has been made to include any other except in a few special cases. The author wishes to express his obligations to Prof. Raymond Burnham for many valuable suggestions and corrections, and to Mrs. Julia Beveridge, librarian at Armour Institute, for assistance in compiling references. iii PREFACE TO SECOND EDITION A number of additional changes have been made to bring this work into accord with more recent practice. All the typographical and other errors discovered in the first edition have been corrected. PREFACE TO THIRD EDITION All obsolete matter has been discarded, considerable new material has been added throughout the book, and many of the chapters have been entirely rewritten. IV CONTENTS Page CHAPTER I. — Elementary Steam Power Plants 1-13 1. General 1 2. Elementary Non-Condensing Plant 2 3. Non-Condensing Plant. Exhaust Steam Heating 5 4. Elementary Condensing Plant 7 5. Condensing Plant with Full Complement of Heat-Saving Appliances . . 10 CHAPTER II. — Fuels and Combustion 14-67 6. General 14 7. Classification of Fuels 14 8. Solid Fuels 14 9. Coal 15 10. Anthracite 15 11. Semi-Anthracite 16 12. Semi-Bituminous 16 13. Bituminous 16 14. Lignite ■. 17 15. Peat or Turf 18 16. Wood, Straw, Sawdust, Bagasse, Tanbark 19 17. Composition of Coal 24 18. Combustion 24 19. Temperature due to Combustion 27 20. Air Required for Combustion 28 21. Calorific Value of Coal 31 22. Heat Losses in Burning Coal 32 23. Loss in the Dry Chimney Gases 33 24. Loss due to Incomplete Combustion 36 25. Loss of Fuel through Grate 37 26. Superheating the Moisture in the Air 37 27. Loss due to Moisture in the Fuel 37 28. Loss due to Presence of Hydrogen in the Fuel 38 29. Loss due to Visible Smoke 38 30. Radiation and Minor Losses 39 31. Size of Coal 39 32. Washed Coai 40 33. Purchasing Coal • 42 34. Powdered Coal 44 35. Depreciation of Powdered-Coal Furnaces 44 36. Storing Powdered Fuel 45 37. Rate of Combustion with Powdered Coal 45 38. Cost of Pulverizing Coal 45 39. Efficiency of Powdered-Coal Furnaces 46 v vi CONTENTS CHAPTER II — Continued P AGE 40. Furnaces for Burning Powdered Coal 47 41. Draft for Powdered Fuel 47 42. Types of Powdered-Coal Burners 47 43. Pinther Apparatus for Burning Powdered Coal 48 44. Schwartzkopff Apparatus for Burning Powdered Coal 49 45. Aero-Pulverizer Apparatus for Burning Powdered Coal 49 46. Triumph Apparatus for Burning Powdered Coal 50 47. Fuel Oil 51 48. Chemical and Physical Properties of Fuel Oil 52 49. Efficiency of Boilers with Fuel Oil 54 50. Comparative Evaporative Economy of Oil and Coal 55 51. Types of Oil Burners 55 52. Furnaces for Burning Oil Fuel 59 53. Air vs. Steam as an Atomizing Medium 60 54. Oil Pressure 63 55. Oil Storage and Transportation 64 56. Conclusions of the U. S. Naval Liquid Fuel Board 65 57. Gaseous Fuels 66 CHAPTER III. — Boilers 66-123 58. General. . 68 59. Classification 68 60. Vertical Tubular Boilers 68 61. Fire-Box Boilers 71 62. Scotch-Marine Boilers 72 63. Robb-Mumford Boiler 73 64. Horizontal Return Tubular Boiler 74 65. Babcock & Wilcox Boiler 81 66. Heine Boiler 82 67. Wickes Vertical Water-Tube Boiler 82 67a. Parker Boiler 86 68. Stirling Boiler 87 69. Unit of Evaporation 88 69a. Heat Transmission 90 70. Heating Surface 92 71. Horse Power of a Boiler 93 72. Grate Surface 95 73. Boiler and Furnace Efficiency . 98 74. Boiler Performances 99 75. Effect of Capacity on Efficiency 104 76. Thickness of Fire 109 77. Influence of Initial Temperature on Efficiency . . Ill 78. Cost of Boiler and Settings 112 79. Selection of Type , 112 80. Grates 114 81. Rocking Grates 115 82. Blow-Offs 116 83. Dampers 118 84. Water Gauge 119 CONTENTS vii CHAPTER III — Continued Page 85. Fusible Plugs . . 121 86. Mechanical Tube Cleaners . . 121 CHAPTER IV. — Smoke Prevention, Furnaces, Stokers ,124-151 87. General 124 88. Mechanical Stokers 125 89. Chain Grates 126 90. Step Grates, Front Feed 130 91. Step Grates, Side Feed 136 92. Underfeed Stokers 138 93. Down-Draft Furnaces 139 94. Sprinkling Furnaces 141 95. Dutch Ovens , 141 96. Twin-Fire Furnace 142 96a. Chicago Settings 143 97. Wooley Smokeless Furnace 149 98. Kent's Wing-Wall Furnace 149 99. Burke's Smokeless Furnace 151 100. Admission of Air above Fire 151 101. Cost of Stokers and Furnaces 151 CHAPTER V. — Superheated Steam; Superheaters 152-180 102. General. 152 103. Economy of Superheat 153 104. Limit of Superheat 154 105. Specific Heat of Superheated Steam 155 106. Types of Superheaters 162 107. Babcock & Wilcox Superheater 163 108. Stirling Superheater 164 109. Foster Superheater 165 110. Independently Fired Superheaters 166 111. Materials for Superheaters 170 112. Extent of Superheating Surface . . 170 113. Performance of Superheaters 174 113a. Properties of Superheated Steam 180 CHAPTER VI. — Coal and Ash-Handling Apparatus .181-206 114. General 181 115. Coal Storage . . ; .... 181 116. Coal Conveyors 183 117. Hand Shoveling . 183 118. Bucket Conveyors 184 119. Belt Conveyors 192 120. Elevating Tower, Hand-Car Distribution 193 121. Overhead Storage, Bucket Hoist 195 122. Elevating Tower, Cable-Car Distribution 196 123. "Vacuum " Ash-Handling System .".'..■ 196 124. Cost of Handling Coal and Ashes 201 125. Coal Hoppers 202 126. Coal Valves 205 vm CONTENTS Page CHAPTER VII. — Chimneys 207-244 127. Chimney Draft 207 128. Chimney Formulas 212 129. Height of Chimneys for Boilers Using Oil Fuel 218 130. Classification of Chimneys 218 131. Guyed Chimneys 219 132. Self-Sustaining Steel Chimneys 219 133. Thickness of Plates , 220 134. Riveting 223 135. Stability of Steel Chimneys 223 136. Brick Chimneys 224 137. Thickness of Walls 226 138. Core and Lining 230 139. Materials for Brick Chimneys 230 140. Stability of Brick Chimneys ; 231 141. Custodis Radial Brick Chimney 234 142. Steel-Concrete Chimneys 234 143. Breeching 240 144. Chimney Foundations 240 145. Chimney Efficiencies 241 146. Cost of Chimneys 243 CHAPTER VIII. — Mechanical Draft 245-266 147. General 245 148. Steam Jets 245 149. Parsons Smokeless Furnace 248 150. Heinrich Smokeless Furnace 249 151. Fan Draft 249 152. Theory of Fans 252 153. Determination of the Size of Fan 258 154. Chimney vs. Mechanical Draft 261 155. Balanced Draft 264 CHAPTER IX. — Steam Engines 267-326 156. Introductory 267 157. The Ideal Engine 267 158. Thermal Efficiency of the Actual Engine 273 159. Mechanical Efficiency 275 160. Heat Losses in the Actual Engine 278 161. Loss due to Moisture in the Steam at Admission 278 162. Loss due to Leakage 279 163. Loss due to Cylinder Condensation 279 164. Loss due to Clearance Volume 281 165. Loss due to Incomplete Expansion and Compression 282 166. Loss due to Wire Drawing , 284 167. Loss due to Friction 284 168. Effect of Increasing Boiler Pressure 286 169. Receive r-Reheaters 287 170. Jackets 288 171. Single and Double Acting Engines 290 172. High and Low Speed Engines 290 CONTENTS ix CHAPTER IX — Continued Page 173-4. High-Speed Single-Valve Engines 291 175. High-Speed Multi- Valve Engines 297 176. Medium and Low Speed Engines 299 177. Compound Engines „ 300 178. Triple and Quadruple Engines 305 179. Influence of Condensing 307 180. Throttling vs. Automatic Cut-Off 310 181. Influence of Superheat 313 182. Binary Vapor Engines 321 183. Cost of Engines 326 CHAPTER X. — Steam Turbines 327-396 184. Classification 327 184a. General Elementary Theory 328 185. De Laval Turbine 331 186. Elementary Theory, De Laval Turbine .... 333 187. Terry Turbine 345 188. Kerr Turbine 346 189. Curtis Turbine 350 190. Elementary Theory, Curtis Turbine 358 191. Hamilton-Holzworth Turbine 362 192. Westinghouse-Parsons Turbine 365 192a. Allis Chalmers Turbine 372 193. Elementary Theory, Parsons Turbine 373 194. Low and Mixed Pressure Turbines 376 195. Advantages of the Steam Turbine 382 196. Simplicity 382 197. Economy of Space and Foundation 382 198. Absence of Oil in Condensed Steam 384 199. Regulation 384 200. Overload Capacity 384 201. Efficiency and Economy 384 202. First Cost 389 203. Cost of Operation 392 204. Influence of Superheat 393 205. Influence of High Vacua 395 CHAPTER XI. — Condensers 397-469 206. General 397 207. Function of the Condenser 398 208. Classification of Condensers 400 209. Common Jet Condenser 401 210. Condensing Water, Jet Condensers 404 211. Effect of Aqueous Vapor upon the Degree of Vacuum 405 212. Injection Orifice 407 213. Volume of Condenser Chamber 408 214. Injection and Discharge Pipes . . 408 215. Siphon Condensers 408 216. Size of Siphon Condensers 409 217. Ejector Condensers 410 x CONTENTS CHAPTER XI — Continued Page 218. Barometric Condensers 411 219. Water-Cooled Surface Condensers 416 220. Cooling Water, Surface Condensers 420 221. Extent of Water-Cooling Surface 421 222. Dry-Air Surface Condensers 428 223. Quantity of Air for Cooling (Dry-Air Condenser) 429 224. Saturated-Air Surface Condensers 430 225. Evaporative Surface Condensers 433 226. Location and Arrangement of Condensers 433 226a. Independent System . . . . 434 227. Central Condensing Systems 439 228. High- Vacuum Systems 441 229. Power Consumption of Condenser Auxiliaries . . . 447 230. Cost of Condensers 450 231. Most Economical Vacuum . 451 232. Choice of Condensers 452 233. Water-Cooling Systems 453 233a. Cooling Pond 454 233b. Spray Fountain .' 455 234. Cooling Towers . 456 235. Parallel Comparison of Fan and Natural-Draft Cooling Towers 460 236. Cooling-Tower Calculations 460 236a. Hygrometry . . . . . 468 237. Tests of Cooling Towers . 468 CHAPTER XII. — Feed-Water Purifiers and Heaters 471-522 238. General . . 471 239. Chemical Purification 476 240. Boiler Compounds 476 241. Use of Kerosene and Petroleum Oils in Boiler Feed Water 477 242. Use of Zinc in Boilers 478 243. Methods of Introducing Compounds 478 244. Weight of Compound Necessary 478 245. Mechanical Purification 479 246. Thermal Purification 479 247. Purifying Plants 480 248. Economy of Preheating Feed Water 484 249. Classification of Feed-Water Heaters 485 250. Open Heaters 486 251. Open Heaters and Purifiers 489 252. Temperatures in Open Heaters 489 253. Pan Surface Required in Open Heaters 491 254. Size of Shell, Open Heaters 491 255. Classification of Closed Heaters 492 256. Closed Heaters, Water-Tube : .' 493 257. Closed Heaters, Steam-Tube 494 258. Heating Surface, Closed Heaters 496 259. Heat Transmission, Closed Heaters 497 260. Open vs. Closed Heaters. 504 261. Through Heaters 505 CONTENTS xi CHAPTER XII — Continued Page 262. Induced Heaters 506 263. Live-Steam Heaters and Purifiers 507 264. Economizers ! 508 265. Value of Economizers 511 266. Factors Determining Installation of Economizers 512 267. Feed-Water Temperature due to Use of Economizers 512 268. Choice of Feed-Water Heating Systems 516 CHAPTER XIII. — Pumps 522-574 269. Classification of Pumps 522 270. Boiler-Feed Pumps, Direct-Acting Duplex 524 271. Boiler-Feed Pumps, Direct- Acting, Steam-Actuated Gear 527 272. Air and Vacuum Chambers 529 273. Water Pistons and Plungers 530 274. Performance of Piston Pumps 531 275. Size of Boiler-Feed Pump 539 276. Steam-Pump Governors „• 541 277. Feed- Water Regulators, Steam Pumps 541 278. Power Pumps 543 279. Injectors 545 280. Positive Injectors 547 281. Automatic Injectors 547 282. Performance of Injectors 548 283. Injector vs. Steam Pump as a Boiler Feeder 548 284. Air Pumps 552 285. Dean Wet- Air Pump 552 286. Size of Wet-Air Pumps for Jet Condensers 553 287. Edwards Air Pump 555 288. Mullan Valveless Air Pump 556 289. Alberger Rotative Dry-Air Pump 557 290. Size of Wet-Air Pumps for Surface Condensers 558 291. Size of Dry- Air Pumps for Surface Condensers 558 292. Centrifugal Pumps 560 292a. Hot-Well Pumps 560 293. Volute Centrifugal Pumps 561 294. Turbine Centrifugal Pumps 561 295. Performance of Centrifugal Pumps 563 296. Rotary Pumps 567 297. Circulating Pumps 572 298. Air Lift 572 CHAPTER XIV. — Separators, Traps, and Drains 575-605 299. Live-Steam Separators , 575 300. Classification of Separators 576 301. Reverse-Current Steam Separators 577 302. Centrifugal Steam Separators 578 303. Baffle-Plate Steam Separators 579 304. Mesh Steam Separators 580 305. Location of Separators 580 306. Exhaust-Steam Separators and Oil Eliminators 581 307. Exhaust Heads 585 xii CONTENTS CHAPTER XIV — Continued Page 308. Drips 586 309. Low-Pressure Drips 586 310. Size of Pipes for Low-Pressure Drips 588 311. High-Pressure Drips 588 312. Classification of Traps 588 313. Float Traps 589 314. Bucket Traps 590 315. Dump or Bowl Traps 591 316. Expansion Traps 592 317. Differential Traps 594 318. Location of Traps 596 319. Drips under Vacuum 597 320. Drips under Alternate Pressure and Vacuum 599 321. The Steam Loop 600 322. The Holly Loop 602 323. Returns Tank and Pump 602 324. Office Building Drains 603 CHAPTER XV. — Piping and Pipe Fittings 606-668 325. General 606 326. Drawings 606 327. Material for Pipes and Fittings 606 328. Size and Strength of Commercial Pipe 608 329. Screwed Fittings 610 330. Flanged Fittings 610 331. Pipe Coverings 616 332. Expansion due to Temperature Variation 618 333. Pipe Supports and Anchors 621 334. General Arrangement of High-Pressure Steam Piping 622 335. Main Steam Headers 629 336. Flow of Steam in Pipes 632 337. Equation of Pipes 636 338. Friction through Valves and Fittings ^ 639 339. Exhaust Piping, Condensing Plants 642 340. Exhaust Piping, Non-Condensing Plant, Webster Vacuum Heating System 642 341. Exhaust Piping, Non-Condensing Plant, Paul Vacuum Heating System 643 342. Automatic Temperature Control 646 343. Feed-Water Piping 647 344. Flow of Water through Orifices, Nozzles and Pipes 650 345. Stop Valves 655 346. Automatic Non-Return Valves 658 347. Emergency Valves 658 348. Check Valves 660 349. Blow-off Cocks and Valves 661 350. Safety Valves 663 351. Back-Pressure and Atmospheric Relief Valves 665 352. Reducing Valves; Pressure Regulators 666 353. Foot Valves 668 CONTENTS xiii Page CHAPTER XVI. — Lubricants and Lubrication 669-689 354. General 669 355. Vegetable Oils 669 356. Animal Oils and Fats 669 357. Mineral Oils ■ 670 358. Solid Lubricants 671 359. Greases 671 360. Qualifications of Good Lubricants 671 361. Identification of Oils 672 362. Gravity 672 363. Viscosity 673 364. Flash Point 673 365. Burning Point or Fire Test 674 366. Acidity 674 367. Cold Test 674 368. Friction Test 674 369. Atmospheric Surface Lubrication • 675 370. Intermittent Feed 675 371. Restricted Feed 675 372. Oil Bath 675 373. Oil Cups 677 374. Telescopic Oiler 677 375. Ring Oiler 678 376. Centrifugal Oiler 678 377. Pendulum Oiler 679 378. "Splash " Oiling 679 379. Gravity Oil Feed 680 380. Low-Pressure Gravity System 680 381. Compressed-Air Feed . 680 382. Cylinder Lubrication 682 383. Cylinder Cups 682 384. Hydrostatic Lubricator 683 385. Forced-Feed Cylinder Lubricator 684 386. Siegrist System of Lubrication 685 387. Oil Filters 687 CHAPTER XVII. — Finance and Economics — Cost of Power 690-729 388. Records 690 389. Output 690 390. Load Factor 691 391. Cost of Operation 693 392. Fixed Charges 693 393. Interest 693 394. Depreciation 694 395. Maintenance 699 396. Taxes and Insurance 699 397. Operating Costs 699 398. Labor, Attendance, Wages 699 399. Fuel 700 400. Oil, Waste, and Supplies 703 401. Repairs and Maintenance 703 402. Cost of Power 703 xiv CONTENTS Page CHAPTER XVIII. — Testing and Measuring Instruments 730-749 403. General ; 730 404. Weighing Fuel 730 405. Measurement of Water 730 406. Steam Meters 734 406a. Pressure Gauges 735 407. Temperature Measurements 736 408. Power Measurements 741 409. Flue-Gas Analysis 741 410. Measurement of Moisture in Steam 745 411. Fuel Calorimeters 747 411a. Hamler-Eddy Smoke Recorder 749 CHAPTER XIX. — Typical Specifications 750-773 412. Specifications for a Cross Compound Non-Condensing Engine 750 413. Specifications for a Return Tubular Boiler 754 414. Specifications for a Condenser Plant 758 415. Specifications for a Piping System 760 416. Government Specifications for Purchasing Coal 769 CHAPTER XX. — A Typical Steam Turbine Station — Commonwealth Edison Company, Chicago 774-787 CHAPTER XXI. — A Typical Isolated Station — West Albany Power Station of the New York Central Railroad Company, West Albany, N. Y 788-797 APPENDIX A. — General Bibliography — Power Plant Engineering and Design 798-821 APPENDIX B. — A. S. M. E. Rules for Conducting Boiler Trials, Code of 1899 822-845 APPENDIX C. — A. S. M. E. Rules for Conducting Steam Engine Tests 846-872 APPENDIX D. — Steam Tables 873-876 APPENDIX E. — Equivalent Values of Mechanical and Electrical Units 877 APPENDIX F. — Miscellaneous Conversion Tables 878 APPENDIX G. — Rules for Firemen Using Illinois and Indiana Coal in Hand-Fired Furnaces 879-880 APPENDIX H. — Mollier Diagram 881-885 LIST OF TABLES Page 1. Composition of Typical American Coals 23 2. Data Relative to Elements Most Commonly Met with in Connection with Combustion 26 3. Weight of Air per Pound of Combustible as Indicated by the Percentage of C0 2 in the Flue Gas 31 4. Heat Carried away by the Dry Chimney Gases per Pound of Combustible.. 34 5. Loss Due to Incomplete Combustion of Carbon to Carbon Monoxide 35 6. Effect of Washing on Bituminous Coals . 41 7. Comparative Tests of Babcock & Wilcox Boiler. Lump Coal vs. Powdered Coal 46 8. Analyses of Typical American Fuel Oils 52 9. Boiler Efficiencies, Fuel Oil 54 10. Tests of Fuel-Oil Burners 62 11. Characteristics of Gaseous Fuels 67 12. Required Hourly Evaporation per Boiler Horse Power at Various Feed Temperatures and Steam Pressures 96 13. Ratio of Heating Surface to Grate Surface in Recent Boiler Installations .... 98 14. Examples of Steam Boiler Tests 100 15. Cost of Evaporating Water, Results of Actual Tests 105 16. Air Spaces and Thickness of Grate Bars 114 17. Values of c p at Atmospheric Pressure by Various Authorities 157 18. Average Yearly Expense for Repairs for Cast-iron Superheaters 171 19. Difference in Heat Efficiency of Superheaters Installed in Flue and Sepa- rately Fired Superheaters 176 20. Decrease in Temperature of Gases of Combustion due to Superheater Installed in Flue 177 21. Increase in Heat Efficiency of the Boiler due to Superheater 178 22. Comparative Boiler Tests, Saturated vs. Superheated Steam, at Spring Creek Pumping Station of the Brooklyn Waterworks 179 23. Density and Weight of Air and Chimney Gas at Various Temperatures. . . . 209 24. Theoretical Draft Pressures in Inches of Water for Various Chimney Tem- peratures 210 25. Test of a 100-Foot Steel Chimney ..213 26. Chimney Formulas » 215 27. Size of Chimneys for Steam Boilers 216 28. Approximate Weight and Cost of Guyed Steel Stacks 219 29. Steel Stack Dimensions 222 30. Dimensions of Steel Chimney Foundations 242 31. Dimensions of Brick Factory Chimneys 244 32-33. Test of Steam Jet Blowers 248 34. Sizes of Forced-Draft Fans 261 35. Sizes of Induced-Draft Fans 262 xv xvi LIST OF TABLES Page 36. Steam Engine Efficiencies 274 37. Mechanical Efficiencies of Engines 276 38. Distribution of Friction Losses in Engines 285 39. Performance of High-Speed Engines 296 40. Performance of Saturated- Steam Engines, Compound 306 41. Effect of Condensing on Engine Economy 310 42. Per Cent Moisture Evaporated by Throttling 312 43. Performance of Superheated-Steam Engines 315 44. Effect of Superheat on Simple Engines 316 45. Effect of Superheat on Compound Engines 317 46. Effect of Superheat on Triple-Expansion Engines 318 47. Record Steam Engine Performance. Superheated Steam 321 48. Performance of Steam Turbines 390 49. Pressures of Aqueous Vapor, Regnault 399 50. Ratio by Weight of Cooling Water to Steam Condensed 406 51. Size of Siphon Condensers . . 409 52. Square Feet Cooling Surface Necessary to Condense One Pound of Steam under Different Conditions 428 53. Test of a Pennel Saturated- Air Condenser 432 54. Test of a Cast-iron Evaporative Surface Condenser 434 55. Power Consumption of Condenser Auxiliaries 449 56. Most Economical Vacuum for Steam Turbines 452 57. Most Economical Vacuum for Piston Engines 453 58. Properties of Saturated Air 468 59. Influence of Thickness of Scale on Heat Transmission 472 60. Water and Boiler Scale Analyses 473 61. Boiler Defects, Report of Hartford Steam Boiler Inspection and Insurance Company 474 62. Percentage of Saving for each Degree of Increase in Temperature of Feed Water 485 63. Feed-Water Temperatures, Open Heaters 490 64. Extent of Heating Surface, Closed Heaters 498 65. Mean Temperature Difference, Closed Heaters 499 66. Heat Transmission, Closed Heaters 502 67. Economizer Tests 515 68. Pump Duties for Various Efficiencies and Steam Consumptions 537 69. Maximum Height to which Pumps can Raise Water by Suction — Tem- perature Constant 538 70. Maximum Height to which Pumps can Raise Water — Temperature Variable 539 71. Range of Working Pressure — Metropolitan Injectors 550 72. Commercial Sizes of Air Pumps for Condensers 558 73. Data Pertaining to Single-Stage Centrifugal Pumps 566 74. Data Pertaining to Multi-Stage Centrifugal Pumps 567 75. Tests of Steam Separators 576 76. Dimensions of Standard Wrought-Iron Pipes 611 77. Comparative Costs of Different Types of Flanges 614 78. Dimensions of Standard Flanges 614 ! 79. Dimensions of Extra Heavy Flanges 615 80. Loss of Heat from Bare Pipes in Still Air 615 81. Experiments on Pipe Coverings , 617 LIST OF TABLES xvii Page 82. Coefficients of Expansion — Piping Materials 620 83. Comparison of Formulas for the Flow of Steam in Pipes 634 84. Comparison of Formulas for the Flow of Steam in Pipes 635 85. Flow of Steam in Pipes, Babcock 637 86. Flow of Steam in Pipes, Sickles 638 87. Equation of Pipes 640 88. Specific Gravity of Lubricating Oils 673 89. Properties of Lubricating Oils 670, 676 90. Approximate Useful Life of Various Portions of Steam Power Plant Equipments 694 91. Rates of Depreciation 695 92. Depreciation Percentages — Chicago Traction Valuation Commission 696 93. Cost of Labor for Street Railway Plants 701 94. Cost of Labor for Tall Office Buildings 702 95. Operating Costs per Kilowatt Hour, Typical British Electric Light and Power Plants , 709 96. Operating Costs per Kilowatt Hour, Typical United States Railway Plants 709 97. Operating Costs per Kilowatt Hour, Average of all Stations, Boston Ele- vated • . 710 98. Operating Costs (1907), First National Bank Building, Chicago 710 99. Cost of One Horse Power per Year, Simple Engine, W. O. Webber 711 100. Cost of One Horse Power per Year, Compound Engine, W. O. Webber. . . 711 101. Cost of One Horse Power per Year, H. von Schon 712 102. Cost of Electrical Power per Year, W. M. Wilson 713 103. Cost of Electrical Power per Year, R. C. Carpenter 715 104. Cost of Electrical Power per Year, Oil Fuel, C. C. Moore & Co 716 105-7. Cost of Power, Typical Isolated Stations 720, 722 108. Cost of Power, First National Bank Building, Chicago 724 109. Temperature Ranges of Thermometers in General Use 739 Appendix D . — Steam Tables 873 Additional Tables — Third Edition. 00. Physical and Chemical Properties of Woods, Straw and Tanbark 19 0. Heat Values of Bagasse and Variation with Degree of Extraction 20 2a. Ratio of Total Air Supplied to that Theoretically Required for Various Analyses of Flue Gases 30 14a. Principal Data and Results, Boiler Unit No. 10, Fisk Street Station, Com- monwealth Edison Company 102 15a. Pounds of Water Evaporated per Hour from and at 212° F. per Pound of Fuel 108 17a. Mean Specific Heat of Superheated Steam 159 17b. Specific Volume of Superheated Steam 160 94a. Distribution of Maintenance and Operating Costs in Large Stations. .... 707 107a. Yearly Operating Costs in Four" Typical Central Stations 723 107b. Power Costs — Steam Electric Central Stations 723 ILLUSTRATIONS. Chapter I Elementary Steam Power Plants. Fig. 1. Elementary Non-Condensing Plant 2. Elementary Non-Condensing Plant with Heating System. 3. Simple Condensing Plant. 4. Condensing Plant with Full Complement of Heat-Saving Appliances. 4a. Sectional Elevation of the Myers Furnace for Burning Bagasse. 4b. End and Side Sectional Elevation of the Myers Furnace for Burning Tanbark, Chapter II. — Fuels and Combustion. 5. Relation of Gas Composition in Combustion Chamber to Temperature. 6. Influence of Size of Coal on Boiler Capacity and Efficiency. 7. Influence of Ash on Fuel Value of Dry Coal. 8. Pinther Coal-dust Feeder. 9. Schwartzkopff Coal-dust Feeder. 10. Aero-Pulverizer Coal-dust Feeder. 11. Triumph Coal-dust Feeder. 12. Korting Fuel Oil Burner. 13. Booth Fuel Oil Burner. 14. Hammel Fuel Oil Burner. 15. Branch Fuel Oil Burner. 16. Kirkwood Fuel Oil Burner. 17. Williams Fuel Oil Burner. 18. Warren Fuel Oil Burner. 19. Furnace for Burning Fuel Oil, Front Feed. 20. Furnace for Burning Fuel Oil, Rear Feed. 21. International Gas and Fuel Company's Fuel Oil System. 22. Hydraulic Oil Storage Company's Fuel Oil System. Chapter III. — Boilers. 23. Vertical Tubular Boiler with Submerged Tube Sheet. 24. Manning Vertical Fire-Tube Boiler. 25. Typical Fire-box Boiler, — Stationary Type. 26. Stationary Scotch-Marine Boiler. 27. Robb-Mumford Boiler. 28. Return Tubular Boiler Setting, — Extended Front. 29. Return Tubular Boiler Setting, — Flush Front. 30. Return Tubular Boiler Setting, — Steel Beam Suspension. 31. Boiler Setting, "Wood" Mill of the American Woolen Company, Lawrence, Mass. 32. Furnace Arch Bars. 33. Back Connection made with Cast-iron Plate. 34. Babcock and Wilcox Boiler and Setting. xx ILLUSTRATIONS Fig. 35. Details of Header, — Babcock and Wilcox Boiler. 36. Front Section, — Babcock and Wilcox Boiler. 37. Heine Boiler and Setting. 38. Wickes Vertical Water-Tube Boiler. 38a. 1200-H.P. Parker Boiler. 38b. Boiler Room Area for Various Types of Boilers. 39. Stirling Boiler and Setting. 39a. Heat Transmission Through Boiler Plate. 39b. Influence of Draft on Capacity, Torpedo Boat "Biddle." 40. Influence of Draft on the Efficiency and Capacity of a 350-Horse-power Babcock and Wilcox Boiler with Chain Grate. 41. Effect of Rate of Driving on Economy of a 150-Horse-power Stirling Boiler, Hand Fired. 41a. Relation Between Efficiency and Capacity, 500-Horse-power Babcock and Wilcox Boiler. 41b. Effect of Rate of Driving on Efficiency of a 600-Horse-power Babcock and Wilcox Boiler. 41c. Influence of Draft on the Capacity of a 600-Horse-power Babcock and Wilcox Boiler. 42. Effect of Thickness of Fire on the Capacity and Efficiency of a 350-Horse- power Stirling Boiler, equipped with Chain Grate. 43. Effect of Thickness of Fire on the Capacity and Efficiency of a 150-Horse- power Water-Tube Boiler. 44. Effect of Thickness of Fire on the Capacity and Efficiency of a 500-Horse- power Babcock and Wilcox Boiler. 45. Types of Grate Bars. 46. A Typical Rocking Grate. 47. Horizontal Blow-off Connection to Head. 48. Vertical Blow-off Connection to Shell. 49. Blow-off Connection with Circulating Pipe. 50. Blow-off Tank and Connections. 51. Surface Blow-off. 52. Buckeye Skimmer. 53. Kitts Hydraulic Damper Regulator. 54. Tilden Steam Actuated Damper Regulator. 55. Simple Water Column. 56. Water Gauge with Self-closing Valve. 57. Combined Water Column and High and Low Water Alarm. 58. Types of Fusible Plugs. 59. Mechanical Tube Cleaner, — Hammer Type. 60. Mechanical Tube Cleaner, — Turbine Type. Chapter IV. — Smoke Prevention, Furnaces, Stokers. 61. Green Chain Grate. 62. Babcock and Wilcox Boiler, Chain Grate, Ordinary Setting. 63. Babcock and Wilcox Boiler, Chain Grate, Fire-tile Roof. 64. Section of Fire Tile. 65. Section of Fire Tile. 66. Application of "Economy" Fire Files to Stirling Boiler. 67. Method of Anchoring "Economy" Fire Tiles to Tubes. 67a. Chain Grate Fired from Rear End of Setting. ILLUSTRATIONS xxi Fig. 67b. Smokeless Setting, Chain Grate and Babcock and Wilcox Boiler. 68. Details of Roney Stoker. 68a. Double Stoker Setting. 69. Details of Wilkinson Stoker. 70. Murphy Furnace, Front Section. 71. Murphy Furnace, Side Section. 72. Jones Underfeed Stoker. 73. American Underfeed Stoker. 74. Hawley Down-Draft Furnace. 75. Plain Dutch Oven. 76. "Twin Fire Arch," Applied to a Return-Tubular Boiler. 76 a, b, c. Chicago Setting, Hand-Fired Return-Tubular Boiler. 77. Wooley Smokeless Furnace. 78. Kent's Wing-Wall Furnace. 79. Burke's Smokeless Furnace, Front Section. 80. Burke's Smokeless Furnace, Side Section. 81. Split Bridge Wall. Chapter V. — Superheated Steam; Superheaters. 82. Specific Heat of Superheated Steam, Knoblauch and Linde. 83. Specific Heat of Superheated Steam, A. R. Dodge. 84. Specific Heat of Superheated Steam, C. E. Burgeon. 85. Specific Heat of Superheated Steam, Thomas. 86. Babcock and Wilcox Superheater. 87. Stirling Superheater. 88. Details of Stirling Superheater. 89. Foster Superheater in Babcock and Wilcox Boiler. 90. Schmidt Independently Fired Superheater. 91. Foster Independently Fired Superheater. 92. Schmidt System of Combined Superheater, Economizer and Feed-Water Heater. 92a. Relation between Gas Temperature, Heating Surface passed over and Amount of Steam Generated. 93. Relation of Superheat to Total Output of Boiler. 94. Relation of Output of Superheater to Boiler Output. 95. Relation of Superheat to Output of Superheater. Chapter VI. — Coal and Ash Handling Apparatus. 96. Link-Belt Coal-Handling Apparatus. 97. Typical Coal and Ash Handling Equipment. 98. Steel Cable Company's Coal-Handling Apparatus. 99. Coal and Ash Handling System of S. S. Elevated Railway. 100. Crusher and Conveyor of S. S. Elevated Railway. 101. Driving Mechanism of Hunt Conveyor. 102. Hunt Coal Conveyor System at Baltimore, Md. 103. Bucket and Screw Conveyor at Commercial National Bank Building, Chicago, Illinois. 104. Guide Pulleys, Robins Belt Conveyor. 105. Coal and Ash Handling System of Aurora and Elgin Railway. 106. Coa] and Ash Handling System of Cincinnati Traction Company. xxii ILLUSTRATIONS Fig. 107. Coal and Ash Handling System of Detroit Edison Company. 108. Vacuum Ash Handling System. 108a. Vacuum Ash Handling System at the Armour Glue Works. 108b. Coal and Ash Handling System, Norfolk Traction Company. 109. Stationary Coal-Weighing Hoppers. 110. Traveling Coal Hoppers. 111. Common Slide Coal Valve. 112. Simplex Coal Valve. 113. Duplex Coal Valve. 114. Flap Coal Valve. 115. Seaton Coal Valve. Chapter VII. — Chimneys. 116. Relation between Draft and Rates of Combustion. 117. Steel Chimney of S. S. Elevated Railway Power House, Chicago. 118. Stability of Steel Chimneys. 119. Custodis Radial Brick Chimney. 119a. Custodis Radial Perforated Brick. 120. Circular Brick Chimney at Armour Institute of Technology. 121. Design of Brick Chimneys, Thickness of Walls. 122. Design of Brick Chimneys, Stability. 123. Weber Steel-Concrete Chimney. 124. Weber Steel-Concrete Chimney. Chapter VIII. — Mechanical Draft. 125. Ring Steam Jet. 126. Bloomsburg Jet. 127. McClaves Argand Blower. 128. Hollow Bridge Wall and Steam Jet. 129. Parsons Smokeless Furnace. 130. Heinrich Smokeless Furnace. (Sectional Elevation.) 131. Heinrich Smokeless Furnace. (Sectional Plan.) 132. Typical Forced-Draft System. 133. Typical Induced-Draft System. 134. Pitot Tubes; Orifice Closed. 135. Pitot Tubes; Orifice Wide Open. 136. Pitot Tubes; Orifice Partly Closed. 137. Performance of Steel Plate Fans. 138. Performance of Pressure Blower, Speed Constant. 139. Performance of Pressure Blower, Speed Variable. 140. Comparative Costs of Chimneys and Mechanical Draft. 141. Influence of Rate of Combustion on Air Supply; Forced Draft. 142. Balanced Draft System. Chapter IX. — Steam Engines. 142a. Rankine Cycle. 142b. Side Elevation, Typical Corliss Engine. 142c. Plan View, Typical Corliss Engine. 142d. A Modern Piston Engine Plant. 143. Mechanical Efficiencies of Engine and Generator. 144. Status of the Steam Engine. 145. Condensation and Leakage Losses in Simple Engines. ILLUSTRATIONS xxiii Fig. 145a. Influence of Increasing Back Pressure. 146. Typical Curves of Steam Engine Friction. 147. Influence of Increasing Initial Pressure. 148. Typical Economy Curves — High-Speed Engines. 148a. Assembly of Valve Gear, Typical Corliss Engine. 148b. Section through Cylinder, Typical Corliss Engine. 148c. Assembly of Governor and Link Mechanism, Corliss Engine. 149. Test of Reeves Simple Engine; Condensing vs. Non-Condensing. 150. Typical Economy Curves of Single- Valve vs. Four- Valve High-Speed Engines. 150a. 3500-K.W. Vertical Cross Compound Corliss Engine. 150b. 7500-K.W. Vertical Horizontal Cross Compound Corliss Engine. 151. Effect of Compounding on High-Speed Non-Condensing Engines. 152. Performance of Corliss Compound; Condensing vs. Non-Condensing. 153. Performance of a 5500-Horse-power Engine. 154. Increase in Power Due to Vacuum. 155. Increase in Power Due to Vacuum. 156. Performance of a 5500-Horse-power Engine. 157. Indicator Cards — High-Speed Throttling Engines. 158. Indicator Cards — High-Speed Automatic Engines. 159. Effect of Superheat on Steam Consumption. 159a. 3000-Horse-power Sulzer Engine for Highly Superheated Steam. 159b. Fleming-Harrisburg Four- Valve Tandem Compound. 160. Effect of Superheat on Steam Compounds. 161. Influence of Superheat on Economy. 162. Diagrammatic Arrangement, Binary- Vapor Engine. 163. Cost of Simple High-Speed Engines. 164. Cost of High-Speed Compound Engines. 165. Cost of Simple and Low-Speed Compound Engines. Chapter X. — Steam Turbine. 166. Horizontal Section of De Laval Turbine. 167. Details of Blades of De Laval Turbine. 168. Details of Nozzle of De Laval Turbine. 169. Details of Governor of De Laval Turbine. 170. Theoretically Proportional Expanding Nozzle. 171. Theoretical Performance of a Divergent Nozzle. 172. Characteristic Performance of a Divergent Nozzle. 172a. Velocity Diagram, Ideal Impulse Turbine. 172b.. Velocity Diagram, as Modified by Friction Losses. 173. Section through Terry Turbine. 174. Arrangement of Buckets and Reversing Chambers, Terry Turbine. 175. Longitudinal Section through Kerr Turbine. 176. Sectional End Elevation, Kerr Turbine. 177. Details of Governor, Kerr Turbine. 178. Four-Stage Vertical Curtis Turbo-Generator. 179. 3500-K.W. Horizontal Curtis Turbine. 180. Arrangement of Nozzles and Blades, Curtis Turbine. 181. Section through Curtis Governor. 182. Mechanical Valve Gear, Curtis Turbine. 183. Hydraulic Valve Gear, Curtis Turbine. 183a. Steam Belt Area in Five-Stage Curtis Turbine. xxiv ILLUSTRATIONS Fig. 184. Velocity Diagram, Curtis Turbine. 185. Section through Hamilton-Holzworth Turbine. 186. Details of Vanes, Hamilton-Holzworth Turbine. 187. Details of Bearings, Hamilton-Holzworth Turbine 188. Details of Governor, Hamilton-Holzworth Turbine. 189. Section through Westinghouse-Parsons Standard Turbine. 190. Flow of Steam in Parsons Turbine. 191. Details of Governor, Westinghouse-Parsons Turbine. 192. Indicator Cards, Westinghouse-Parsons Turbine. 193. By-Pass Valve, Westinghouse-Parsons Turbine. 193a. Method of Fastening Blades, Westinghouse-Parsons Turbine. 193b. High-Pressure Double-Flow, Westinghouse-Parsons Turbine. 193c. Allis-Chalmers Steam Turbine. 194. Velocity Diagram, Multi-Stage Reaction Turbine. 194a. Low-Pressure Turbine, 59th St. Station, Interborough Rapid Transit Company. 194b. Low-Pressure Double-Flow Westinghouse-Parsons Turbine. 194c. Performance of 7500-K.W. Engine at 59th St. Station. 194d. Comparison of Economy Curves, Combined Engine and Turbine. 195. Rateau Low-Pressure Steam Turbine at South Chicago. 196. Rateau Regenerator Accumulator. 196a. Typical Double-Deck Turbine Installation. 197. Curve of Performance of Rateau Low-Pressure Turbine. 198. Comparative Floor Space, Engines vs. Turbines. 198a. Typical Performance of 90-Horse-power Terry Turbine. 198b. Typical Performance 9000-K.W. Curtis Turbine. 198c. Typical Performance Small Non-Condensing Turbines. 198d. Typical Correction Curves, 125-K. W. Turbines. 199. Reciprocating Engine vs. Turbine Economy. 200. Effect of Superheat on Economy. 201. Effect of Vacuum on Economy, Westinghouse-Parsons Turbine. 201a. Effect of a Vacuum on Economy. 202. Effect of Vacuum and Superheat on Economy. Chapter XL — Condensers. 203. Worthington Jet Condenser. 204. Blake Jet Condenser. 205. Baragwanath Siphon Condenser. 206. Schutte Ejector Condenser. 207. Piping for Schutte Condenser. 208. Weiss Counter-Current Condenser. 209. Alberger Barometric Condenser. 210. Worthington Barometric Condenser. 211. Tomlinson Type B Barometric Condenser. 211a. Centrifugal Pump Applied to Tail Pipe of a Barometric Condenser. 212. Baragwanath Surface Condenser. 213. Wheeler Surface Condenser and Pumps. 214. Wheeler Multi-flow Surface Condenser. 215. Weighton Multi-flow Surface Condenser. 216. Relation between Hot-well Temperature and Vacuum in Surface Condensers. 216a. Application of Weighton Dry-Tube Surface Condenser to Vertical Marine Ensrine. ILLUSTRATIONS xxv Fig. 216b. Heat Transfer in Condenser Tubes, Steam to Water. 216c, d. Heat Transfer in Condenser Tubes, Steam to Air. 217. Pennel Saturated-Air Surface Condenser. 218. Pennel Flask Type Atmospheric Condenser. 219. Jet Condenser located below Engine-Room Floor. 220. Surface Condenser located below Engine-Room Floor. 221. Surface Condenser Connected with Pumping Engine. 222. Jet Condenser located above Engine-Room Floor. 223. Typical Arrangement, Westinghouse-Leblanc Condenser and Curtis Turbine, 224. Elevation of Condenser Piping, Des Moines City Railroad Power House. 225. Plan of the Condenser Piping, Des Moines City Railroad Power House. 226. Plan of Condenser Piping, Northwestern Elevated Railroad Power House, Chicago. 226a. Condenser Installation, Quincy Point Power Plant. 227. Worthington High-Vacuum System. 228. Wheeler High- Vacuum System. 229. Parsons Vacuum Augmenter. 229a. Westinghouse-Leblanc High Vacuum Multi-Jet Condenser. 229b. Tomlinson Type C High Vacuum Jet Condener. 229c. Korting Multi-Jet Condenser. 230. Power Consumption of Auxiliaries, Parsons Turbine. 231. Power Consumption of Auxiliaries, Curtis Turbine. 232. Relative Cost of High- Vacuum Condensing Systems. 233. Performance of Spray Fountain. 234. Barnard-Wheeler Cooling Tower. 235. Worthington Cooling Tower. 236. Alberger Cooling Tower Installation. Chapter XII. — Feed-Water Heaters. 237. Scaife System for Feed-Water Purification. 238. We-Fu-Go System for Feed-Water Purification. 238a. Anderson System for Feed-Water Purification. 239. Cochrane Feed-Water Heater and Receiver. 240. Webster Star Vacuum Heater. 241. Hoppes Horizontal Heater and Purifier. 242. Goubert Single-flow Heater. 243. Expansion Joint, Goubert Heater. 244. Wainwright Multi-flow Closed Heater. 245. Coil Heater. 246. Otis Steam-Tube Feed-Water Heater. 247. Baragwanath Steam-Jacketed Heater. 248. Heat Transmission in Feed-Water Heater Tubes. 249. Open Heater Connected as a Through Heater 250. Through Heater with By-Pass. 251. Open Induced Heater, Non-Condensing Plant. 252. Closed Induced Heater, Condensing Plant. 253. Hoppes Live-Steam Heater and Purifier. 254. Installation of a Live-Steam Purifier. 255. Typical Installation of Primary and Secondary Heater. xxvi ILLUSTRATIONS Fig. 256. Green Economizer. 257. Economizer Installation at Weehawken, New Jersey. 258. Heat Transmission, Economizers. Chapter XIII. — Pumps. 259. Duplex Direct-Acting Boiler-Feed Pump. 260. Section through Duplex Boiler-Feed Pump. 261. Method of Obtaining Lost Motion, — Duplex Valve Gear. 262. Method of Obtaining Lost Motion, — Duplex Valve Gear. 263. Position of Valve and Piston at the Beginning of Stroke. 264. Position of Valve and Piston at the End of Stroke. 265. Pump Disk Valve. 266. Section through a Compound Duplex Pump. 267. Section through a Simplex Pump with Steam-Actuated Gear. 268. Forms of Vacuum Chambers. 269. Different Arrangements of Vacuum Chambers. 270. Types of Water Pistons. 271. Plunger with Metal Packing Ring. 272. Plunger with Hydraulic Packing. 273. Horizontal Fly-Wheel Pump with Outsicre Packed Plunger. 274. Performance of Direct-Acting Pressure Pumps. 275. Performance of Boiler-Feed Pump at the Armour Institute of Technology. 276. Fisher Pump Governor. 277. Kitts Feed- Water Regulator. 278. Rowe Feed-Water Regulator. 279. Triplex Pump. 279a. Performance of Triplex Pump, Direct Connected. 280. Performance of Triplex Pump, Geared. 281. Elementary Form of Ejector. 282. The Hancock Inspirator. 283. The Penberthy Automatic Injector. 284a. Performance of a Desmond Automatic Injector with Varying Initial Pressure. 284b. Performance of a Desmond Automatic Injector with Varying Suction Temperature. 284c. Performance of a Desmond Automatic Injector with Varying Discharge Pressure. 285. Dean Jet Condenser Air Pump. 286. Edwards Air Pump. 287. Mullan Valveless Air Pump. 288. Hewes and Phillips Air Pump. 289. Alberger Dry- Air Pump. 290. Air Pump Indicator Diagram. 291. Types of Impellers, Centrifugal Pumps. 292. A Typical Centrifugal Pump. 293. Direction of Water from Impeller of Volute Pumps without Diffusion Vanes. 294. Effect of Diffusion Vanes on the Direction of Water. 295. Three-Stage Lea-Degan Turbine Pump. 296. Six-Stage Rateau Turbine Pump. 297. Test of Centrifugal Pump at Armour Institute. 298. Centrifugal Pump Characteristic for Boiler-Feed Pumps. ILLUSTRATIONS xxvii Fig. 299. Centrifugal Pump Characteristic for Dry-Dock Service. 300. Centrifugal Pump Characteristic for Waterworks Service. 301. Performance of a 6-inch Worthington Conoidal Centrifugal Pump. 302. Performance of a Single-Stage De Laval Volute Pump. 303. Performance of a Two-Stage Turbine Pump. 304. Performance of a Two-Stage De Laval Centrifugal Pump. 305. Two-Lobe Cycloidal Rotary Pump. 306. Rotary Pump with Movable Butment. 307. Performance of a Rotary Pump. 308. High-Duty Circulating Pump, New York Rapid Transit Company. 309. Pulsometer. 310. Air Lift. Chapter XIV. — Separators, Traps, Drains. 311. Hoppes Live-Steam Separator. 312. Stratton Live-Steam Separator. 313. Keystone Live-Steam Separator. 314. Bundy Live-Steam Separator. 315. Austin Live-Steam Separator. 316. Direct Live-Steam Separator. 317. Baum Oil Separator. 318. Loew Grease Extractor. 319. Typical Exhaust Head. 320. Closed Heater as a Blow-off Tank. 321. Piping Drips to Exhaust Pipe. 322. McDaniel Float Trap. 323. Acme Bucket Trap. 324. Bundy Tilting Trap. 325. Columbia Expansion Trap. 326. Geipel Expansion Trap. 327. Dunham Expansion Trap. 328. Heintz Expansion Trap. 329. Flinn Differential Trap. 330. Simple Siphon Trap. 331. Location of Return Trap. 332-3. Drainage for Jackets and Receivers of Triple Expansion Pumping Engines. 334. Gravity Drainage; Vacuum Heater. 335. Method of Draining Heater under Vacuum. 336. Method of Draining Receivers under Alternate Pressure and Vacuum 337. Steam Loop. 338. Holly Loop. 339. Section through Holly Receiver. 340. Returns Tank and Pump. 341. Shone Ejector. Chapter XV. — Piping and Pipe Fittings. 342. United States Standard Pipe Thread. 343. Types of Pipe Flanges. 344. Efficiencies of Various Pipe Coverings. 345. Pipe Bends. 346. Double-Swing Screwed Fittings for Expansion. xxviii ILLUSTRATIONS Fig. 347. Slip Expansion Joint. 348. Typical Pipe Hanger. 349. Typical Wall Bracket with Roll Binder. 349a. A Typical Floor Stand. 350. Typical Pipe Anchor. 351. Arrangement of Steam Piping, Princeton University Power Plant. 352. Typical "Duplicate" Header System. 353. Typical "Loop Header" System. 354. Typical "By-Pass" Piping System. 355. General Arrangement of Steam and Exhaust Piping, Heyworth Building, Chicago, Illinois. 356. General Arrangement of Piping, Manhattan Elevated Station, New York. 357-8. Piping Arrangement at the Yonkers Power House of the New York Central. 359. Overhead Piping of Boilers, Quincy Point Power Plant of the Old Colony Street Railway Company, Quincy Point, Mass. 360. Main Stream Header and Branches, Grand Rapids, Grand Haven and Muskegon Railway Power House. 361. Main Header and Branches, Des Moines City Railway Power House. 362. Drop in Pressure in Steam Pipes of Various Diameters at Different Velocities. 363. Diagrammatic Arrangement of Piping in the Webster Vacuum Heating System. 364. Webster Vacuum Seal Valve. 365. Automatic Vacuum Valve, Illinois Engineering Company. 366. Diagrammatic Arrangement of Piping in the Paul Vacuum Heating System. 367. Paul Exhauster. 368. Paul Vacuum Valve. 369. Powers Thermostat. 370. Typical Diaphragm Valve. 371. Diagram of Feed-Water Piping, Condensing Plant. 372. Diagram of Feed- Water Piping, Non-Condensing Plant. 373. Arrangement of Valves in Feed-Water Branches. 374. Globe Valve, Screw-Top, Inside Screw. 375. Globe Valve, Bolt-Top, Outside Screw. 376. Gate Valve, Solid-Wedge, Screw-Top, Outside Screw. 377. Gate Valve, Solid-Wedge, Bolt-Top, Inside Screw. 378. Gate Valve, Split-Wedge, Bolt-Top, Inside Screw. 379. Ludlow Angle Valve, Gate Pattern. 380. Anderson Automatic Non-Return Valve. 381. Crane Hydraulic Emergency Gate Valve. 382. Anderson Triple-Duty Emergency Valve. 383. Pilot Valve, Anderson Triple-Duty Emergency Valve. 384. Types of Check Valves. 385-7. Types of Blow-off Valves. 388. Blow-off and Feed-Water Piping, South Side Elevated Railway Station, Chicago, Illinois. 389. Dead-Weight Safety Valve. 390. Common Lever Safety Valve. 391. Consolidated Pop Safety Valve. 392. Foster Back-Pressure Valve. 393. Davis Back-Pressure Valve. 394. Crane Atmospheric Relief Valve. 395. Acton Atmospheric Relief Valve. ILLUSTRATIONS xxix Fig. 396. Kieley Reducing Valve. 397. Foster Pressure Regulator. 398. Types of Foot Valves. Chapter XVI. — Lubricants and Lubrication. 399. Oil Cup Lubrication. 400. Nugent's Telescopic Oiler. 401. Ring Oiler. 402. Centrifugal Oiler. 403. Pendulum Oiler. 404. Simple Gravity Feed System. 405. Low-Pressure Gravity Oil Feed. 40G. Compressed-Air Oiling System at First National Bank Building, Chicago, Illinois. 407. Leyland Automatic Oil Cup. 408. Common Sight-Feed Hydrostatic Lubricator. 409. Lunkenheimer Sight-Feed Lubricator. 410. Central Hydrostatic Lubricating System. 411. Rochester Forced-Feed Lubricator, Single Feed. 412. Forced-Feed Cylinder Lubricator, Multi-feed. 413. Siegrist System. 414. Siegrist Sight-Feed Lubricator. 415. White Star Oil Filter. 416. Turner Oil Filter. Chapter XVII. — Finance and Economics — Cost of Power. 417. Influence of Load Factor on Cost of Power. 418. Cost of Power, Manufacturing Plant. 418 a, b, c. Cost of Power in Large Central Stations. Chapter XVIII. — Testing and Measuring Instruments. 419. Piston Water Meter. 420. Disk Water Meter. 421. Venturi Meter. 422. St. John's Steam Meter. 422a. Burnham Steam Meter. 423. Different Forms of Manometer Draft Gauges. 424. Bourdon Pressure Gauge. 425. Bristol Recording Air Thermometer. 426. Bristol Thermo-Electric Pyrometer. 427. Element for Callendar Resistance Pyrometer. 428. Wanner Optical Pyrometer. 429. Fery Radiation Pyrometer. 430. Orsat Apparatus. 431. Arndt's Econometer. 432. Ados C0 2 Recorder — Gas-Weighing Apparatus. 433. Sarco C0 2 Recorder. 434. Separating Calorimeter. xxx ILLUSTRATIONS Fig. 435. Throttling Calorimeter. 435a. Universal Calorimeter. 436. Mahler Bomb Calorimeter. 437. Parr Fuel Calorimeter. Chapter XX. — A Typical Steam Turbine Station. 438. General Arrangement of Plant and Grounds. 439. North Elevation of Building. 440. General Plan of Boiler and Turbine Room. 441. Section through Boiler and Turbine Room. 442. Section through Boiler Room. 443. General Plan Quarry Street Station. 444. Side Elevation of Quarry Street Station. 445. Sectional Elevation of Unit No. 4 Quarry Street Station. Chapter XXI. — A Typical Isolated Station. 446. Plan of Ground Floor. 447. Sectional Elevation through Line DD of Fig. 446. 448. Cross Section through Line CC of Fig. 446. 449. Longitudinal Section through BB of Fig. 446. 450. Plan of Basement. 451. Diagram of Switchboard Connections. Appendix B. — A. S. M. E. Rules for Conducting Boiler Tests. 452. Ringlemann Smoke Chart. 453. Graphical Log Boiler Test. Appendix C. — A. S. M. E. Rules for Conducting Steam Engine Tests. 454-5. Rope Brakes. 456. Alden Absorption Dynamometer. 457. Indicator Card, Simple Engine. 458. Indicator Card, Four- Valve Engine, Slow Speed. 459. Indicator Card, Single- Valve Engine, High Speed. 460-1. Temperature-Entropy Diagrams. Appendix H. — Mollier Diagram. 462. General Outline of Mollier's Diagram. 463. Complete Reproduction of Diagram. . STEAM POWER PLANT ENGINEERING CHAPTER I. ELEMENTARY STEAM POWER PLANTS. 1. General. — An equipment for the generation of power is known as a station or plant. When equipped to generate electricity for the production of light or power it is known as an Electrical Station or Electric Light and Power Station. The term Heating Plant refers to a plant in which the heat energy of fuel is made available for heating purposes through the medium of steam or hot water. In general, plants or stations are designated according to the manner in which the energy of the fuel is utilized. When a station distributes power to a number of consumers more or less distant, it is called a Central Station. When the distances are very^ great, electrical current of high tension is frequently employed, and is transformed and distributed at convenient points through Sub- stations. A plant designed to furnish power or heat to a building or a group of buildings under one management is called an Isolated Station. For example, the power plant of an office building is usually called an isolated station. When the exhaust steam from the engines is discharged at approxi- mately atmospheric pressure, the plant is said to be operating non- condensing. When the exhaust steam is condensed, reducing the back pressure on the piston by the partial vacuum thus formed, the plant is said to operate condensing. When the exhaust steam may be used for manufacturing, heating, or other useful purposes, as is frequently the case in various manufac- turing establishments, and in large office buildings, it is usually more economical to run non-condensing, while power plants for electric lighting and power, pumping stations, air compressor plants, and others, in which the load is fairly constant and the exhaust steam is not required for heating, are generally operated condensing. 2 STEAM POWER PLANT ENGINEERING 2. Elementary Non-Condensing Plant. — Fig. 1 gives a diagrammatic outline of the essential elements of the simplest form of steam power plant. The equipment is complete in every respect and embodies all the accessories necessary for successful operation. Where a small r+l INJECTOR Fig. 1. Elementary Non-Condensing Plant. amount of power is desired at intermittent periods, as in hoisting systems, threshing outfits and traction machinery, the arrangement is substantially as illustrated. The output in these cases seldom exceeds 50 horse power and the time of operation is usually short, so the cheapest of appliances are installed, simplicity and low first cost being more important than economy of fuel. Such a plant has three essential elements: (1) The furnace, (2) the boiler and (3) the engine. Fuel is fed into the furnace, where it is burned. A portion of the heat liberated from the fuel by combustion is absorbed by the water in the boiler, converting it into steam under pressure. The steam being admitted to the cylinder of the engine does work upon the piston, and is then exhausted through a suitable pipe to the atmosphere. The process is a continuous one, fuel and water being fed into the furnace and the boiler in proportion to the power demanded. ELEMENTARY STEAM POWER PLANTS 3 In such an elementary plant, certain accessories are necessary for successful operation. The grate for supporting the fuel during com- bustion consists of a cast iron grid or of a number of cast iron bars spaced in such a manner as to permit the passage of air through the fuel from below. The solid waste products fall through or are " sliced " through the grate bars into the ash pit, from which they may be removed through the ash door. The latter acts also as a means of regulating the supply of air below the grate. Fuel is fed into the furnace through the fire door, and when occasion demands, air may be supplied above the bed of fuel by means of this door. The combustion chamber is the space between the bed of fuel and the boiler heating surface, its office being to afford a space for the oxidation of the combustible gases from the solid fuel before they are cooled below ignition temperature by the com- paratively cool surfaces of the boiler. The chimney or stack discharges the products of combustion into the atmosphere, and serves to create the draft necessary to draw the air through the bed of fuel. Various forced draft appliances are sometimes used to assist or to entirely replace the chimney. The heating surface is that portion of the boiler area which comes into contact with the hot furnace gases, absorbs the heat and transmits it to the water. In the small plant, illustrated in Fig. 1, the major portion of the heating surface is composed of a number of fire tubes below the water line, through which the heated gases pass. The vol-ume above the water level is called the steam space. Water is forced into the boilers either by a feed pump or an injector. In small plants of the type considered, steam pumps are seldom em- ployed; the injector answers the purpose and is considerably cheaper. A safety valve connected to the steam space of the boiler automatically permits steam to escape to the atmosphere if an excessive pressure is reached. The water level is indicated by try cocks or by a gauge glass, the top of which is connected with the steam space and the bottom with the water space. Try cocks are small valves placed in the water column or boiler shell, one at normal water level, one above it, and one below. By opening the valves from time to time the water level is approximately ascertained. They are ordinarily used in case of acci- dent to the gauge glass. Fusible plugs are frequently inserted in the boiler shell at the lowest permissible water level. They are com- posed of an alloy having a low fusing point which melts when in con- tact with steam, thus giving warning by the blast of the escaping steam if the water level gets dangerously low. The blowoff cock is a valve fitted to the lowest part of the boiler to drain it of water or to discharge the sediment which deposits in the bottom. The steam outlet of a boiler is usually called the steam nozzle. 4 STEAM POWER PLANT ENGINEERING The essential accessories of the simple steam engine include: A throttle valve for controlling the supply of steam to the engine; the governor, which regulates the speed of the engine by governing the steam supply; the lubricator, attached to the steam pipe, which is usually of the "sight feed" class and provides for lubrication of piston and valve. Lubrication of the various bearings is effected by oil cups suitably located. Drips are placed wherever a water pocket is apt to form in order that the condensation may be drained. The apparatus to be driven by the engine may be direct connected to the crank shaft or belted to the fly wheel or geared. In small plants of this type no attempt is made to utilize the exhaust steam except in instances where the stack is too short to create the necessary draft, in which case the, exhaust may be discharged up the stack. If the draft is produced by convection of the heated gases in the chimney, the fuel is said to be burned under natural draft; if the natural draft is assisted by the exhaust steam, the fuel is said to be burned under forced draft. The power realized from a given weight of fuel is very low and seldom exceeds 2\ per cent of the heat value of the fuel. The distribution of the various losses in a plant of, say, 40 horse power is approximately as follows: B.T.U. Heat value of 1 pound of coal 14,500 Boiler and furnace losses, 50 per cent 7,250 Heat of the steam, 50 per cent 7,250 Heat equivalent of one horse power hour 2,545 Heat used to develop one horse power hour (50 pounds steam per horse power hour, pressure 80 pounds gauge, feed water 62 degrees F.) 57,500 Per cent. 2 545 Percentage of heat in the. steam, realized as work, ' .... 4.4 O i ,OUU 2 545 Percentage of heat value of the coal realized as work, __ _ ' — rr-pz. 2.2 • 57,500 -;- 0.50 The power plant of the modern locomotive is very much like that illustrated in Fig. 1, the main difference lying in the type of boiler and engine. The entire exhaust from the engine is discharged up the stack through a suitable nozzle, since the extreme rate of combustion requires an intense draft. The engine is a highly efficient one compared to that in the illustration, and the performance of the boiler is more economical. In average locomotive practice about 6 per cent of the heat value of the fuel is converted into mechanical energy at the draw bar. In general, a non-condensing steam plant in which the heat of the exhaust is wasted is very uneconomical of fuel, eA^en under the ELEMENTARY STEAM POWER PLANTS 5 most favorable conditions, and seldom transforms as much as 7 per cent of the heat value of the fuel into mechanical energy. 3. Non-Condensing Plant. Exhaust Steam Heating. — Fig. 2 gives a diagrammatic arrangement of a simple non-condensing plant differ- ing from Fig. 1 in that the exhaust steam is used for heating pur- poses. This shows the essential elements and accessories, but omits a number of small valves, by-passes, drains, and the like for the sake of simplicity. The plant is assumed to be of sufficient size to warrant the installation of efficient appliances. Steam is led from the boiler to the engine by the steam main. The moisture is removed from the steam before it enters the cylinder by a steam separator. The moisture drained from the separator is either discharged to waste or returned to the boiler. The exhaust steam from the engine is discharged into the exhaust main where it mingles with the steam exhausted from the steam pumps. Since the exhaust from engines and pumps contains a large portion of the cylinder oil introduced into the live steam for lubricat- ing purposes, it passes through an oil separator before entering the heating system. After leaving the oil separator the exhaust steam is diverted into two paths, part of it entering the feed water heater where it condenses and gives up heat to the feed water, and the remainder flowing to the heating system. During warm weather the engine generally exhausts more steam than is necessary for heating purposes, in which case the surplus steam is automatically discharged to the exhaust head through the back pressure valve. The back pressure valve is, virtually, a large weighted check valve which remains closed when the pressure in the heating system is below a certain prescribed amount but which opens automatically when the pressure is greater than this amount. During cold weather it often happens that the engine exhaust is insufficient to supply the heating system, the radiators condensing the steam more rapidly than it can be supplied. In this case live steam from the boiler is automatically fed into the main heating supply pipe through the reducing valve. The condensed steam, and the entrained air which is always present, are automatically discharged from the radiators by a thermostatic valve into the returns header. The thermostatic valve is so constructed that when in contact with the comparatively cool water of condensation it remains open and when in contact with steam it closes. The vacuum pump or vapor pump exhausts the condensed steam and air from the returns header and discharges them to the returns tank. The small pipe S admits cold water to the vacuum pump and serves to condense the heated vapor, and at the same time supply the necessary make up water to the system. The returns tank is open to the atmosphere so STEAM POWER PLANT ENGINEERING ELEMENTARY STEAM POWER PLANTS 7 that the air discharged from the vacuum pump may escape. From the returns tank the condensed steam gravitates to the feed water heater where its temperature is raised to practically that of the exhaust steam. The feed water gravitates to the feed pump and is forced into the boiler. There are several systems of exhaust steam heating in current practice which differ considerably in details, but, in a broad sense, are similar to the one just described. The more important of these will be described later on. During the summer months when the heating system is shut down, the plant operates as a simple non-condensing station and practically all of the exhaust steam, amounting to perhaps 60 per cent of the heat value of the fuel, is wasted. The total coal consumption, therefore, is charged against the power developed. During the winter months, however, all, or nearly all of the exhaust steam may be used for heating purposes and the power becomes a relatively small percentage of the total fuel energy utilized. The percentage of heat value of the fuel chargeable to power depends upon the size of the plant, the number and character of engines and boilers, and the conditions of operation. It ranges anywhere from 50 per cent to 100 per cent for the summer months and may run as low as 6 per cent for the winter months. This is on the assumption, of course, that the engine is debited only with the difference between the coal necessary to produce the heat entering the cylinder and that utilized in the heating system. 4. Elementary Condensing Plant. — Under the most favorable con- ditions a non-condensing plant can never be expected to realize more than 7 per cent of the heat value of the fuel as power. In large non- condensing power stations the demand for exhaust steam is usually limited to the heating of the feed water, and as only 12 per cent or 15 per cent can be utilized in this manner, the greater portion of the heat in the exhaust is lost. Non-condensing engines require from 20 to 60 pounds of steam per hour for each horse power developed. On the other hand in condensing engines the steam consumption may be reduced to as low as 10 pounds per horse power hour. The saving of fuel is at once apparent. Fig. 3 gives a diagrammatic arrangement of a simple condensing plant in which the back pressure on the engine is reduced by condens- ing the exhaust steam. A different type of boiler from that in Fig. 1 or Fig. 2 has been selected for the purpose of bringing out a few of the characteristic elements. The products of combustion instead of passing directly through fire tubes to the stack as in Fig. 1 are deflected back and forth across a number of water tubes, by the bridge wall and a series of baffles. After imparting the greater part of their heat to the STEAM POWER PLANT ENGINEERING fS* ..... t HH ^^ ■ numm'm^ ssssssss ^^^ ^^^ ELEMENTARY STEAM POWER PLANTS 9 heating surface the products of combustion escape to the chimney through the breeching or flue. The rate of flow is regulated by a damper placed in the breeching as indicated. The steam generated in the boiler is led to the engine through the main header. The steam is exhausted into a condenser in which its latent heat is absorbed by injection or cooling water. The process condenses the steam and creates a partial vacuum. The condensed steam, injection water, and the air which is invariably present are withdrawn by an air pump and discharged to the hot well. In case the vacuum should fail as by stoppage of the air pump the exhaust steam is automatically discharged to the exhaust head by the atmospheric relief valve, and the engine will operate non-condensing. The atmos- pheric relief valve is a large check valve which is held closed by atmos- pheric pressure as long as there is a vacuum in the condenser. When the vacuum fails the pressure of the exhaust becomes greater' than that of the atmosphere and the valve opens. The feed water may be taken from the hot well or from any other source of supply and forced into the heater. In this particular case it is taken from a cold supply and upon entering the heater is heated by the exhaust steam from the air and feed pumps. From the heater it gravitates to the feed pump and is forced into the boiler. Various other combinations of heaters, pumps, and condensers are necessary in many cases, depending upon the conditions of operation. Feed pumps, air pumps, and in fact all small engines used in connection with a steam power plant are usually called auxiliaries. A well-designed station similar to the one illustrated in Fig. 3 is capable of converting about 10 per cent of the heat value of the fuel into mechanical energy. The various heat losses are approximately as follows: BOILER LOSSES. Per Cent. Loss due to fuel falling through the grate 2 Loss due to incomplete combustion 2 Loss due to heat carried away in chimney gases 23 Radiation and other losses "8 Total V. ~35 Heat used by engines and auxiliaries (16 pounds of steam per ' ' ' I.H.P. hour, pressure 150 pounds, feed water 210° F.) . . . . 16,250 Engine and generator friction, 5 per cent 812 Leakage, radiation, etc., 2 per cent 325 Total 17,387 Heat equivalent of one electrical horse power 2,545 Percentage of the heat value of the steam converted into electrical Cent. energy 14 .7 Percentage of heat value of fuel converted into electrical energy 2545 X 0.65 17,387 ' •■ " y ' 5 10 STEAM POWER PLANT ENGINEERING 5. Condensing Plant with Full Complement of Heat-saving Appli- ances. — When fuel is costly it frequently becomes necessary for the sake of economy to reduce the heat wastes as much as possible. The chimney gases, which in average practice are discharged at a tem- perature between 450 degrees F. and 550 degrees F., represent a loss of 20 per cent to 30 per cent of the total value of the fuel. If part of the heat could be reclaimed without impairing the draft the gain would be directly proportional to the reduction in temperature of the gases. Again, in some types of condensers all of the steam exhausted by the engine is condensed by the circulating water and discharged to waste. If provision could be made for utilizing part of the exhaust steam for feed-water heating, the efficiency of the plant could be correspondingly increased. In many cases the cost of installing such heat-saving devices would more than offset the gain effected, but occasions arise where they give marked economy. Fig. 4 gives a diagrammatic arrangement of a condensing plant in which a number of heat-reclaiming devices are installed. The plant is assumed to consist of a number of engines, boilers, and auxiliaries. Coal is automatically transferred from the cars to coal hoppers placed above the boiler, by a system of buckets and conveyors. These hoppers store the coal in sufficient quantities to keep the boiler in continu- ous operation for some time. From the hoppers the coal is fed intermittently to the stoker by means of a down spout. The stoker feeds the furnace in proportion to the power demanded and auto- matically rejects the ash and refuse to the ash pit. The ashes are removed from the ash pit when occasion demands, and are transferred to the ash hopper by the same system of buckets and conveyor which handles the coal. The ash hopper is usually placed alongside the coal hoppers and is not unlike them in general appearance and construction. The products of combustion are discharged to the stack through the flue or breeching. Within the flue is placed a feed-water heater called an economizer, the function of which is to absorb part of the heat from the gases on their way to the chimney. The heat reclaimed by the economizer varies widely with the conditions of operation and ranges between 5 per cent and 20 per cent. Since the economizer acts as a resistance to the passage of the products of combustion it is sometimes necessary to increase the draft either by increasing the height of the chimney or, as is the usual practice, by using a forced draft system. Part of the heat of the exhaust steam is reclaimed by a vacuum heater which is placed in the exhaust line between engine and condenser. For example, if the feed water has a normal temperature of 60 degrees F. and the vacuum in the condenser is 26 inches, the vacuum heater will ELEMENTARY STEAM POWER PLANTS 11 12 STEAM POWER PLANT ENGINEERING raise the temperature of the feed to say 120 degrees F. ; thereby effecting a gain in heat of approximately 6 per cent. If the feed supply is taken from the hot well the vacuum heater is without purpose, as the temperature of the hot well will not be far from 120 degrees F. Referring to the diagram, the path of the steam is as follows: From the boiler it flows through the boiler lead to the main header or equalizing pipe. From the main header it flows through the engine lead to the high-pressure cylinder. The exhaust steam discharges from the low- pressure cylinder through the vacuum heater and into the condenser. Part of the exhaust steam is condensed in the vacuum heater and gives up its latent heat to the feed water. The remainder is condensed by the injection water which is forced into the condenser chamber by the circulating pump. The condensed steam and circulating water gravitate through the tail pipe to the hot well. The air which enters the con- denser either as leakage or entrainment is withdrawn by the air pump. The steam exhausted by the feed pump, air pump, stoker engine, and other steam-driven auxiliaries is usually discharged into the atmospheric heater, which still further heats the feed water. Referring to the feed water, the circuit is as follows: The pump draws in cold water at a temperature of say 60 degrees F., and forces it in turn through the vacuum heater, the atmospheric heater, and the economizer into the boiler. The vacuum heater raises the temperature to 120 degrees F., the atmospheric heater increases it to 212 degrees F., and the economizer still further to about 300 degrees F. The heat reclaimed by this series of heaters is evidently the equivalent of that necessary to raise the feed water from 60 degrees F. to 300 degrees F., or approximately 24 per cent of the total heat supplied. In some plants the economizer only is installed, in others the economizer and atmospheric heater are deemed desirable, still others utilize all three. The distribution of the heat losses in a plant of this type operating under favorable conditions is approximately as follows: Per Cent. B.T.U. Delivered to engine, 15 pounds steam per I.H.P. hour; pressure 150 pounds, feed 60° F 100 17,482 Delivered to feed pump 1.5 262 Delivered to circulating pump • 1.5 262 Delivered to air pump 2 349 Delivered to small auxiliaries 1-5 262 Loss in leakage and drips • • 0-5 87 Engine and generator friction 5 874 Radiation and minor losses 1 175 Total 19.753 ELEMENTARY STEAM POWER PLANTS 13 PerCent. B.T.U. Returned by vacuum heater 5.5 1,086 Returned by atmospheric heater 7.9 1,560 Returned by economizer 9.7 1,916 Total 23.1 4,562 Net heat delivered to engine in the form of steam to pro- duce one electrical horse power, 19,753—4,562 .... 15,191 2545 Percentage converted to electrical power - ■ .... 16.7 Boiler efficiency 70 Percentage of heat value of fuel necessary to produce one electrical horse power at switchboard rriaT — * * H • 7 15,191 The preceding figures give the results of very good practice. So much depends upon the size and character of the prime movers, the nature of the fuel, and the conditions of operation that no definite figure can be given for the percentage of heat converted to power in a given type of station. Six per cent represents good average practice in a non-condensing plant and 10 per cent in a condensing plant. Pumping stations operating continuously under full load have realized as much as 15 per cent of the total heat value of the fuel, but such performances are practically unobtainable in connection with steam- driven electrical power plants. Steam power plants as a class are very wasteful of fuel at the best. One of the best recorded performances to date (March, 1909) of a steam-electric power plant is that of the Pacific Light and Power Company at Redondo, Cal. When operating under regular commer- cial conditions approximately 14 per cent of the available heat of the fuel (crude oil) is realized as power at the switchboard. For a detailed description of the plant and the results of the acceptance tests, see Jour, of Elec. Gas and Power, Aug. 22, 1908. CHAPTER II. FUELS AND COMBUSTION. 6. General. — The subject of fuels and combustion has been so extensively treated by various authorities that a comprehensive dis- cussion would be without purpose here, but in order to bring out more clearly the matter pertaining to the commercial design and operation of steam power plants a few of the essential elements will be briefly treated. The fuels used for steam making are coal, coke, wood, peat, mineral oil, natural and artificial gases, refuse products such as straw, manure, sawdust, tan bark, bagasse, and occasionally corn and molasses. In most cases that fuel is selected which develops the required power at the lowest cost, taking into consideration all of the circumstances that may affect its use. Occasionally the disposition of waste products is a factor in the choice, but such instances are uncommon. The boilers and furnaces are designed to suit the fuel selected. 7. Classification of Fuels. — Fuels may be divided into three classes as follows: 1. Solid fuels. a. Natural fuels: straw, wood, peat, coal. b. Prepared: charcoal, coke, peat and other briquettes. 2. Liquid fuels. a. Natural: crude oils. b. Prepared: distilled oils, alcohol, molasses. 3. Gaseous fuels. a. Natural: natural gas. b. Prepared: coal gas, water gas, producer gas, oil gas. 8. Solid Fuels. — Solid fuels are of vegetable origin and exist in a variety of forms between that of a comparatively recent cellulose growth and that of nearly pure carbon as anthracite coal. They owe their forms to the conditions under which they were created or to the geological changes which they have undergone. With each succeeding stage the percentage of carbon increases. The chemical changes are approxi- mately as follows: 14 FUELS AND COMBUSTION 15 Substance. Pure cellulose . . . Wood Peat Lignite Brown coal .... Bituminous coal . . , Semi-bituminous coal. Anthracite Graphite Carbon. Hydrogen. Per Cent Per Cent 44.44 6.17 52.65 5.25 59.57 5.96 66.04 5.27 73.18 5.58 75.06 5.84 89.29 5.05 91.58 3.96 100.00 Oxygen. Per Cent 49.39 42.10 34.47 28.69 21.14 19.10 6.66 4.46 All natural solid fuels contain more or less earthy or inorganic matter which is not combustible and therefore remains as ash, while the organic matter is consumed. Sometimes the percentage of ash is so great as to render them valueless for steam-making purposes. Origin and Formation of Fuel: Engng, Aug. 23, 1901; Am. Geol., Feb., 1899; Col. Guard, Sept. 10, 1897, Oct. 1, 1897, Jan. 14, 1898, Jan. 28, 1898, March 18, 1898, Sept. 14, 1900; Ec. Geol., Oct., 1905; Eng. U.S., April 1, 1903; Ir. and Coal Td. Review, Feb. 4, 1898, July 13, 1906. 9. Coal. — Coals are most satisfactorily classified according to the constituents of the combustible, as Fixed Carbon. Volatile Matter. Anthracite Semi-anthracite Semi-bituminous Bituminous, Eastern Bituminous, Western Per Cent 97 to 92.5 92.5 to 87.5 87.5 to 75 75 to 60 65 to 50 Under 50 Per Cent 3 to 7.5 7.5 to 12.5 12.5 to 25 25 to 40 35 to 50 Lignite Over 50 Classification of Coals: Am. Inst, of Min. Engrs., May, 1906, Sept., 1905; Mines and Minerals, Dec, 1906; Min. Rept., Apr. 26, 1906; Col. Guard, July 6, 1900; Power, Oct., 1906. 10. Anthracite. — This is the most perfect form of coal and consists almost entirely of carbon; it contains very little hydrocarbon and burns with little or no smoke, is slow to ignite, burns slowly, and breaks into small pieces when rapidly heated. It requires a very large grate of about twice the surface necessary for bituminous coal. Large sizes may be burned in almost any kind of a furnace and with moderate draft. For small sizes a thinner bed has to be carried unless a strong draft is used. There is difficulty in keeping it free from air holes. When 16 STEAM POWER PLANT ENGINEERING possible, the coal should be at least six inches deep on the grates. On account of the large percentage of ash in the smaller size, the fire requires frequent cleaning. Anthracite does not require " slicing " and should be disturbed only when cleaning is necessary. Small Size Anthracite : Eng. and Min. Jour., Dec. 22, 1904. Heat Value of Anthra- cite, Small Sizes, and the Best Way of burning it : Col. Guard, Nov. 26, 1897. Anthra- cite Coal Mines and Coal Mining : Rev. of Rev., July, 1902. The Screening of Anthra- cite : Col. Guard, Sept. 20, 1901. Preparation of Anthracite : Mines and Min., March, 1905. Anthracite Washeries : Am. Inst, of Min. Engrs., Nov., 1905. Anthracite Coal Fields of Pennsylvania: Min. Mag., March, 1905. Virginia Anthracite : Eng. News, Oct. 20, 1904; Mines and Min., March, 1906. Burning of Anthracite Culm of Poor Quality : Trans. A.S.M.E., 7-390. The Use of Electricity in Anthracite Mining : Eng. and Min. Jour., Feb. 2, 1907. Anthracite is classed and marketed according to sizes, the following division of mesh being adopted as standard at Wilkesbarre in 1891 : Egg Stove Chestnut Pea Buckwheat Rice coal must pass through 2| inch mesh and not through 2 inch 2 H Sizes over " pea coal " are prohibitive in price for steam power plant use and consequently the demand is limited to the smaller sizes. 11. Semi-Anthracite. — This coal kindles more readily and burns more rapidly than anthracite. It requires little attention, burns freely with a short flame, and yields great heat with little clinker and ash. It is apt to split up on burning and waste somewhat in falling through the grates. It swells considerably but does not cake. It has less density, hardness, and metallic luster than anthracite, and can generally be dis- tinguished by its tendency to soil the hands while pure anthracite will not. 12. Semi-Bituminous. — This coal is softer than semi-anthracite. It ignites easily and burns freely under a moderate draft. It gives an intense fire and is an excellent steam coal, but is apt to smoke con- siderably unless special provision is made to prevent it. 13. Bituminous. — This coal contains a large and varying amount of volatile matter and requires careful firing to prevent smoke and clinker. Its physical properties vary widely, so much so that it is usually divided into three grades: 1. Dry Bituminous coal is sometimes known as the free-burning bituminous. It is hard and dense, black in color, but brittle and FUELS AND COMBUSTION 17 splintery. It ignites somewhat slowly, burns freely with a short, clean, bluish flame, little smoke, and without caking. 2. Bituminous Caking coal swells up, becomes pasty, and fuses together in burning. It contains less fixed carbon and more volatile matter than the free-burning grades. Caking coal is rich in hydrocarbon and is particularly adapted to gas making. The flame is of a yellowish color. 3. Long Flaming Bituminous coal is similar in many respects to the caking coal but contains a larger percentage of volatile matter. It is free burning, with a long, yellowish flame. It may be either caking or non-caking. COAL FIELDS OF THE UNITED STATES. Alabama: Mines and Minerals, May, 1901. Arkansas: Eng. and Min. Jour., Sept. 12, 1903, Oct. 28, 1905. Colorado: Min. Rept., Jan. 19, 1905, March 2, 1905; Jour. W. S. Engrs., Dec, 1903 ; Mines and Minerals, May, 1905. Illinois: Min. Mag., March, 1905; Eng. and Min. Jour., Jan. 13, 1906. Indiana: Eng. Rec, Jan. 27, 1906; Min. Mag., March, 1905; Power, July and Aug., 1902. Indian Territory : Min. Rept., May 17, 1906. Kansas: Eng. and Min. Jour., Dec. 7, 1901. Kentucky : Col. Guard, Sept. 7, 1900. Michigan: Eng. and Min. Jour., June 30, 1900; Min. World, Feb. 9, 1907. Missouri: Am. Inst, of Min. Engrs., Jan., 1905. Montana: Min. Mag., March, 1905; Min. World, Nov. 24, 1906. Ohio : Min. Mag., March, 1905. Pennsylvania : Eng. and Min. Jour., Aug. 24, 1901; Trans. A.S.M.E., 4-217; Min. Mag., March, 1905; Pro. Eng. S. W. Penn., Jan., 1907. Texas : Mines and Min., Oct., 1905. Virginia: Mines and Min., March, 1906; Eng. News, Oct. 20, 1904. West Virginia: Eng. and Min. Jour., May 12, 1904. Wyoming : Min. World, May 6, 1905. GENERAL. Coal Mines of the United States : Peabody Atlas, A. Bement, Chicago, 111., Min. World, May 6, 1905. Coal Resources of the Pacific : Eng. Mag., May, 1902. Rocky Mountain Coal Fields: Min. Rept., Jan. 5, 1905; Jour. Asso. Eng. Soc, Dec, 1902. Coal Fields, U.S. Northwest : Rev. of Rev., Feb., 1903. Coal Fields, U.S. Southwest : Eng. and Min. Jour., Oct. 17, 1903. U.S. Coal Fields : Steam Boiler Practice, Wm. Kent. Report of Coal Testing Plant : U.S. Geological Survey, Washington, D.C. 18 STEAM POWER PLANT ENGINEERING 14. Lignite, or brown coal, is a substance of more recent geological formation than coal and represents a stage in development intermediate between coal and peat. Its specific gravity is low, 1.2, and when freshly mined it contains as high as 50 per cent of moisture. It is non-caking and gives a bright but slightly smoky flame. It is a low-grade fuel and is used where good coal is difficult to get. Exposure to the air causes it to split into fine pieces like air-slacked lime. It is very fragile and will not bear much handling in transportation. Eng. and Min. Jour., Nov. 22, 1902, Feb. 7, 1903; Mines and Min., July, 1901. Lignite of Northeastern Wyoming : Mines and Min., Feb., 1907. 15. Peat, or Turf, is formed by the slow carbonization under water of a variety of accumulated vegetable materials. It is unsuitable for fuel until dried. Peat as ordinarily cut and dried is too bulky for com- mercial competition with coal, and is used only where coal is prohibitive in price. When properly prepared and compressed into briquettes peat is an excellent fuel. In Russia, Germany, and Holland peat briquettes have passed the experimental stage and several millions of pounds are manufactured annually. Peat is used but little in this country at present, but its possibilities are beginning to attract the attention of engineers. The proportion in which the various primary constit- uents exist in dried peat is approximately as follows: m Per Cent. Fixed carbon 35 Volatile matter 60 Ash 5 Peat: Power, Sept., 1907; Engr. U.S., April 1, 1905; Min. World, Sept. 30, 1905; Col. Guard, Nov. 30, 1900 ; Mines and Min., July, 1901 ; Eng. and Min. Jour., Nov. 22, 1902; Feb. 7, 1903, Eng. Rec, 52-191; Sci. Am. Sup., March 2, 1907; Elec. Engr., Lond., Dec. 6, 1907. Fuel Briquetting : Jour. Assn. Eng. Soc, Jan., 1906; Iron Age, April 19, 1906; Power, Dec, 1902; Eng. and Min. Jour., Nov. 8, 1902; Mines and Min., Oct., 1904; Power, March, 1905; Engr. U.S., May, 1905. 16. Wood, Straw, Sawdust, Bagasse, Tanbark. — In certain locali- ties cord wood is still used as a fuel, but the steadily increasing values of even the poorest qualities are rapidly prohibiting its use for steam- generating purposes. Sawdust, shavings, tanbark and other waste products of wood are burned under boilers in situations where such disposition nets the best financial returns. Recent progress, however, in industrial chemistry shows that ethyl and wood alcohols and other valuable by-products can be cheaply made from sawdust, shavings, FUELS AND COMBUSTION 19 slashings and similar waste material, and it is not unlikely that their use for steaming purposes will be unheard of in a comparatively few years. Table 00 gives the physical and chemical characteristics of a number of woods. Wood as Fuel: Prac. Engr. U. S., Jan., 1910, p. 805; Power & Engr., June 30, 1908, p. 1015; Power, Dec, 1908, p. 772. Burning Sawdust: Prac. Engr. U. S., Jan.. 1910, p. 48; Power & Engr., April 7, 1908, p. 536; Oct. 13, 1908, p. 613; Jour, of Elec, Oct., 1905. TABLE 00. PHYSICAL AND CHEMICAL PROPERTIES OF WOODS, STRAW AND TANBARK. (Prac. Engr. U. S., Jan., 1910.) o 3 . 3 -r. ° £ ^ 3 > a ,1 0) CD X O = 3 fP t-, I* < Calorific value, B.T.U. per Pound. >> o "3 < Ash 46 43 45 42 41 3520 3250 2880 3140 2350 2350 1220 4500 3310 3850 3850 3310 1920 2130 2130 1920 3310 1920 1420 1300 1190 1260 940 940 580 1800 1340 1560 1540 1340 970 1050 1050 970 1340 970 5450 5400 5580 5420 5400 5400 6410 5400 .5460 5460 5400 5460 6830 .6660 6660 6830 5460 6830 Beech Birch Cherry 49.36 50.20 6.01 6.20 42.69 41.62 0.91 1.15 1.06 0.81 Sharpless Hutton Sharpless Elm 35 Hemlock 25 Sharpless Hutton Maple /Hard 1 49 Oak Live 59 tt " 'White. 52 " Red 45 49.64 5.92 41.16 1.29 1.97 Rankine Hutton Pine, White 25 " Yellow 3fi a a Poplar Spruce Walnut 36 25 35 25 49.37 6.21 41.60 0.96 1.86 it it tt Willow .... 49.96 5.96 6.06 39.56 41.30 0.96 1.05 3.37 1.80 Rankine Average. '. 49.70 Straw. Wheat . . . Barley . . . o -1-3 CO Water 16.00 15.50 15.75 35.86 36.27 36.06 5.01 5.07 5.04 37.68 38.26 0.45 0.40 0.42 5.00 4.50 4.75 5155 Clark tt Average 37.97 Tanbark Dry 51.80 6.04 40.74 1.42 9500 Mvers Compressed. 20 STEAM POWER PLANT ENGINEERING Bagasse, or megass, is refuse sugar cane and is used as a fuel on the sugar plantations. Its heat value depends upon the proportions of fiber, molasses, sugar and water left after the extraction. The heat furnished by the different constituents is about as follows: Fiber, 8325 B.T.U. per pound; sugar, 7223 B.T.U. per pound; and molasses, 6956 B.T.U. per pound. Table gives the heat value of bagasse and TABLE 0. HEAT VALUES OF BAGASSE AND VARIATION WITH DEGREE OF EXTRACTION. Is 15 2 .1 i o £ Ph Fiber. Sugar. Molasses. £6 •Sh - a . _, "^ o o H Heat required to evaporate the Water present. B.T.U. cs P3 Lbs. Bagasse required to equal lib. Coal of 14,000 B.T.U. Cal- orific Power. Coal Equivalent per Ton of Cane. Pounds. 6 o Ph d ^ CQ CM w d i gj of. Is t>H S3 a En o> P. 90 0.00 28.33 100.00 66.67 8325 5552 8325 5900 339 8325 5561 1.68 2.52 119 119 2465° 85 3.33 240 1.67 116 2236 80 42.50 50.00 4160 5.00 361 2.50 174 4697 509 4188 3.34 120 2023 75 51.00 40.00 3330 6.00 433 3.00 209 3972 611 3361 4.17 120 1862 70 56.67 33.33 2775 6.67 482 3.33 232 3489 679 2810 4.98 120 1732 65 60.71 28.57 2378 7.15 516 3.57 248 3142 727 2415 5.80 121 1612 60 63.75 25.00 2081 7.50 541 3.75 261 2883 764 2119 6.61 121 1513 55 66.12 22.22 1850 7.78 562 3.88 270 2682 792 1890 7.40 121 1427 50 68.00 20.00 1665 8.00 578 4.00 278 2521 815 1706 8.21 122 1350 45 69.55 18.18 1513 8.18 591 4.09 284 2388 833 1555 9.00 122 1284 40 70.83 16.67 1388 8.33 601 4.17 290 2279 849 1430 9.79 123 1222 25 73.67 13.33 1110 8.67 626 4.33 301 2037 883 1154 12.13 124 1077 15 75.00 11.77 980 8.82 637 4.41 307 1924 899 1025 13.66 124 1002 76.50 10.00 832 9.00 650 4.50 313 1795 916 879 15.93 126 906 variation with the degree of extraction, bagasse is shown in Fig. 4a. A typical furnace for burning Bagasse as Fuel: Prac. Engr. U. S., Jan., 1910; Engr. U. S., April 1, 1903; Jour. Assn. Engng. Soc, July, 1901; Engng., Feb. 18, 1910. Tanbark is usually quite moist; the amount of moisture varies with the leaching process used and averages around 65 per cent. In this condition it has a heat value of about 4300 B.T.U. per pound. If perfectly dry its heating power is approximately 6100 B.T.U. per pound. As in the case of all moist fuels, tanbark must be surrounded by heated surfaces of sufficient extent to insure drying out the fresh fuel as thrown on the fire. A successful furnace for burning tanbark is shown in Fig. 4b. Tanbark as a Boiler Fuel: Jour. A.S.M.E., Feb., 1910, p. 181; Jour. A.S.M.E., Oct., 1909, p. 951 ; Prac. Engr. U. S., Jan., 1910. FUELS AND COMBUSTION 21 22 STEAM POWER PLANT ENGINEERING H FUELS AND COMBUSTION 23 17. Composition of Coal. — The uncombined carbon in coal is known as fixed carbon, while the hydrocarbons and other gaseous compounds which distill off on application of heat constitute the volatile matter. Refractory earths and moisture are found in varying quantities in different classes of coal and as they are incombustible tend to reduce the heat value of the fuel. That part of the fuel which is dry and free from ash is called the combustible, though the nitrogen and oxygen in the volatile matter are not actually combustible. The term " pure coal " has been suggested in this connection and is meeting with much favor. (Jour. W.S.E. 11-757.)* The various elementary constituents of a fuel must be determined by a careful chemical analysis, but in most cases it is only necessary to know the heating value, the per cent of moisture and ash and perhaps the per cent of sulphur. Table 1 shows the composition of a number of American coals and gives a good idea of their chemical characteristics. TABLE 1. COMPOSITION OF TYPICAL AMERICAN COALS. (U.S. Geological Survey.) Anthracite. Semi-Bituminous. ii . . * tuo 3 r 3 a n * § !^ 3 X £ PS Is o III H - 3 .2 Proximate analysis Water Volatile matter . . . Fixed carbon Ash Sulphur Ultimate analysis Carbon Hydrogen ....... Nitrogen 1.97 4.35 86.49 7.19 0.64 85.66 2.78 0.77 2.87 13963 1.50 7.84 81.07 9.59 0.50 83.20 3 = 29 0.95 2.45 i3954 2.08 7.27 74.32 16.33 0.77 75.21 2.81 0.80 4.08 12472 12395 0.65 18.80 75.92 4.63 0.57 85.91 4.58 1.07 3.24 15190 15104 0.59 18.52 74.31 6.58 0.81 81.05 4.91 2.15 4.57 14484 1.28 12.82 73.69 12.21 2.01 77.29 3.74 1.39 Oxygen Calorific value Calorimeter Dulong's Formula 3.36 13406 13831 * H. J. Williams. * See also " Unit Coal and the Composition of Coal Ash, No. 37, Aug. 9, 1909. Univ. of 111. Bulletin 24 STEAM POWER PLANT ENGINEERING TABLE 1.- - Continued. Bituminous. Lignite. cu ■s as . pq •ill ill gas GO o o -2 CO << ft * § 3 >> £> CO O 9 fl s 43 a o o c o Pi Per Cent of C0 2 in the Flue Gas by Volume. 6 8 10 12 14 16 0.2 328 2.2 248 1.7 199 1.3 168 1.1 144 1 126 0.8 0.4 635 4.3 484 3.3 390 2.6 327 2.2 282 1.9 248 1.7 0.6 925 6.3 709 4.8 575 3.9 474 3.2 417 2.8 367 2.5 0.8 1.0 1192 8.1 923 6.3 750 5.1 635 4.3 549 3.7 495 3.4 1494 10.2 1128 7.7 923 6.3 780 5.3 676 4.6 596 4.1 1.2 1690 11.5 1321 9 1085 7.4 923 6.3 801 5.4 708 4.8 1.4 1920 13.1 1512 10.3 1248 8.5 1061 7.2 924 6.3 819 5.6 1.6 2104 14.3 1693 11.5 1400 9.5 1193 8.1 1040 7.1 924 6.3 1.8 2340 16 1865 12.7 1549 10.5 1321 9.0 1151 7.8 1025 7 2.0 2537 17.2 2030 13.8 1690 11.5 1450 9.9 1270 8.6 1129 7.7 Large type gives the loss in B.T.U. per pound of carbon. Small type gives the per cent loss, assuming a calorific value of 14,650 B.T.U. per pound of carbon. should consist of carbon dioxide and nitrogen only, and in the ratio by volume of 21 to 79; therefore 21 per cent of C0 2 in the flue gas is indicative of complete combustion and theoretical air supply. In other words, the ratio by volume of C0 2 to N after complete com- bustion is practically the same as the ratio of the oxygen to the nitrogen in the air before combustion. Table 3 gives the approximate weight 36 POWER PLANT ENGINEERING of air used per pound of combustible for different percentages of C0 2 in the flue gas. In practice, 15 per cent is all that can be expected under the best conditions, with an average between 10 per cent and 12 per cent. Any- thing less than 10 per cent shows an excessive amount of air supplied. Traveling grates, unless carefully operated, are apt to show as low as 5 per cent of C0 2 . 24. Loss due to Incomplete Combustion. — If the volatile gases are not completely oxidized, as when the air supply is insufficient or the mixture of air and gases is not thorough, some of the carbon may escape as CO. The presence of even a small amount of CO in the flue gas is indicative of a very appreciable loss, as will be seen from Table 5. Carbon monoxide is a colorless gas, and its presence in the chimney gases cannot be detected by the fireman, consequently the absence of smoke is not an infallible guide for perfect combustion. This loss may be expressed , _ „ 10,150 CO K ~ ° X (C0 2 + CO)' (7) in which h 2 = the loss in B.T.U. per pound of carbon, C0 2 and CO are percentages by volume of the flue gases and C is the proportional part of carbon in the combustible. 15 14 13 C J-i 12 11 10 9 S 7 6 5 4 Relation of Gas Composition in Rear Combustion Chamber To Temperature at Same Place ■2i 3 2 1 _C u_ 0.5 o 0.4 t 0.3 J 0.2 O o.i a 1900 2000 2100 2200 2300 2400 2500 2600 Combustion Chamber Temperature.Deg.Fah. 2700 Fig. 5. Relation of Gas Composition in Combustion Chamber to Temperature. This loss, however, may be wholly avoided in a properly designed and carefully operated furnace. In fact the loss from this cause is often exaggerated and seldom exceeds 2 per cent of the total heat value of the fuel except during the few moments following the replenishing FUELS AND COMBUSTION 37 of a burned-down fire with fresh fuel or when the supply of air is checked to meet a sudden reduction in load. In improperly designed furnaces in which the volatile gases are brought into contact with the cooler boiler surface before combustion is complete, the carbon monoxide may be reduced in temperature below its ignition point and consequently will fail to combine with the oxygen. In such a case the loss may prove to be a serious one. Fig. 5 shows the relation between the com- position of the products of combustion in the rear combustion chamber of a 250-horse-power Heine boiler, hand fired, and the temperature at the same place. (For an extended discussion of this subject see Jour. West. Soc. Engrs., June, 1907, p. 285.) 25. Loss of Fuel through Grate. — The refuse from a fuel is that portion which falls into the pit in the form of ashes, unburned or partially burned fuel, and cinders. The loss from this cause depends upon the size of the fuel, the width of opening in the grate bars, and the type of grate. Coal which necessitates frequent slicing is apt to give greater loss than a free-burning coal. Under good conditions of operation it ought not to exceed 2 per cent of the total heat value of the fuel. In traveling grates in which a large percentage of the fine fuel falls through the front end of the grate a special hopper is ordinarily installed in the ash pit which reclaims most of it. (See Fig. 99.) Loss of Fuel in Ashes : Power, March, 1905. Experiments on Fuel Value of Bitumi- nous Coal Ashes: Technology Quarterly, Dec, 1905. Coal Ash: Jour. Soc. Chem. Ind., Jan. 15, 1904. 26. Superheating the Moisture in the Air. — The loss due to this cause is a minor one, though on hot, humid days it may be appreciable. This loss may be expressed h 3 = 0.48 M(T - t)* (8) in which h 3 = B.T.U. lost per pound of combustible. M = weight of moisture introduced with the air per pound of combustible. t = temperature of air entering the furnace, degrees F. T = temperature of chimney gases, degrees F. 27. Moisture in the Fuel. — Moisture in the fuel represents an appre- ciable loss in economy if present in large quantities, since the heat necessary to evaporate it into superheated steam at chimney temperature is lost. Firemen occasionally wet the coal to assist coking or to reduce * The latest accepted value for the mean specific heat of water vapor at atmos- pheric pressure is 0.46 in place of 0.48. 38 STEAM POWER PLANT ENGINEERING the dust, but moisture thus added necessarily reduces the furnace efficiency. The loss due to this cause is expressed: h 4 = M [(212 - t) + 966* + 0.48f (T - 212)], (9) in which h 4 = B.T.U. lost per pound of combustible. M = weight of moisture per pound of combustible. Other notations as in preceding equation. For example, the heat loss due to the moisture in a pound of fuel containing 10 per cent water, temperature of fuel 80 degrees F., chimney temperature 480 degrees F., is h t = 0.1 [(212 - 80) + 966* + 0.48f (480 - 212)] = 122.6 B.T.U. A rough rule is to allow a loss of 1 per cent of the total heat value of the dry fuel for each 10 per cent of moisture present. 28. Loss due to the Presence of Hydrogen in the Fuel. — The hydro- gen in any fuel which is not rendered inert by oxygen burns to water and in so doing liberates 62,032 B.T.U. per pound. All of this heat is not available for producing steam in the boiler, since the water formed by combustion is discharged with the flue gases as superheated steam at chimney temperature. This loss is equal to h 5 = 9H [(212 - t) + 966* + 0.48f (T - 212)], (10) in which h 5 = B.T.U. lost per pound of combustible. H = weight of hydrogen per pound of combustible. All other notations as in equations (8) and (9). With anthracite coal this loss is approximately 2.5 per cent of the total heat value of the combustible and with bituminous coal it runs as high as 4.5 per cent. 29. Loss due to Smoke. — Visible smoke consists of carbon in a flocculent state mixed with the products of combustion. It is seldom evident in connection with anthracite coal and is generally associated with bituminous fuel. A smoky chimney does not necessarily indicate an inefficient furnace, since the losses due to visible smoke generation seldom exceed 2 per cent; as a matter of fact, a smoky chimney may be much more economical than one which is smokeless. That is to say, a furnace operating with minimum air supply may cause dense clouds of smoke and still give a higher evaporation than one made smokeless by a very large excess of air. There will be some loss due to carbon monoxide and unburned carbon or soot in the former case, but this may be more than offset by the excessive losses caused by the * See footnote, p. 88. f See footnote, p. 37. FUELS AND COMBUSTION 39 heat carried away in the chimney gases in the latter. Judging from the results of the majority of steam power plants using bituminous coal, even those recently installed, smokeless and efficient combustion is not readily effected and the problem is far from being satisfactorily solved. Smoke has become such a public nuisance, particularly in the larger cities, that special ordinances prohibiting its production have been enacted and violators are subject to heavy fines. Effective enforce- ment of these ordinances renders smoke production very costly and the problem of smokeless combustion becomes a momentous one. The subject of smoke prevention and smoke-prevention devices is discussed at some length in Chapter V. 30. Radiation and .Minor Losses. — These losses are usually deter- mined by difference. That is, the difference between the heat repre- sented in the steam and the losses just mentioned is charged to radia- tion, leakage, and unaccounted for. Summing up the various losses we have Heat given to steam Loss in chimney gases Loss due to carbon burning to CO Loss of fuel through grate Loss due to moisture in coal, moisture in air, and hydrogen in fuel Smoke, soot, etc Radiation and minor losses Excellent Good Average Practice. Practice. Practice. Per Cent. Per Cent. Per Cent. 80 70 60 12 18 24 1 2 0.5 1 2 3.0 3 3 0.5 1 4.5 6.5 8 Poor Practice. Per Cent. 50 30 3 3 3.5 1.5 9 31. Size of Coal. — Bituminous. Coal is usually marketed in different sizes, ranging from lump coal to screenings. The latter furnish by far the greater part of the stoker fuel used. For maximum efficiency coal should be uniform in size. With hand-fired furnaces there is usually no limit to its fineness and larger sizes can be used than with stokers. As a rule the percentage of ash increases as the size of coal decreases. This is due to the fact that all of the fine foreign matter separated from larger coal, or which comes from roof or floor of the mine, naturally finds its way into the smaller coal. The size best adapted for a given case is dependent upon the intensity of draft, kind of stoker or grate, and the method of firing, and its proper selection often affords an opportunity to effect considerable economy. Some idea of the influence of the size of screenings on the capacity and efficiency of a 40 STEAM POWER PLANT ENGINEERING boiler in a specific case is illustrated in Fig. 6. The curves are plotted from a series of tests conducted with Illinois screenings on a 500-horse- power B. & W. boiler, equipped with chain grates, at the power house of the Chicago Edison Company. Influence" of Thickness of Fire. — See paragraph 76. Size of Coal: Some Characteristics of Coal as affecting Performances with Steam Boilers : Jour. West. Soc. Engrs., Oct., 1906, p. 528. Small Size Anthracite : Eng. and Min. Jour., Dec. 22, 1904. The Economy of Small Size Coals for Power Plants : Eng. Mag., Feb., 1905. 1000 1.25 1.00 0.75 Size of Coal in Inches 0.50 0.25 Fig. 6. Influence of Size of Coal on Boiler Capacity and Efficiency. 32. Washed Coal. — The washing of coal is for the purpose of separating from it such impurities as slate, sulphur, bone coal, and ash. All of these impurities show themselves in the ash when the coal is burned. Screenings contain anywhere from 5 per cent to 25 per cent FUELS AND COMBUSTION 41 of ash and from 1 per cent to 4 per cent of sulphur. Washing eliminates about 50 per cent of the ash and some of the sulphur. Table 6 gives some idea of the effects of washing upon a number of grades of coal. The evaporative power of the combustible is practically unaffected by washing and the greater part of the water taken up by the coal is removed by thorough drainage. Many coals otherwise worthless as steam coals are rendered marketable by washing. Washed coals are usually graded as follows : Size. Screens. No. 1 Over If Under 2£ 2 11 U 3 1 It 4 * I 5 i Numbers 3 and 4 are excellent sizes for use in connection with stokers and No. 5 is well adapted for hand furnaces where smoke prevention is essential. TABLE 6. EFFECT OF WASHING ON BITUMINOUS COALS. (Journal W.S.E., December, 1901.) Before Washing. (Per Cent.) Ash. Sul- phur. Fixed Carbon. After Washing. (Per Cent.) Ash. Sul- phur. Fixed Carbon. Belt Mountain, Mont Wellington Colliery Co., Van- couver Island (new coal) . . . Alexandria Coal Co., Crabtree, Pa DeSoto, 111 , Northwestern Improvement Co., Roslyn, Wash Luhrig Coal Co., Zaleski, Ohio Rocky Ford Coal Co., Red Lodge, Mont Buckeye Coal and Ry. Co., Nelsonville, Ohio New Ohio Washed Coal Co., Carterville, 111 18.74 35.00 10.60 18.00 16.30 15.80 25.30 13.77 9.48 3.34 43.72 38.00 1.30 0.57 1.90 44.00 45.90 1.05 0.78 37.80 49.04 55.00 5.56 8.90 6.21 4.20 9.70 8.00 8.50 4.30 4.85 2.40 0.61 0.40 0.87 0.89 0.69 48.39 56.90 57.00 47.86 50.90 47.20 54.82 63.00 42 STEAM POWER PLANT ENGINEERING Modern Method of Coal Washing : Eng. and Min. Jour., May 9, 1903, Oct. 13, 1904. Principles of Coal Washing : Mines and Min., Aug., 1903. Bituminous Coal Washing : Mines and Min., April, 1905. Washing of Bituminous Coals by Luhrig Process : Jour. West. Soc. Engrs., Dec., 1901. Coal Washing : Jour. Soc. Chem. Ind., April 30, 1904. Anthracite Washeries : Eng. and Min. Jour., April 28, 1906; Col. Guard, April 20, 1906; Trans. Am. Inst. Min. Engrs., Nov., 1905. Studies on Coal Wash- ing : Col. Guard, Nov. 21, 1902. Coal Washing by Stuart System : Mines and Min., Dec, 1903. Coal Washing at Collinsville, III: Mines and Min., Sept., 1901. Bellevue Washery of D. L. and W. R.R. Co., Scranton, Pa.: Mines and Min., June, 1903. Coal Washery at Howe, Indian Territory : Mines and Min., March 14, 1904. Eastern Coal and Coke Co.'s Washery at Kansas : Eng. and Min. Jour., Sept. 20, 1902. 33. Purchasing Coal.* — Engineers fail to agree as to the specifica- tions best suited for the purchase of coal. Some extensive purchasers require elaborate analyses and others specify only the size and grade of the fuel. Every essential requirement of the purchaser may be fulfilled by confining them to the four following characteristics : Moisture. Ash. Size of coal. Calorific value of the coal. Although moisture is a great and uncertain variable, and the producer can exercise no control over this factor, still the purchaser should pro- tect himself against excessive moisture by stipulating an amount con- sistent with the average inherent moisture in the coal, and proper penalty should be fixed for delivery in excess of the amount allowed, a corre- sponding bonus being paid for delivery of less than contract amount. Considerable attention should be given to the percentage of earthy matter contained. The amount of earthy matter usually fixes the heating value of the coal, since the heating value of the combustible is practically constant. The effect of ash on the heat value of Illinois screenings as fired under a B. & W. boiler with chain grate is shown in Fig. 7. This value varies with the different types of boilers, grates, and furnaces, but is substantially as illustrated. The amount of refuse in the ash pit is always in excess of the earthy matter as reported by analysis. The maximum allowable amount of sulphur is sometimes specified, since some grades of coal high in sulphur cause considerable clinker- ing. But sulphur is not always an indication of a clinker -producing ash, and a more rational procedure would be to classify a coal as clinkering or non-clinkering according to its behavior in the particular furnace in question, irrespective of the amount of sulphur present. An analysis of the various constituents of the ash is necessary to * See also Selection of Coal for Boiler Furnaces, by D. T. Randall, Power & Engr., Apr. 6, 1909, p. 642. FUELS AND COMBUSTION 43 determine whether or not the sulphur unites with them to produce a fusible slag, and as such analyses are usually out of the question on account of the expense attached, they may well be omitted. The heating value of the coal as determined by a sample burned in an atmosphere of oxygen does not give its evaporative power, since this 100 90 80 TO g50 a 8 40 •20 10 Influence of Ash on Fuel Value of Dry Coal. (Illinois Screenings) B.& W. Boiler, Chain Grate. Screenings with 12.5 Per Cent Ash taken at 100. \ \ \ Jour.S.W.E .Oct.l 906 P. 542. \ 10 20 30 40 Per Cent of Ash in Dry Coal Fig. 7. Influence of Ash on Fuel Value of Dry Coal. depends largely upon the composition of the fuel, character of grate, and conditions of operation. It merely serves as a basis upon which to determine the efficiency of the furnace. In large plants where a number of grades of fuel are available it is customary to conduct a series of tests with the different grades and sizes, and the one which evaporates 44 STEAM POWER PLANT ENGINEERING the most water for a given sum of money, other conditions permitting, is the one usually contracted for. In designing a new plant particular attention should be paid to the performance of similar plants already in operation, and that fuel and stoker should be selected which are found to give the best returns for the money. Where smoke prevention is a necessity the smoke factor greatly influences the choice of fuel and stoker. See paragraph 416. * Testing and Purchasing Coal for Steam Plants : Eng. News, Feb. 7, 1907; Eng. Rec, Sept. 22, 1906, p. 326; Engr. U.S., Aug. 15, 1907; Bulletin No. 339 U.S. Geological Survey, 1908. Coal for Hand-Fired Furnaces: Nat. Engr., July, 1909. 34. Powdered Coal. — The value of powdered coal as a fuel for steam boiler plants has long been known, and appliances for pulverizing and feeding the coal have been on the market for a number of years. How- ever, despite the many advantages of powdered fuel and the apparent success of some of the systems of burning it, little progress has been made toward its general adoption. Some of the advantages obtained in burning powdered coal are: a. Complete combustion and total absence of smoke. The coal in the form of dry impalpable dust is induced or forced into the zone of combustion, where each minute particle is brought into contact with the necessary amount of air and complete oxidation is effected without the excess of air which accompanies the firing with lump coal, provided the furnace is properly proportioned. With a properly designed setting there is complete absence of smoke. b. A cheaper grade of bituminous coal may be burned, since the per cent of ash and moisture has little effect on the completeness of combustion and the full value of the fuel is more nearly realized than with ordinary firing. c. The plant may be rapidly forced above its rated capacity and sudden demands for power readily met. d. The labor of firing is reduced to a minimum. Pulverized Fuel : Eng. Mag., Jan., 1908; Jour. West. Soc. Engrs., Feb., 1904; Am. Elecn., Sept., 1901. Coal Dust for Steam Making : Engr. U.S., Feb. 15, 1899. Burn- ing Pulverized Coal: Eng. and Min. Jour., Dec. 31, 1903, May 12, 1906; Jour. Assn. Eng. Soc., July, 1903. Use of Pulverized Coal under Steam Boilers: Eng. News, April 1, 1904; Power, March, 1904, April, 1904. Coal Dust Firing: Eng. and Min. Jour., Dec. 16, 1905. Coal Dust Fuel: Engr., Lond., Jan. 31, 1896; Engr. U.S., April 1, 1903; Eng. News, Feb. 20, 1902. 35. Depreciation of Powdered-Coal Furnaces. — To withstand the intense heat of combustion, brickwork of the highest quality is essential, since common fire brick are soon reduced to a liquid slag. A good quality FUELS AND COMBUSTION 45 of fire brick will withstand the heat for several months without renewals provided the furnace is properly enclosed, otherwise the strain of expansion and contraction due to alternate heating and cooling will crack the brick. Excellent results have been obtained from the use of bricks composed chiefly of the refuse from a carborundum slag, but the high cost has prevented their general use. 36. Storing Powdered Fuel. — Most cities limit the storage of pow- dered coal to such a small quantity as to prohibit the use of fuel feeders of the " dust feed " type in plants of any size not provided with a pulverizing and crushing system. Coal dust mixed with air is often claimed to be of an explosive nature and many accidents are reported to have resulted from this cause. Many engineers, however, refute this on the basis of experiments which show that explosion can only occur at temperatures high enough to drive off the volatile gases.,* 37. Rate of Combustion with Powdered Fuel. — In forcing large quan- tities of dust into the furnace the velocity imparted to the particles may be so great as to carry them beyond the zone of combustion before oxida- tion is complete, with the result that the flues, and the back of the fur- nace, become covered with unconsumed carbon. So much depends upon the depth of the furnace and the arrangement of the regenerative surface that no specific figures can be given as to the maximum rate of combus- tion that can be efficiently effected. At ordinary rates of combustion the small particles of fuel are completely oxidized while in the combustion chamber and there is total absence of smoke. The use of anthracite coal is practically excluded from this type of stoker unless mixed with coal high in volatile matter. This is due to the fact that fixed carbon burns more slowly than the hydrocarbon gases and the temperature of ignition is higher, hence the most gentle draft will carry away the particles before they are completely consumed. With fuels high in vola- tile matter the hydrocarbons are distilled at a comparatively low tem- perature, forming an inflammable gas which burns rapidly with the fixed carbon. A mixture of 30 per cent bituminous and 70 per cent anthra- cite has been successfully burned in the powdered form. 38. Cost of Pulverizing Coal. — In stokers of the " Aero Pulverizer " type in which the grinding and feeding are carried on simultaneously in a self-contained apparatus, the power consumed varies from 2 per cent to 10 per cent of the total power developed, depending upon the nature of the fuel, the load factor, the efficiency of the driving mechan- ism, and the degree of fineness of the powdered fuel; 5 per cent is a fair average. The best results are obtained when 95 per cent of the dust will pass a 100 mesh and 75 per cent a 200 mesh, though satis- factory results have been obtained with as low as 40 mesh. Powdered * See Fuel, Jan. 12, 1909, p. 294. 46 STEAM POWER PLANT ENGINEERING coal in the open market ranges from 25 cents to 50 cents a ton above the price of the same coal in the form of screenings. 39. Efficiency of Powdered-Coal Furnaces. — Table 7 gives the results of a comparative test of a 140-horse-power Babcock & Wilcox boiler, hand fired, vs. coal-dust feeder. The test was conducted by the engineering staff of the McCormick Harvester Company at Chicago, Illinois, and the results were obtained with boilers working under normal conditions. The dust apparatus was a modified " Ruhl " feeder, and was installed by the C. 0. Bartlett & Snow Company of Cleveland, Ohio. In this particular test the efficiency of the boiler is very low for both hand fired and dust-feed, but the dust-feed test shows an efficiency of 10 per cent greater than that of the hand fired. TABLE 7. COMPARATIVE TEST OF 140-HORSE-POWER BABCOCK & WILCOX BOILER. Hand Fired vs. Pulverized Fuel. Boilers fired by Date Duration of test Total water evap. into dry steam from and at 212 degrees Average gauge pressure Average feed-water temperature, Fahr. . . . Average stack-gas temperature, Fahr Kind of coal used Cost of coal delivered in boiler room ready to fire Total weight of dry coal consumed Per cent of ash in coal determined by lab- oratory analysis Per cent of ash as removed from ash pit and furnace Heating value of coal Water evap. per pound of fuel, actual con- ditions Equivalent water evap. from and at 212 degrees per pound of dry fuel Equivalent water evap. per pound com- bustible Horse power developed Dry fuel per hour per square foot grate surface Equivalent water evap. per hour per square foot heating surface Cost per 1,000 pounds water evaporated (for fuel ready to fire only) Efficiency of boiler and furnace based on coal ; machine 2-24r-04 8 hours hand 2-3-04 8 hours 70,070 79.8 1b. 169.7 506 Westville, Indi- 45,673 79.4 1b. 172.1 458 Westville ana, screenings pulverized to 40 mesh screenings. $2.10 9,373 $1.72 8,413 17.5 19.54 none 12,555 20.57 11,300 6.822 1b. 4.595 lb. 7.476 1b. 5.429 1b. 9.132 lb. 254 6.941 lb. 165.5 19.27 18.65 3.128 2.039 $0.1455 $0,177 55.5 per cent 41.5 per cen FUELS AND COMBUSTION 47 A comparison of a number of tests of hand-fired and powdered-coal furnaces with different types of feeders shows a decided gain in efficiency of the powdered coal over the hand-fired where the fuel is of a low grade. The gain becomes less marked with fuel of fair quality and disappears entirely with good fuel and properly manipulated automatic stokers. A test made by G. H. Barrus on a 250-horse-power B. &. W. boiler at the General Electric Works in connection with a coal-dust feeder manu- factured by the Phcenix Investment Company of New York gave a boiler and furnace efficiency of 75.3 per cent. Subtracting from this the power consumption of 5 per cent for operating the crusher and feeder, the net efficiency was 70.1 per cent. A test of a 135-horse-power return tubular boiler with this same stoker gave a combined efficiency of boiler and furnace of 80 per cent. These figures, however, have been equaled and even exceeded in special hand-fired automatic stoker tests, and only a comparative test of the two systems under similar conditions will show their respective efficiencies. Tests of Pulverized Fuel: Engr. U.S., April 1, 1904; Engr. Lond., Jan. 31, May, 1904; Power, May, 1904. Comparative Boiler Tests with Ordinary and Pulverized Coal Firing : Eng. Rec, March 12, 1904. 40. Furnaces for Burning Powdered Coal. — In burning ordinary bulk coal the mass of incandescent fuel stores up a sufficient quantity of heat to effect the distillation and ignition of the volatile matter in the green fuel. With pulverized coal a refractory lining is necessary to bring about the same result. In arranging a furnace for burning powdered coal in connection with a burner of the forced draft type, the grate bars are removed, ash and fire doors bricked up, and the nozzle bricked in tightly. The lower surfaces of the tubes are covered and the whole forms a reverberatory furnace. With the natural draft system of burner, a suitable opening is left in the brick lining of the ash door to allow the necessary amount of air for combustion to enter. Considerable difficulty is found with delivery nozzles in the formation of slag in the outlet and in their rapid destruction on account of the intense heat. A water-jacketed cast-iron nozzle is said to satisfactorily overcome these objections. 41. Draft for Powdered Fuel. — A study of a number of tests of boilers burning powdered coal shows that the necessary draft is very low and ranges from 0.05 to 0.2 of an inch of water and averages not far from 0.1 inch. 42. Types of Powdered-Coal Burners. — Powdered-coal burners may be grouped into two general classes: , 1. The dust-feed burner, in which the coal is supplied in the powdered form, and 48 STEAM POWER PLANT ENGINEERING 2. The self-contained burner, in which the coal is crushed, pulver- ized, and fed to the furnace simultaneously. The dust may be fed into the furnace by 1. Natural draft. 2. Mechanical means, or by 3. Forced draft. The following outline gives a classification of a few of the best known coal-dust burners : Natural Draft Forced Draft Natural Draft Feed Brush Feed Blower Feed fpinther j Wegener Schwartzkopff /Cyclone "(Triumph Dust Feed Compressed Air /Eng and Powdered (Fuel Company Paddle Wheel /Ideal (.Aero-Pulverizer Self-contained 43. Pinther Apparatus. — Fig. 8 shows a section through a Pinther coal-dust feeder, illustrating the principles of the " natural draft feed " type. The powdered coal is placed into hopper B, from which it is fed by rollers a, a into the chamber leading to the furnace C. The dust falls in a thin stream and is caught up by the current of air and drawn into the furnace as indicated. The furnace is lined with refractory material heated to a sufficiently high temperature to ignite the fuel and burn it in suspension. The chief drawback to a burner of this type is its limited capacity. Any attempt to feed large quantities of fuel into the furnace necessitates Fig. 8. Pinther Coal-dust Feeder. guc ] 1 & s t r0n g current of air as to carry the particles of dust beyond the zone of combustion before they are completely consumed. Within the limits of its capacity it is an efficient and simple apparatus, but is open to the same objection as all FURNACE FUELS AND COMBUSTION 49 FURNACE burners of this type in that it necessitates the storage of powdered coal. This apparatus is not much in evidence in boiler plants. 44. Schwartzkopff Apparatus. — Fig. 9 shows a section through a Schwartz kopff feeder, illustrating the principles of the brush-feed, naturaWraft system. It is a very simple and practical dust feeder, though open to the objec- tion of all systems which require the coal to be ground and pulver- ized in separate machines. The fuel is placed in a hopper and its supply to the brush is regulated by the hand screw A and the spring plate bottom of the hopper. The brush, consisting of a number of flat steel leaves ^ inch by J inch wide, revolves at a high speed, 1000 to 1200 r.p.m. and forces the dust into the furnace. The air for combus- tion is admitted either through the grates in the ordinary way or through the lower chamber of the burner. To prevent the dust from bridging in the hopper, a small hammer C is fitted to the brush so that it will strike the plate D and agitate the dust. This apparatus is meeting with much success in connection with annealing furnaces, but is still in the experimental state as far as boiler firing is concerned. 45. Aero-Pulverizer Apparatus. — Fig. 10 gives a general view of the Aero-Pulverizer Company's apparatus, and is a typical example of a self- contained system. It is very compact, occupying a floor space of only 30 by 77 inches, and is capable of burning 300 to 1500 pounds of coal per hour. It consists essentially of four interior communicating chambers of successively increased diameter in which paddles revolve on. arms with corresponding increased radii. The largest chamber contains a fan, the function of which is to draw the pulverized material successively from one chamber to another and to finally deliver it through the exit in the fan chamber under the impetus of a forced draft. There are two adjustable inlets for air at the feed of the machine through which is introduced the amount of air required for pulverizing purposes. The apparatus may be belt driven or direct connected and runs at about Fig. 9. Schwartzkopff Coal-dust Feeder. 50 STEAM POWER PLANT ENGINEERING 1600 r.p.m., requiring from 6 to 15 horse power for its operation. It is a complete dust fuel feeding system on one bed plate comprising a pul- verizer, fan, coal feeder, hopper, and air dampers. The operation is as follows: Coal previously crushed to nut size is fed to the hopper from TOP CASING THROWN BACK FOR INSPECTIO COAL HOPPER AUTOMATIC FEEDER Fig. 10. Aero-Pulverizer Coal-dust Feeder. the bottom of which it is transferred, with the necessary air for com- bustion, to the pulverizer chamber. The coal, passing into the pul- verizer, is thrown out radially by centrifugal force, due to the rapidly revolving arms and bats, and is reduced to a dust by percussion and attrition. The dust is moved over the ends of the bats and into the fan chamber from which it is blown into the furnace. This apparatus will successfully pulverize and feed coal containing as much as 10 per cent moisture. Fig. 11. Triumph Coal-dust Feeder. 46. Triumph Apparatus. — Fig. 11 illustrates the Triumph coal-dust feeder as designed by the C. O. Bartlett & Snow Company, Cleve- land, Ohio. FUELS AND COMBUSTION 51 The coal is fed from storage bin to hopper A and feed worm B. The latter forces it down spout F directly to delivery tube D, where it is caught by the air draft and fed into the furnace. The amount of feed depends upon the speed of the feed worm, which is driven by the friction disk / against the flange plate H. This disk is moved in or out by handle so as to get any speed desired. The air is furnished by fan C, the amount being controlled by valve E. DESCRIPTION OF COAL-DUST BURNERS. Aero-Pulverizer System: Eng. News, Nov. 28, 1901, p. 415; Eng. Rec, May 25, 1901, p. 506; Power, March, 1904. Cyclone System: Engr. U.S., April 1, 1903, p. 272; Eng. News, Nov. 28, 1901, p. 415; Power, March, 1904. Davis Pulverizer : Jour. Asso. Eng. Soc, July, 1903; Engr. U.S., April 1, 1903. Ideal: Am. Elecn., April, 1902, p. 196; Power, March, 1904. Miscellaneous Coal Dust Burners: Am. Elecn., Sept., 1901, p. 434; Engr., Lond., Sept. 11, 1896; Engng., Jan. 15, 1897; Power, Aug., 1903; St. Ry. Rev., Vol. 8-187, 1898; Engr. U.S., April 1, 1904. Rowe : Engr. U.S., Jan. 1, 1903, p. 93; April 1, 1903, p. 272; Eng. News, Dec. 25, 1902, p. 548; Eng. Rec, Dec. 20, 1902, p. 591. Schwartzkopff : Am. Elecn., Jan., 1902; Eng. News, Feb. 20, 1902; Power, March, 1904. Wegener : Cassier's, March, 1896, p. 501; Power, March, 1904; Eng. Mag., March, 1896, p. 1158, Aug., 1896, p. 964, Oct., 1898, p. 125; Eng. News, Sept. 16, 1897, p. 189, 47. Fuel Oil. — The recent development of oil wells in the Western and Gulf States, with the consequent enormous increase in production, has given a marked impulse to the use of crude oil for fuel purposes in steam power plants. Where economic and commercial conditions permit, it is the most desirable substitute for coal. The total absence of smoke and ashes, prompt kindling and extinguishing of fires, extreme rate of combustion, and ease with which it can be handled and con- trolled are marked advantages in favor of fuel oil. The reduction in volume and weight over an equivalent quantity of coal for equal heat- ing values and the increase in boiler efficiency are factors of no mean importance, particularly in connection with marine or locomotive work. In stationary work the chief objections are the difficulty in securing ample storage capacity and the increased rate of insurance. An objec- tion sometimes raised against oil fuel is the increased depreciation of the setting, but in a well-designed setting this figure is only nominal and of secondary importance. However, in spite of the many advantages presented in the use of fuel oil for power plant purposes, the limited supply and constant fluctuation in price prevent its adoption as a general fuel, and limit its use to the plants most favorably located. 52 STEAM POWER PLANT ENGINEERING Crude Oil Burning : Power, March, 1907; Engr. U.S., Dec. 15, 1905, March 1, 1906; Am. Elecn., Aug., 1903, p. 396; Engng., March 28, 1902, p. 140; Eng. Mag., May, July, Sept., 1903; Eng. News, June 19, 1902, p. 501; Cassier's, May, 1901, p. 61; Engr., Lond., Dec. 9, 1904; St. Ry. Jour., May 10, 1902, p. 588; Am. Gas Light Jour., May 12, 1902, p. 695, . TABLE 8. ANALYSES OF TYPICAL AMERICAN FUEL OILS. Authority. Physical Properties. Location. Specific Gravity. 60°-70° F. Flash Point. Deg. F. Burn- ing Point. Deg. F. Specific Viscosity. 60° F. 185° F. California — Crude Ed. O'Neill ....do 0.9533 0.9572. 299.6 373 1.17 4.7 Do Do do 0.7825 0.9670 0.866 62 196 52 64.5 221 77 Do ....do Kansas — Crude B. F. McFarland. C. E. Coates Louisiana — Crude Ohio — Distillate Deville 0.887 0.838 0.826 0.886 0.841 Do N. W. Lord Deville 177 212 Pennsylvania — Crude . . Pennsylvania — Distillate W. Virginia — Crude. . . . Wyoming — Crude Texas — Crude do ..do Colburn Denton 0.92 0.926 142 216 181 240 Texas — Distillate U. S. Naval Re- port Location. Authority. Chemical Properties. C H O+N S B.T.U. per Lb. California — Crude Do Ed. O'Neill do 85.75 86.3 11.3 10.7 0.668 0.8 18,797 18,646 Do ... .do Do .. ..do Kansas — Crude B. F. McFarland. C. E. Coates 85.4 13.07 Louisiana — Crude 0.34 19,814 18,718 19,880 17,930 19,210 18,400 19,590 19,060 Ohio — Distillate Deville 84.2 13.1 2.7 Do N. W. Lord Pennsylvania — Crude . . Pennsylvania — Distillate W Virginia — Crude .... Deville 82 84.9 84.3 14.8 13.7 14.1 3.2 1.4 1.6 ....do ....do Wyoming — Crude Texas — Crude Denton 84.6 83.26 10.9 12.41 2.87 3.83 1.63 0.50 Texas — Distillate U. S. Naval Re- port 19,481 48. Chemical and Physical Properties of Fuel Oil. — From Table 8 it will be seen that the physical properties of oils from different localities in the United States differ widely, while the chemical constituents vary FUELS AND COMBUSTION 53 but slightly. For example, the oils given in the table differ greatly in volatility, specific gravity, and viscosity, but have nearly a constant ratio of carbon and hydrogen and consequently vary but slightly in heating value. A good deal of the oil produced is unfit for fuel purposes unless refined. The chief impurities are sulphur, earthy matter, and water. Besides interfering with the free burning of the oil, moisture and sulphur have a deleterious effect upon the boiler and furnaces, and should not be present in large quantities. Where the percentage of sulphur is greater than 4 per cent, the depreciation of the boiler and furnace offsets the gain in using the lower grade fuel. Many successful processes of removing the water and sulphur are on the market, and consequently crude oil high in sulphur should not be used unless the depreciation element has been taken into consideration. Oil that is to be transported or stored or used for fuel inside of buildings should be of the " reduced " variety, from which the naphtha and higher illuminating products have been distilled. The gravities of such distillates vary from 20 to 25 degrees Baume, or close to 0.9 specific gravity, and their flash points range from 240 degrees. F. to 270 degrees F. This variation in volatility has little effect on the heat value of the oil, since the ratio of carbon to hydrogen varies but slightly in the various distillates. One barrel of oil contains 42 gallons and weighs from 310 to 332 pounds according to the specific gravity. Compared with coal, oil occupies about 50 per cent less space and is 35 per cent less in weight for equal heat values. The comparative heat values of coal and oil are approximately as follows: B.T.U. per Pound of Coal. Pounds of Coal Equal to 1 Barrel of Oil. Barrels of Oil Equal to 1 Short Ton of Coal. 10,000 11,000 12,000 13,000 14,000 15,000 620 564 517 477 443 413 3.23 3.55 3.87 4.19 4.52 4.84 Technical Aspects of Oil as Fuel: Junge, Power, Oct., 1907, p. 665. Petroleum Oil Fields : Jour. Soc. Chem. Ind., Oct. 15, 1902, p. 1228. Investiga- tion on American Petroleum : Am. Chem. Jour., March, 1896, p. 215. The Constit- uents of Pennsylvania, Ohio, and Canada Petroleums: Amer. Chem. Jour., Vol. 19-419. Composition of California Petroleum : Amer. Chem. Jour., Vol. 19-796, Vol. 25-253. Composition of Petroleum: Amer. Chem. Jour., Vol. 28-165, 33-251. Composition 54 STEAM POWER PLANT ENGINEERING of Texas Petroleum: Jour. Amer. Chem. Soc, Feb. 9, 1901, p. 264; Soc. Chem. Ind., 19-121, 20-237, 690. Flashing Points of Petroleum : Jour. Soc. Chem. Ind., 15-341. Origin of Petroleum : Jour. Soc. Chem. Ind., 16-727, 1898. Influence of Water on Flash Test and Combustion Point of Petroleum : Chem. News, 85-267. The Relation between Some Physical Properties of Bitumens and Oils : Eng. Rec, Aug. 18, 1906. 49. Efficiency of Boilers with Fuel Oil. — From Table 14, it will be seen that 70 per cent is a high figure for boiler efficiency in regular service when good coal is burned, and 65 per cent a fair average. With liquid fuel an average efficiency of 4 to 6 per cent above this is readily attained. (See Table 9.) This increase in efficiency is partly due to the fact that the oil is readily broken up and brought into intimate contact with the necessary air for combustion and the loss due to excess of air is correspondingly reduced. The results of tests made by the Liquid Fuel Board of the U.S. Navy show that oil has an efficiency of 5 per cent greater than coal for the same rate of evaporation, and that the boiler capacity may be increased 50 per cent above that possible when burning coal and still maintain the same efficiency. The max- imum efficiency with oil was attained at a higher rate of evaporation than was the maximum efficiency when coal was burned. TABLE 9. BOILER EFFICIENCIES, OIL FUEL. X 0) a i— ( Authority. Reference. Quality of Oil. Evapora- tion from and at 212° F. Pounds. Efficiency of Boiler and Fur- nace. Per Cent. 1 2 Pacific Light and Power Co., Los Angeles, Cal. Eng. Record, Aug. 6, 1904. California Crude, 18,607 B.T.U. per pound. California Crude, 18,760 B.T.U. 16.02 15.66 83.06 80.6 3 4 U.S. Naval Board 1902 Report of U.S. Naval Liquid Fuel Board. Reduced Beau- mont, 19,480 B.T.U. 20,000 B.T.U 14.43 16.9 71.5 77.8 5 6 Prof. Williston. . . Engineering Mag- azine, July, 1903. West Va. Crude, 20,960 B.T.U. Texas Crude, 18,850 B.T.U. 16.5 15.9 76 76.8 7 Prof. Denton. . . . Power, Feb., 1902. Beaumont Texas, 19,060 B.T.U. 15.5 78.5 8 Wallsend Engineering, Nov. 6, 1902. Not stated 14.45 FUELS AND COMBUSTION 55 50. Comparative Evaporative Economy of Oil and Coal. — In deter- mining the comparative economy of coal and oil, the fixed and operating charges must be considered in addition to the cost and efficiency of the fuel. From the market quotation on oil and coal and the com- parative heating values of each the actual cost per B.T.U. is readily obtained, and by combining this with the relative efficiencies from the furnace standpoint the net cost of the fuel is obtained. The fixed charges vary with the location and size of the plant and are approxi- mately the same per boiler horse power for a given location in both cases. The insurance rates may be greater with the oil fuel and the depreciation of the boiler setting may be somewhat larger, but in a well- constructed furnace the latter item should be the same in both instances for average rates of combustion. The operating charges are decidedly in favor of the oil fuel, since no ash handling is necessary. Oil fuel is readily fed to the furnace, and the cost of attendance may be materially less than with coal firing, and one man may safely control from eight to ten boilers. Table 106, Chapter XVII, gives data relative to the cost of producing electrical power in connection with oil-fired steam plants. Tests of Crude Oil as a Fuel : Cassier's, May, 1901, p. 61 ; Power, Feb., 1902, p. 8; Eng. U.S., Jan. 16, 1905, p. 90, Feb. 1, 1905; Eng. Rec., Dec. 20, 1902;Eng. Mag., July, 1902, p. 615; Eng. News, July 11, 1901, p. 23; Eng. Rec, Aug. 6, 1904, p. 175; Oct. 29, 1904, p. 502. 51. Oil Burners. — The function of the burner is to atomize the oil to as nearly a gaseous state as possible. Classification of a few well-known burners: Mechanical Spray : Korting. Vapor or Carburettor : Durr. Harvey. Spray Burners : Outside Mixers. a. Booth. b. Warren. Inside Mixers. a. Hammel. b. Kirkwood. c. Branch. d. Williams. 56 STEAM POWER PLANT ENGINEERING Oil burners for burning liquid fuel may be divided into three general classes : 1. Mechanical spray, in which the oil previously heated to a tem- perature of about 150 degrees F. is forced under pressure through nozzles so designed as to break it up into a fine spray. The Korting Liquid Fuel Burner, Fig. 12, is an example of this type. In this (~\ \ ffi|yr~ ""jtffHil design a central spindle, spirally grooved, imparts a rotary motion to the oil and causes it to fly into a spray by centrifugal force on issuing from the nozzle. The particles of oil are burned in the furnace when they come in contact with the neces- sary air to effect combustion. This type of burner is little used in this country in connection with power plant work, but is meeting with much success on the continent. 2. Vapor burners, or carburettors, in which the oil is volatilized in a heater or chamber and then admitted to the furnace, are seldom used except in connection with refined oils, as the residuals from crude oil are vaporized only at a high temperature. The Durr and Harvey gasifiers are the best known of this type. 3. Spray burners are by far the most common in use. In this type the oil is held in suspension and forced into the furnace by means of a jet of steam or compressed air. Spray burners are designed either as outside mixers, in which the oil and atomizing medium meet outside the apparatus, or inside mixers, in which the oil and atomizing medium mingle inside the apparatus. Fig. 12. OIL Korting Fuel Oil Burner. OIL '&aaawzzz£azzz££&zuaa&4. izEzzzzzznzzzzzzzzzzmzzzzzzzzzzzzzzzpzz L STEAM Fig. 13. Booth Fuel Oil Burner. The Booth burner, Fig. 13, illustrates the principles of the " outside mixer " type of apparatus and is in use on the Santa Fe Railroad. In this type the oil flows through a thin slit and falls upon a jet of steam FUELS AND COMBUSTION 57 which atomizes it and forces it into the furnace. A feature of this apparatus is its simplicity of construction and freedom from clogging. Fig. 14 illustrates the Hammel burner as used at the power house of the Pacific Light and Power Company, Los Angeles, Cal. Oil enters the G.I. H.- Fig. 14. Hammel Fuel Oil Burner. burner under pressure and flows through opening D to the mouth of the burner where it is atomized by the steam jets issuing from slots G, H, STEAM 5sr«»dl3 p OIL Fig. 15. Branch Fuel Oil Burner. and 7. The oil is preheated to facilitate its flow through the supply system. Plates K-K are removable and are easily replaced when worn out or burned. The Hammel burner belongs to the " inside mixers." 58 STEAM POWER PLANT ENGINEERING A few well-known types of " inside mixers " are illustrated in Figs. 14 to 16. The operation is practically the same in all of them and they differ only in mechanical details. Fig. 16. Kirkwood Fuel Oil Burner. Fig. 17. Williams Fuel Oil Burner. The Williams burner, Fig. 17, differs somewhat from the others in that the air supply passes through the burner and mingles with the oil and steam before entering the furnace. FUELS AND COMBUSTION 59 The simplest and most reliable burners are of the Hammel type and are much in evidence in the Pacific States. Notes on Oil Burners using Compressed Air : Power, Nov., 1904. Report of U.S. Naval Liquid Fuel Board : Engr. U.S., Dec. 1, 1904. Oil Burners : Engng., April 15, 1904; Am. Engr. and R.R. Jour., Sept., 1901. • J40ILPIPE % STEAM PIPE OIL-* Fig. 18. Warren Fuel Oil Burner. 52. Furnaces for Burning Oil Fuel. — Fig. 19 shows the construction of a typical oil-burning furnace as applied to a 250-horse-power B. & W. water tube boiler in the power plant of the Union Loop Elevated Station, Chicago. For the successful burning of oil the furnace should be so constructed that oxidation of the fuel is complete before it reaches the tubes. This is effected by arranging the refractory lining to form a sort of reverberatory furnace in which the atomized oil is vaporized and mixed with the necessary air for combustion. The air is preheated in its passage beneath the lower lining of the furnace and the supply is regulated by a suitable damper. The regulation of air, steam, and oil for the burner is a very delicate operation and requires considerable skill for efficient results.* In the particular furnace illustrated in Fig. 19 the flame impinges against a cellular wall of fire brick before it reaches * For a modern and highly efficient system of oil fuel feeding and regulation, see Power, Dec. 29, 1908, p. 1108. 60 STEAM POWER PLANT ENGINEERING the bridge wall. The bricks are loosely stacked and are readily removed when burned out. They tend to save the lining of the bridge wall arch and insure a more intimate mixture of air and oil in the combustion chamber. The supply of crude oil is limited compared with that of coal and the price is subject to sharp fluctuations, and conse- quently the cost may prove prohibitive for fuel purposes. To be prepared for such an emergency, many engineers design the furnaces for coal burning and arrange them with loose brick-work for the temporary use of oil. Fig. 20 illustrates the application of a Hammel burner at the rear end of a furnace. Oil~Fired Furnaces : Engr. U.S., July 1, 1902, p. 491, Nov. 15, 1905; Eng. Mag., May, 1903; Engng., April 15, 1904, p. 523, April 29, 1904, p. 594. Liquid Fuel Com- bustion: Trans. A.S.M.E., May, 1902. Equipment for Oil Fuel: Eng. and Min. Jour., Oct. 7, 1905. 53. Air vs. Steam as an Atomizing Medium. — Table 10 gives the results of a series of tests made by the U.S. Naval Liquid Fuel Board in 1902 on different types of burners using air, steam, or both for atomizing the fuel. The first eight tests were made with compressed air as the spraying medium and under pressures varying from 0.78 pounds to 4.68 pounds per square inch. The most economical results were °5 FUELS AND COMBUSTION 61 obtained with the lower pressures, and the total steam used to compress the air varied from 1.06 per cent to 7.45 per cent of the total steam generated, but not all burners would work with air at this low pres- sure. With steam as the spraying medium the steam required to Fig. 20. Furnace for burning Fuel Oil, Rear Feed. operate the burner varied from 3.98 per cent to 5.77 per cent of the total generated, while the burner using both steam and air required 6.09 per cent to 8.54 per cent of the total. The results of recent tests with the latest types of burners give somewhat lower steam consumption than the tests conducted by the Naval Board and a good average is not far from 3 per cent.* Table 10 contains also the results of a few scattering tests made with different types of burners. In general it may be said that where a supply of low-pressure air is available, air is unquestionably more economical than steam as an atomizing medium, but in the average boiler plant the use of steam obviates complication and risk of interrupted service. Where it is necessary to use high-pressure air the economy decreases with the increase in pressure, since the cost of each cubic foot of compressed air increases rapidly with the pressure, but its ability to atomize the oil does not increase proportionately. Steam vs. Air for Liquid Fuel and Oil Burners : Am. Mach., Vol. 27, No. 51. * See Proceedings, A. S. M. E., Dec., 1908, p. 1698. 62 STEAM POWER PLANT ENGINEERING •fttrao J9d) Jsnog jo iCouaptga •p9}T3J9U9S UlB91g imoj, JO ^1190 J9d t no Sui^jdg ui pasri uii39ig QOMlOrHTttOOO^OSNt^WlHN 000«OOONW>0>ONOOiHCOtJIOS f CO CTiNiOOOiO©»CO>CON N0»0'* (•^ S99lg9Cl) Uinip8J\[3UI -A'-Bidg jo ajnj'Bjaduiax \\q Sui^Bidg joj pasn ranip8J\[ JO 9JT1SS8JJ (93ni3£) ipui gjenbg J9d SpunOj) 9JT1SS91J UI129Jg •^S9X jo J9qmnx 00 t-h 10 o o O O t~ T-H ^ ^ "^ CO ^ c^ lO lO lO to «o lO lO iOiHN(Nl>-* O©NQ005O>-<(NC0'1<>O!ON00 FUELS AND COMBUSTION 63 54. Oil Pressure. — This varies with the different types of burners and ranges from a few pounds to 60 pounds or more per square inch. The low-pressure systems are ordinarily operated under standpipe pres- sures as in Fig. 21, which illustrates the arrangement of apparatus as advocated by the International Gas and Fuel Company. A steam pump B draws the oil from the buried tank through pipe Z and delivers it to the standpipe E. Thence it flows through pipe / to the burners under a Fig. 21. International Gas and Fuel Company's Fuel Oil System. head of about 10 feet. The pump runs constantly, the surplus oil flow- ing back to the tank through the pipe T. The oil is heated by the exhaust pipe Z'. The oil pump is provided with a device D having a piston connected by a chain with a cock S, which automatically opens when the boiler is not under steam pressure, so that the standpipe will be emptied, the oil flowing to the storage tank. The high-pressure systems are invariably operated by steam pumps, usually in duplicate, and are so arranged that the oil pressure will be kept practically constant irrespective of the steam pressure. The adjust- ment of steam and oil is a very delicate operation, and fluctuation in the steam pressure disturbs the proportion of oil and steam; to prevent this 64 STEAM POWER PLANT ENGINEERING the steam pressure at the burner is reduced several pounds below that of the boiler by suitable reducing valves and is thereby kept at a nearly constant value. 55. Oil Storage and Transportation. — Distillates or reduced oils are readily stored and transported, but the crude oils, on account of the inflammability of the highly volatile elements, offer a different problem. In most cities distillates may be stored in large quantities but only in tanks sunk below the lowest level of the surrounding territory. This is a protection against flooding the district with burning oil in case of a fire. In the country the oil is ordinarily stored in tanks above the ground level and at some distance from the plant. Fig. 22 illustrates the Hydraulic Oil Storage Company's system of storing oil and delivering it to the burners. The oil reservoirs are placed SIPHON BREAKER 1% OIL TO BURNERS DISCHARGE TO SEWER FLOOR LINE Fig. 22. Hydraulic Oil Storage Company's Fuel Oil System. below grade as indicated to minimize fire risk. The operation is as follows: Water enters the " float box" and flows through a " three-way cock " to the bottom of the reservoir until all of the oil and water pipes are filled up to the level of the float box, when the float auto- matically cuts off the supply. This flooding of the entire system FUELS AND COMBUSTION 65 drives out all of the air. The three-way cock is then turned to " discharge " and part of the water flows to the sewer. The tank- car or wagon is next attached to the " oil inlet " and the oil flows into the tank and displaces the water until the level of the " filler float " is reached, when the supply is automatically cut off. The inlet is so placed that the head of oil in the tank-car is sufficiently great to over- come the opposing head of water. The three-way valve is next turned to the first position and the head of water forces the oil to the burners. After the oil has been withdrawn from the storage tank the water can only rise to the level of the water in the float box and therefore cannot be fed to the furnace. The small steam pipe admits steam into the tank and heats the oil, thereby making it flow more freely. Storing Oil Fuels : Eng. News, Oct. 29, 1903, p. 396. Petroleum Reservoirs : Jour. Soc. Chem. Ind., Jan. 31, 1899. Handling Fuel at Railway Terminals : Eng. News, Sept. 25, 1902, p. 232. 56. Conclusions of U. S. Naval Liquid Fuel Board. — After a series of elaborate tests it was concluded a. That oil can be burned in a very uniform manner. b. That the evaporative efficiency of nearly every kind of oil per pound of combustible is probably the same. While the crude oil may be rich in hydrocarbons, it also contains sulphur, so that, after refining, the distilled oil has probably the same calorific value as the crude product. c. That a marine steam generator can be forced to even as high a degree with oil as with coal. d. That up to the present time no ill effects have been shown upon the boiler. e. That the firemen are disposed to favor oil, and therefore no impediment will be met in this respect. /. That the air requisite for combustion should be heated if possible before entering the furnace. Such action undoubtedly assists the gasification of the oil product. g. That the oil should be heated, so that it could be atomized more readily. h. That when using steam higher pressures are undoubtedly more advantageous than lower pressures for atomizing the oil. i. That under heavy forced draft conditions, and particularly when steam is used, the Board has not yet found it possible to prevent smoke from issuing from the stack, although all connected with the tests 66 STEAM POWER PLANT ENGINEERING made special efforts to secure complete combustion. Particularly for naval purposes, it is desirable that the smoke nuisance be eradicated in order that the presence of a war ship might not be detected from this cause. As there has been a tendency of late to force the boilers of industrial plants, the inability to prevent the smoke nuisance under forced-draft conditions may have an important influence upon the increased use of liquid fuel. /. That the consumption of liquid fuel cannot probably be forced to as great an extent with steam as the atomizing agent as when com- pressed air is used for this purpose. This is probably due to the fact that the air used for atomizing purposes, after entering the furnace, supplies oxygen for the combustible, while in the case of steam the rarefied vapor simply displaces air that is needed to complete combustion. k. That the efficiency of oil-fuel plants will be greatly dependent upon the general character of the installation of auxiliaries and fittings, and therefore the work should be intrusted only to those who have given careful study to the matter and who have had extended experience in burning the crude product. The form of the furnace will play a very small part in increasing the use of crude petroleum. The method and character of the installation will count for much, but where burners are simple in design and are constructed in accordance with scientific principles there will be very little difference in their efficiency. Con- sumers should principally see that they do not purchase appliances that have been untried and have been designed by persons who have had but limited experience in operating oil devices. 57. Gaseous Fuels. — These fuels offer all of the advantages of liquid fuels and but few of the disadvantages. The gases most commonly met with in connection with steam power plants are outlined in Table 11. The artificial gases for steam purposes are prohibitive in cost in most cases, and even in blast-furnace installations, where the gases are waste products, the gas engine has virtually supplanted the steam engine for power purposes. In the immediate locality of natural gas wells gas- fired furnaces may prove to be more economical than coal furnaces, but the limited supply and constant fluctuation in price limit its use as a general fuel. From the market quotations on coal and gas and the comparative heating value of each the actual cost per B.T.U. is readily obtained, and by combining this with the relative efficiencies from the furnace standpoint the net cost of the fuel is obtained. The following table, based upon the assumption that one cubic foot of natural gas under standard conditions has a heating value of 1,000 B.T.U. , will enable an approximate comparison to be made: FUELS AND COMBUSTION 67 B.T.U. per Pound of Coal. Pounds of Coal Equal to 1,000 Cu. Ft. of Gas. No. of 1,000 Cu. Ft. of Gas Equal to One Short Ton of Coal. 10,000 11,000 12,000 13,000 14,000 15,000 100 91 83 77 71 67 20 22 24 26 28 30 Fuel Economy : Fuel Economy in Steam Power Plants : Cassier's Mag., May, 1904; Inst, of Elec. Engrs., Jan. 12, 1905; Engr. U.S., April 1, 1905; Engng., Aug. 7, 1903; Eng. Mag., June, 1907. The Province of the Fuel Expert : Eng. and Min. Jour., May 25, 1905. A Gas-Fired Boiler : Engr. U.S., Feb. 15, 1907, p. 223. See also A.S.M.E. Code for conducting Boiler Tests — reprinted in Appendix B. TABLE 11. CHARACTERISTICS OF GASEOUS FUELS. (Lucke.) Natural gas Cannel-coal gas Common-coal gas Carburetted water-gas. . . Uncarburetted water-gas. , Producer-gas, little steam. Loomis Pettibone coal gas Dowson gas, average Taylor gas, average Mond gas , Coke-oven gas Blast-furnace gas , H 1.7 27.7 39.78 21.8 49.50 9.2 14.0 18.0 21.0 29.0 53.0 3.0 CO 0.55 6.8 7.04- 28.1 35.93 25.3 20.0 25.0 12.0 12.0 6.0 27.5 CH 4 94.16 50.0 45.16 30.7 1.05 3.1 2.0 3.0 2.0 2.0 35.0 C 2 H 4 0.30 13.0 6.38 12.9 0.8 0.20 2.0 0.30 0.06 0.5 0.10 3.6 C0 2 N Cubic Feet of Air per Cubic Foot of Gas. B.T.U. per Cubic Foot of Gas. High. Low. Natural gas 0.29 0.1 1.08 3.8 4.25 3.4 8.2 7.0 6.0 14.5 2.0 10.0 2.80 2.4 0.50 2.2 8.75 58.2 55.5 47.0 57.0 42.5 2.0 59.4 9.13 6.50 6.38 6.00 2.10 1.24 989 843 727 702 295 160 888 Cannel-coal gas 762 Common-coal gas 651 Carburetted water-gas Uncarburetted water-gas Producer-gas, little steam Loomis Pettibone coal gas .... Dowson gas, average 635 265 150 1.32 0.98 1.17 5.06 .81 119 130 156 620 100 115 Taylor gas, average 116 Mond gas 139 Coke-oven gas 524 Blast-furnace gas 99 CHAPTER III. BOILERS. 58. As affecting fuel economy the boiler equipment is by far the most important part of the power plant and involves the largest share of the operating expenses. It matters little how elaborate, modern, or well designed it may be, skill, good judgment, and con- tinued vigilance are required on the part of the operator to secure the best efficiency. Of the various types and grades of boilers on the market experience shows that most of them are capable of practically the same evaporation per pound of coal, provided they are designed with the same proportions of heating and grate surface and are operated under similar conditions. They differ, however, with respect to space occupied, weight, capacity, first cost, and adaptability to particular conditions of operation and location. 59. Classification. — As to design and construction there is an almost endless variety of boilers and furnaces, classified as internally and externally fired; water tube and fire tube; through tube and return tubular; horizontal and vertical. The internally fired type includes the vertical tubular, locomotive, Scotch-marine, and practically all flue boilers. The externally fired includes the plain cylinder, the through tubular, return tubular, and nearly all stationary water-tube boilers. 60. Vertical Tubular Boilers. — Vertical tubular boilers, Figs. 1 and 23, are commonly used where small power, compactness, low first cost, and sometimes portability are the chief requirements, though they are not necessarily restricted to small sizes. The tubes are sometimes arranged so that the spaces between them radiate from a hand hole on one side so that a scraper may readily be inserted to clean the top of the furnace plate. The hand hole in the water leg permits removal of the scale. It is convenient to place a chain in the bottom of the water leg which can be worked around through the hand hole for the purpose of loosening up the scale deposit. The distance between the furnace crown and top of the grate is never less than 24 inches even in the smallest boiler and should be as great as possible to insure good combustion. /Two styles of vertical boilers are in common use, the ordinary vertical 68 BOILERS 69 type, Fig. 1, and the submerged type, Fig. 23. In the former the upper tube sheet and part of the tubes are above the water line, and while this feature may tend to superheat the steam to a slight extent, the difficulty STACK STEAM GAUGE HAND-HOLE WATER COLUMN STAY-BOLTS BLOW OFF Fig. 23. Vertical Tubular Boiler with Submerged Tube Sheet. from unequal expansion and liability to overheating is of sufficient moment to justify the use of the submerged type, particularly where the boiler is likely to be forced above its rated capacity. The advantages of this type of boiler are (1) compactness and portability; (2) requires no setting beyond a light foundation; (3) is a rapid steamer, and (4) is low in first cost. The disadvantages are (1) inaccessibility for thorough inspection and cleaning; (2) small steam space, which results in excessive 70 STEAM POWER PLANT ENGINEERING priming at heavy loads; (3) poor economy except at light loads, as the products of combustion escape at a high temperature on account of the SECTION THROUGH A- SECTION THROUGH ASH PIT SECTIONAL FRONT ELEVATION Fig. 24. Manning Vertical Fire-Tube Boiler. shortness of the tubes; (4) smokeless combustion practically impossible with bituminous coals; (5) the small water capacity results in rapidly fluctuating steam pressures with varying demands for steam. BOILERS 71 Although vertical fire-tube boilers are usually of very small size, being seldom constructed in sizes over 60 horse power, an exception is found in the Manning boiler, Fig. 24, which is constructed in sizes as large as 250 horse power. Many of the disadvantages found in the smaller types are obviated in the Manning boilers, which, as far as safety and efficiency are concerned, rank with any of the other first- class types. They differ from the boiler described above mainly in having the lower or furnace portion of much greater diameter than the upper part which encircles the tubes. This permits a proper proportion of grate, which is not obtainable in boilers like Figs. 1 and 23. The double flanged head connecting the upper and lower shells allows sufficient flexibility between the top and bottom tube sheets to provide for unequal expansion of tubes and shell. The ash pit is built of brick and the water leg does not extend below the grate level, thus doing away with dead water space. Where overhead room permits and ground space is expensive, this boiler offers the advantage of taking up a small floor space as compared with horizontal types. 61. Fire-Box Boilers. — Although vertical fire-tube boilers may be classed as fire-box boilers, yet the term "fire box" is usually associated with the locomotive types, whether used for traction or stationary pur- poses. The usual form of fire-box boiler as applied to stationary work SAFETY VALVE FIRE DOOR ASH DOOR Fig. 25. Typical Fire-box Boiler. — Stationary Type. is illustrated in Fig. 25. The shell is prolonged beyond the front tube sheet to form a smoke box. The front ends of the tubes lead into the smoke box and the rear ends into the furnace or fire box. The fire box 72 STEAM POWER PLANT ENGINEERING is ordinarily of rectangular cross section, and is secured against collapse by stay bolts and other forms of stays. In Fig. 25 the smoke box is of cylindrical cross section and hence requires no staying except at the flat surface. Fire-box boilers are used a great deal in small heating plants where space limitation precludes other types. Their steam capacity gives them an advantage over the vertical tubular form. Being internally fired no brick setting is required. They are usually of cheap construction, designed for low pressure, and seldom made in sizes over 75 horse power. Unless carefully designed and constructed high steam pressures are apt to cause leakage because of unequal expan- sion of boiler shell, tubes, and fire box. Portable fire-box boilers with return tubes are made in sizes as large as 150 horse power and for pressures as high as 150 pounds per square inch, but being more costly than some of the other types of boilers of equal capacity are used only where portability is an essential requirement. 62. Scotch-Marine Boiler. — Where an internally fired boiler is desired for large powers the Scotch-marine type is finding much favor with engineers. A number of the tall office buildings in Chicago are equipped with boilers of this class which are giving good results. They require little overhead room, no brick setting, and are excellent steamers. mt»»w»>w»jwj»»»»»w»>»iMw»w»wm/M»»MM»m r ~ Fig. 26. Stationary Scotch-Marine Boiler. The Continental boiler, Fig. 26, is one of the best known of this type. The boiler is self-contained and requires no brick setting, the only fire brick used being those that form the bridge wall, baffle ring and the layer at the back of the combustion chamber. The furnace and tubes are BOILERS 73 entirely surrounded by water, so that all fire surfaces, excepting the rear of the combustion chamber, are water cooled. The furnace is cor- rugated for its whole length. These corrugations, in addition to giving greater strength to the furnace, act as a series of expansion joints, taking up the strains due to unequal expansion of furnace and shell. Practically all types of mechanical stokers and grates are applicable to these boilers. The advantages of a Scotch boiler and of all internally fired boilers are (1) minimum radiation losses; (2) requires no setting; (3) no leakage of cool air into the furnace as sometimes occurs through cracks or parous brickwork of other types; (4) large steaming capacity for the space occupied. The circulation, however, is not always positive and the water below the furnace may be considerably below the average or normal temperature, giving rise to unequal expansion and contraction which may cause leakage. The boiler proper is relatively costly, but this is offset to some extent by the absence of setting. 63, Robb-Mumford Boiler. — Fig. 27 shows a section through a Robb-Mumford boiler, which is a modification of the Scotch-marine and of the horizontal tubular type. It consists of two cylindrical shells, the lower one containing a round furnace and tubes and the upper one forming the steam drum, the two being connected by two necks. The lower shell has an incline of about one inch per foot from the horizontal, for the purpose of promoting circulation and draft, and also for convenience in washing out the lower shell. Combustion takes place in the furnace, which is surrounded entirely by water, and 74 STEAM POWER PLANT ENGINEERING the gases pass through the tubes and return between the lower and upper shells (this space being enclosed by a steel casing) to the outlet at the front of the boiler. Mingled water and steam circulate rapidly up the rear neck into the steam drum, where the steam is released, the water passing along the upper drum towards the front of the boiler and down the front neck, a semicircular baffle plate around the furnace causing the down-flowing water to circulate to the lowest part of the lower shell under the furnace. The outer casing, which incloses the space between the lower and upper shells, including the rear smoke box and the smoke outlet, is constructed of steel plate, with angle-irqn stiffeners, the various sections being bolted together for convenient removal. The inside of the steel case, including the rear smoke chamber, is lined with asbestos air-cell blocks fitted in between the angle-iron stiffeners. The top of the upper drum and bottom of the lower shell are also covered with non-conducting material after the boiler is erected. Owing to the fact that steam and water spaces are divided between two cylin- drical shells, the thickness of plates is not so great as in the Scotch- marine or horizontal return tubular types; and the rear chamber of the marine boiler is avoided. The chief claim for this type of boiler is compactness. A battery of five 200-horse-power units occupies a floor space of but 33 feet in width by 20 feet in depth and 12.5 feet high. Each unit is entirely independent and may be isolated for cleaning, inspection, and repairs. 64. Horizontal Return Tubular Boilers. — These are the most common in use and are constructed in sizes up to 200 horse power. They are simple and inexpensive and, when properly operated, durable and economical. Figs. 28 to 31 show various forms of standard settings, and Figs. 75, 76, and 76a different " smokeless " settings. The grate is independent of the boiler, and the products of combustion pass beneath the shell to the back end, returning through the tubes to the front, and into the smoke connection. The tubes are 3 to 4 inches in diameter and from 14 to 18 feet long, and are expanded into the tube sheets. The portion of the tube sheets not supported by the tubes is secured against bulging by suitable stays. Access to the interior of the boiler is obtained through manholes. The most convenient arrangement for inspection and cleaning is to have one manhole located at the top of the shell and one at the bottom of the front tube sheet. Return tubular boilers are made either with an extended front (Fig. 28) or flush front (Fig. 29). The latter costs a little more for brick and setting, but it is more convenient to operate and the boiler is less expensive. The shell may be supported by lugs on the brickwork as in Fig. 28 or by steel beams and hangers as in Fig. 30. BOILERS 75 9 -.01 1 ! 76 STEAM POWER PLANT ENGINEERING BOILERS 77 The latter construction permits the brickwork and shell to expand or contract independently, and settling of the brickwork does not affect the boiler alignment. With the side bracket support, the front lugs usually rest directly on iron or steel plates imbedded in the brickwork, and the back lugs on rollers, to permit free expansion and contraction. The brackets are long enough to rest upon the outside wall, so that the inside brick lining can be renewed without disturbing the setting. The distance between the rear tube sheet and wall should be about 16 inches STEEL SUPPORTS Fig. 30. Return Tubular Boiler Setting. — Steel Beam Suspension. for boilers less than 60 inches in diameter and from 20 to 24 inches for larger ones. The distance between grate and boiler shell should not be less than 28 inches for anthracite coal and 36 inches for bituminous coal.* The greater this distance the more complete the combustion, since the gases will have a better opportunity for combining with the air before coming in contact with the comparatively cool surfaces of the shell. The shell should be slightly inclined toward the blow-off end so as to drain freely. The vertical distance between the bridge wall and shell is usually between 10 and 12 inches. The lower part of the combustion chamber behind the bridge wall may be filled with earth and paved with common * For smokeless combustion the setting must be modified. See furnace illustrated and described in paragraph 96. 78 STEAM POWER PLANT ENGINEERING a o O BOILERS 79 brick, as in Fig. 31, or left empty as in Fig. 29. The shape of the walls, whether curved to conform to the shell or flat, appears to have little influence on the economy. The side and end walls are ordinarily constructed of common brick with an inner lining of fire brick, and may be solid as in Fig. 29 or double with air spaces as in Fig. 28. The latter construction is prefer- able and permits the inner and outer walls to expand independently without cracking and settling. The side walls are braced by five pairs of buck-staves, with through rods under the paving and over the tops of the boilers. The connection between the rear wall and the shell is a source of more or less trouble on account of the expansion and contraction of the boiler. Cast-iron supports of T section supporting a fire-brick arch are usually employed as illustrated in Fig. 32, the clearance between the arch and the shell being sufficient to allow the necessary expansion. Fig. 32. Furnace Arch Bars. Fig. 33. Back connection made with Cast-iron Plate. Fig. 33 shows the common method of resting one end of the arch supports on the rear wall and the other end on an angle iron riveted to the boiler. The products of combustion are sometimes carried over the top of the boiler as shown in Fig. 31. This tends to superheat the steam, but the advantage gained is probably offset considerably by the extra cost of the setting and the accumulation of soot on the top of the shell. The arrangement is not common. The steam connection is naturally made to the highest point in the boiler shell. Frequently a steam dome, to which the steam nozzle is connected, is provided as in Fig. 29. The function of the steam dome is to increase the steam space so as to permit the collection of dry steam at a point high above the water level. If a boiler is too small for its work 80 STEAM POWER PLANT ENGINEERING 7e « BOILERS 81 and is forced far above its rating a steam dome is probably an advantage, though its use is less common now than formerly, since a properly designed boiler insures ample steam space without one. A dry pipe inside the boiler above the water line as in Fig. 26 or 27 is commonly used to guard against priming where the nozzle is connected to the shell. For low pressures and small powers the return tubular boiler has the advantage of affording a large heating surface in a small space and large overload capacity. It requires little overhead room and its first cost is low. On the other hand the interior is difficult of access for purposes of cleaning and inspection. Boilers of this type are seldom constructed in sizes above 150 horse power or for pressures over 150 pounds per square inch, since the cost increases rapidly as the pressure rises above this amount. 65. Babcock & Wilcox Boiler. — Fig. 34 shows a longitudinal section through a Babcock & Wilcox boiler, illustrating a typical horizontal water-tube type. The tubes, usually 4 inches in diameter and 18 feet Fig. 35. Details of Header, — Babcock and Wilcox Boiler. Fig Front Section, — Babcock and Wilcox Boiler. in length, are arranged in vertical and horizontal rows and are expanded into pressed-steel headers. Two vertical rows are fitted to each header and are " staggered " as shown in Fig. 35. The headers are connected 82 STEAM POWER PLANT ENGINEERING with the steam drum by short tubes expanded into bored holes. Each tube is accessible for cleaning through openings closed by covers with ground joints held in place by wrought-iron clamps and bolts. The tubes are inclined at an angle of about 22 degrees with the horizontal. The rear headers are connected at the bottom to a cast-iron mud drum. The steam drum is horizontal and the headers are arranged either ver- tically or at right angles to the tubes. The boiler is supported by steel girders resting on suitable columns independent of the brick setting. The grate is placed under the higher ends of the tubes, the products of combustion passing at right angles to the tubes and being deflected back and forth by fire-tile baffles. The feed water enters the front of the steam drum as shown in Fig. 36. A rapid circulation is effected by the difference in density between the solid column of water in the rear header and the mixed steam and water in the front one. B. & W. boilers under 150 horse power have but one steam drum, and the larger sizes have two. The number of tubes varies with the size of boiler, ranging from 6 in width and 9 in height in the 100-horse-power boilers to 14 high and 18 wide in the 500-horse-power boilers. 66. Heine Boiler. — Fig. 37 shows a longitudinal section through a Heine horizontal water-tube boiler. This boiler differs from the B. & W. boiler in that the tubes are expanded into a single large header con- structed of boiler steel. The drum and tubes are parallel with each other and inclined about 22 degrees with the horizontal. The feed water enters at the front of the steam drum and flows into the mud drum, from which it passes to the rear header. Steam is taken from the front of the steam drum and is partially freed from moisture by the dry pipe A. A baffle over the front header prevents an excess of water from being carried into the dry pipe. As the rear header forms one large chamber, no additional mud drum is necessary and the sediment is blown off from the bottom by the blow-off cock. The circulation is somewhat freer than in the B. & W. boiler on account of the large sectional area through the headers. 67. Wickes Boiler. — Fig. 38 shows a section through a Wickes vertical boiler, illustrating the vertical water-tube type. The steam drum and water drum are arranged one directly above the other. The tubes are expanded and rolled into both tube sheets and are divided into two sections by fire-brick tile. The water line in the steam drum is carried about two feet above the tube sheet, leaving a space of five feet between water line and top of the drum. This affords a large steam space and disengagement surface. Feed water is introduced into the steam drum below the water line and flows downward through the tubes of the second compartment. The boiler is supported by four BOILERS 84 STEAM POWER PLANT ENGINEERING brackets riveted to the shell of the bottom drum and is independent of the setting. The entire boiler is enclosed in brickwork and is com- pletely surrounded by the products of combustion. The upper part llil 1 i i Fig. 38. Wickes Vertical Water Tube Boiler. of the steam drum acts as a superheating surface and tends to dry the steam. Wickes boilers are simple in design, easy to inspect and clean, low in first cost, and comparable in efficiency with any water- tube type of boiler. BOILERS 85 86 STEAM POWER PLANT ENGINEERING a 5.9 67a. Parker Boiler. — Fig. 38a shows a longitudinal sectional ele- vation and an end sectional elevation of a 1200-H.P. Parker Down- Flow Boiler with double-ended setting. This type of boiler is finding much favor with engineers for central stations where large units are desired. The Parker boiler differs from the conventional horizontal water-tube boiler principally in circulation and flexibility. Feed water is pumped into the economizer or feed element (1), Fig. 38a, at 0, 0, and flows downward through a series of tubes, discharging finally into the drum through an upcast H. In a large unit, as illus- trated here, there are two feed elements and two drums. The circula- tion in the feed element is indicated by solid lines and arrow points at the left of the end sectional elevation, the tubes having been omit- ted from the drawing for the sake of clearness. The intermediate ele- ments (2) take their water supply from the bottom of the drum through a cross-box V, the circulation being downward, as indicated by arrow points, through four tube wide elements, and finally discharge it through an upcast X into the steam space of the drum. Each element has a " down-comer " and an upcast. In the smaller sized boilers the inter- mediate elements are omitted. The evaporator elements (3) take their water supply from the bottom of the drum at V, the circulation being downwards through two tube wide elements, and finally discharge it into the drum at U. The last two passes of the water are through the two bottom tubes of each ele- ment, thus assuring dry steam without the use of dry pipes. To prevent reversal of flow each element is fitted with a check valve at the admission end. Each drum is equipped with a diaphragm, as indicated, separating the steam and water spaces, thus insuring against foaming and priming. Saturated steam is taken from the drum at A and passes by way of B to C, where it enters the superheater S. The superheated steam leaves the superheater at D and passes by way of E and R to the storage drum N, finally leaving the boiler at G. The superheater is designed 1 1 1 BOILER R SIZE? OF OOM AREA FOR VARIOUS TYPES AND BOILERS, 2 BOILERS IN A BATTERY 1 1 1 B 4 W- 3,5,7,8,9,10,11,12,13,11,15,16 Sterling-2,4,6 Helne-1 Parker-20,21 Hormaby-18,19 1 5 < o 11 5 4 I x 15 6^ 3 R 12 13 in t 500 1000 1500 Capacity of Battery Boiler H.P. Fig. 38b. 2000 BOILERS 87 to maintain an approximately constant degree of superheat for all variations in load. All tubes are connected by malleable-iron junction boxes, the interior of each tube being accessible through hand holes placed opposite the Fig. 39. Stirling Boiler and Setting. end of each tube. The hand-hole cover plates are on the inside of the box and have conical ground joints, thus dispensing with gaskets. The Parker boiler is built single or double ended, with or without superheater, and in sizes ranging from 50 H.P to 2500 H.P. standard rating. 68. Stirling Boiler. — Fig. 39 shows a longitudinal section through a Stirling water-tube boiler, which differs considerably from the types just described. Three horizontal steam drums and one horizontal mud drum are connected by a series of inclined tubes. The tubes are bent at the ends to permit them to enter the drums radially. Short tubes 88 STEAM POWER PLANT ENGINEERING connect the steam spaces of all the upper drums and also the water spaces of the front and middle drums. Suitably disposed fire-tile baffles between the banks of tubes direct the gases in their proper course. The boiler is supported on a structural steel framework in- dependent of the setting. The feed water enters the rear upper drum, which is the cooler part of the boiler, and flows to the bottom or mud drum, where it is heated to such an extent that many of the impurities are precipitated. There is a rapid circulation up the front bank of tubes to the front drum, across to the middle drum, and thence down the middle bank of tubes to the mud drum. The interior of the drums is accessible for cleaning by manholes located in the ends. The Stirling furnace is distinctive in design. A fire-brick arch is sprung over the grates immediately in front of the first bank of tubes. The large tri- angular space between boiler front, tubes, and mud drum forms the combustion chamber. Stirling boilers are somewhat lower in first cost than other types of water-tube boilers on account of the absence of numerous hand holes and the like which are necessary in the hori- zontal type. 69. Unit of Evaporation. — The performance of a boiler and furnace may be expressed in terms of the weight of water evaporated per hour per square foot of heating surface or of the weight evaporated per pound of fuel. To reduce all performances to an equal basis so as to facilitate comparison the evaporation under actual conditions is conveniently referred to the equivalent evaporation from a feed-water temperature of 212 degrees F. to steam at atmospheric pressure. The heat required to evaporate one pound of feed water at a temperature of 212 degrees F. into steam of the same temperature, or " from and at 212 degrees " as it is commonly called, is 965.7 B.T.U.* The ratio of the heat neces- sary to evaporate one pound of water under actual conditions of feed temperature and steam pressure to the heat required to evaporate one pound from and at 212 degrees is called the factor of evaporation. Thus for dry steam, X _ t + 32 F = 965.7 ' (U) in which F = factor of evaporation. X = total heat of one pound of steam at observed pressure. t = temperature of the feed water, degrees F. * Recent redeterminations of the properties of saturated steam give this figure as 970.4. BOILERS 89 Dry Surfaces Heatfrom Fuel j Bed and Hot > Furnace Walls \ If the steam is wet, X - xr + q, (12) in which q = heat in liquid at observed pressure. x = the quality of the steam. r = latent heat of evaporation at observed pressure. If the steam is superheated, X = r + q+ Ct s , (13) in which C = the specific heat of the superheated steam. t s = the degree of superheat, degrees F. 69a. Heat Transmission. — Fig. 39a shows a section through a boiler heating plate and serves to illustrate the accepted theory of heat trans- mission. The outer surface of the plate is covered with a thin layer of soot and a film of gas, and the inner surface is similarly protected by a layer of scale and a film of steam and water. It is therefore reasonable to assume that the dry surface of the plate is located somewhere within the film of gas, and the wet surface within the film of water and steam. The heat is imparted to the dry surface by (1) radiation from the hot fuel bed and furnace walls, and by (2) convection from the moving furnace gases. The heat is transferred through the boiler plate and its coatings purely by conduction. The final transfer from the wet surface to the boiler is mainly by convection. Radiation depends on the temperature, and according to the law of Stephen and Boltzmann is approximately proportional to the difference between the fourth power of the absolute temperature of the fuel bed and furnace walls and the temperature of the dry surface of the heating plate. According to this law the heat transmitted by radiation increases rapidly with the increase in furnace temperature. In the modern boiler the surface exposed to radiation is only a small portion of the total o- 1 Fig. 39a. A =• Average Temperature of.Moving Gases. B= Average Temperature of Dry Surface. C —Average Temperature of Wet Surface . D ^Temperature of Water in Boiler. Heat Transmission through Boiler Plate. 90 STEAM POWER PLANT ENGINEERING heating surface, and, since in well-operated furnaces the temperature of the furnace cannot be increased materially on account of practical con- siderations, there is little hope of increasing the capacity of a boiler by increasing the furnace temperature. The heat imparted to a boiler plate by convection may be determined by the following equation (Prof. Perry, " The Steam Engine," 1906 Ed., p. 588): H = C & - t 2 ) vol, (13a) in which ^ H = B.T.U. transferred per hour per sq. ft. of heating surface. C = a coefficient determined by experiment. t x = temperature of the moving gases, degrees F. t 2 = temperature of the dry plate surface, degrees F. v = velocity of the gases, feet per sec. d = density of the gases, lbs. per cubic foot. Prof. Nicholson gives the following modifications of formula (13a) as applied to boiler tubes or flues (Engr. Lond., Feb. 19, 1908): in which H -\m + wt{ l + i)V k - Qvd ' (13b) t = mean film temperature. m = hydraulic radius = area of tube in square inches -*- perimeter of the tube in inches, other notations as in (13a). Both equations are based upon the same general law except that the latter gives a means of determining coefficient C in terms of the mean film temperature and the dimensions of the flues or tubes. An examination of equation (13a) shows that for a given set of condi- tions the heat imparted to a unit of dry surface of heating plate varies directly as the difference between the temperature of the hot gases and that of the dry surface and directly as the velocity and density of the gases. However, the density of the gases drops with the rise of tem- perature, and increase in furnace temperature does not necessarily imply increase in heat impartation. It is the utilization of the velocity factor, then, which offers a possibility of increasing boiler capacity and efficiency. Experiments by Prof. Nicholson and the U. S. Geological Survey show that by establishing a powerful scrubbing action between the gases and the boiler plate the protecting film of gas is torn off as rapidly as it is formed and new portions of the hot gases are brought into contact with the plate, thereby greatly increasing the rate of heat transmission. Similarly the faster the circulation of the water the greater will be the BOILERS 91 scrubbing action tending to remove the bubbles of steam from the wet surface and the more rapid will be the transfer from the plate to the boiler water. The resistance of the metal itself is so small that it may be neglected in calculating the heat trans- mission, and it may be logically assumed that the plate will take care of all the heat that reaches its dry surface. Prof. Nicholson found that by filling up the flue of a Cornish boiler with an internal water vessel, leaving an annular space of only 1 inch around the latter, an evaporation eight times the ordinary rate was effected at a flow of gases 330 feet per second (8 to 10 times the average flow). The fan for creating the draft consumed about \\% of the total power. The conclusion is that the heating surface for a given evap- oration at the present rating may be reduced as much as 90% for the same output, with a cor- responding reduction in the size, cost and space requirements, or with a given heating surface of standard rating the output may be enormously increased; also the increase in power necessary to create the draft is by no means comparable with the ad- vantages gained. The modern locomotive boiler is the nearest approach to these conditions in practice. Here a powerful draft forces the heated gases through small tubes at a very high velocity and an enormous evapo- ration is effected with a comparatively small heating surface. See Fig. 39b for influence of draft on the capacity of a torpedo boat boiler (Power and Engr., May 24, 1910). JUJ ■ t y i -3 19 ■— t- iz r - 3-1 11 - 1-- 1 ~ 2 J & .-• - t O Si Q ^ ~£ A -S^ -,A h£ n *T z? 05 8 -i - ^-rr ° t ■ o!2 jl •2 S -n _- a 7 Ci ^ - o gwfl : „<."■ PhH , -^ n *- : 100 200 300 Per cent, of Rated Capacity Developed by Boiler. Power Fig. 39b. Influence of Draft on the Capacity of a Normand Water-Tube Boiler on the U. S. Torpedo Boat " Biddle." 92 STEAM POWER PLANT ENGINEERING These principles have been applied to a limited extent to stationary boilers already installed by making the gas passages smaller as compared to the length by means of suitable baffles (Fig. 38a) and by forcing larger weights of gas through the boiler, either by forced draft or by increasing the grate area (Fig. 68a). In a general sense when the capacity of a boiler is doubled or tripled the over-all efficiency of the whole steam-generating apparatus drops, but the advantage gained usually offsets the loss in fuel economy. A close examination of the results, however, will show that the loss in efficiency is due more to low furnace efficiency than to inability of the boiler to absorb the heat generated. In view of recent experiments it is not unlikely that within the next ten years boilers will be constructed capable of developing a boiler horse power with two or three square feet of heating surface instead of ten square feet, as at present, and with high over-all efficiency. (See Figs. 41a and 41b.) Heat Transmission in Boilers, Kreisinger and Ray: Power and Engr., June 29, 1909, p. 1144; U. S. Geological Survey, Bulletin Journ. West Soc. Engrs., Sept. 18, 1907; Am. Inst. Elect. Engrs., Dec. 13, 1907. Heat Transfer and Future Boiler Practice : A. H. Allen, Power and Engr., Sept. 21, 1909, p. 482; Engng., Lond, Feb. 19, 1908. The Heat of Fuels and Furnace Efficiency : W. D. Ennis, Power and Engr., r July 14, 1908, p. 50. A Study in Heat Transmission (The Transmission of Heat to Water in Tubes as Affected by the Velocity of the Water), J. K. Clement and C. M. Garland, Univ. of 111. Bulletin No. 40, Sept. 27, 1909. 70. Heating Surface. — All parts of the boiler shell, flues, or tubes which are covered by water and exposed to hot gases constitute the heating surface. Any surface having steam on one side and exposed to hot gases on the other is superheating surface. According to the recommendations of the American Society of Mechanical Engineers, the side next to the gases is to be used in measuring the extent of the heating surface. Thus measurements are made of the inside area of fire tubes and the outside area of water tubes. The heating surface in a boiler under average conditions of good practice is most efficient when the heated gases leave the uptake at a temperature of 100 to 200 degrees F. above that of the steam. Each square foot of heating surface is capable of transmitting a certain amount of heat, depending upon the conductivity of the material, the character of the surface, the temperature difference between the gas and the water, the location and arrangement of the tubes, the density of the gas, the velocity of the gas, and the time allowed for transmission of the heat. It is customary to assume a uniform heat BOILERS 93 transmission for the entire surface. Thus with most boilers it is found that the best results are obtained with an evaporation of from 3 to 3.5 pounds of water from and at 212 degrees F. per square foot of heating surface, which is equivalent to allowing 12 to 10 square feet per boiler horse power. By increasing the quantity of heat the evaporation may be increased, but at the expense of efficiency, since a smaller percentage of the heat is utilized. For example, an evaporation as high as 20 pounds per square foot per hour has been effected in torpedo-boat prac- tice, and 12 pounds per square foot per hour is not unusual in locomotive work, but such performances are invariably obtained at the expense of economy. The selection of the proper proportion of heating surface to the evaporation required is evidently a very important matter. For maximum economy under average conditions of operation, practice allows a proportion of 1 square foot to every 3.5 pounds of water to be evaporated from and at 212 degrees F. Where economy must be sacri- ficed to capacity, as in locomotive practice, a much higher evaporation is allowed. The maximum evaporation is limited by the amount of coal which can be burned upon the grate. It the draft is sufficient, a good boiler can develop a horse power upon 0.75 to 0.5 of the surface recommended. In the very latest large central stations the gas passages and grate surface are proportioned so that the boiler may be operated at 100% above standard rating with high over-all efficiency. The following table shows approximately the result which may be expected with different rates of evaporation. POUNDS WATER EVAPORATED FROM AND AT 212 DEGREES F. PER SQUARE FOOT OF HEATING SURFACE PER HOUR. 2 2.5 3 3.5 4 5 6 8 10 12 PROBABLE RELATIVE ECONOMY. 100 100 100 100 99 98 95 90 85 80 Efficiency of Boiler Heating Surface: Trans. A.S.M.E., 18-328, 19-571. Kent, Steam Boiler Economy (John Wiley & Son), Chapter IX. The Nature of True Boiler Effi- ciency: Jour. West. Soc. Engrs., Sept. 18, 1907. Heat Transference through Heating Surface: Engineering, 77-1. 71. The Horse Power of a Boiler. — A boiler horse power is equivalent to the evaporation of 34.5 pounds of water per hour from a temperature of 212 degrees F. to steam at atmospheric pressure. This corresponds 94 STEAM POWER PLANT ENGINEERING to 33,305 B.T.U. per hour.* Since the power from steam is developed in the engine and the boiler itself does no work, the above measure of capacity is merely conventional. Thus one boiler horse power will furnish sufficient steam to develop about three actual horse power in the best compound condensing engine, but only one-half horse power in a small non-condensing engine. Boilers should be purchased on the basis of heating surface and not on the horse power rating, since one bidder may offer a boiler with say 5 square feet of heating surface per horse power and another with 10 square feet, both being capable of the required evaporation, but the one with a small heating surface (which will, of course, be the cheaper boiler) will do so only at an increased cost of fuel. Manufacturers ordinarily rate their boilers on the basis of 10 to 12 square feet of heating surface per horse power, and the power assigned is called the builder's rating. As this practice is not uniform, bids and contracts should always specify the amount of heating surface to be furnished. According to the recommendations of the American Society of Mechanical Engineers, " A boiler rated at any stated capacity should develop that capacity when using the best coal ordinarily sold in the market where the boiler is located, when fired by an ordinary fireman, without forcing the fires, while exhibiting good economy. And further, the boiler should develop at least one-third more than stated capacity when using the same fuel and operated by the same fireman, the full draft being employed and the fires being crowded; the available draft at the damper, unless otherwise under- stood, being not less than one-half inch water column. In determining the boiler horse power required for a given engine horse power it is convenient to estimate the steam consumption of the engine under actual conditions and then ascertain the equivalent evaporation from and at 212 degrees F. For example, assume a single non-condensing engine developing 20 horse power to use 50 pounds of steam per horse power hour, or 1000 pounds steam per hour; steam pres- sure, 80 pounds per square inch; feed-water temperature, 120 degrees F. Required the boiler horse power necessary to furnish this quantity of steam. From equation (11), the factor of evaporation is -, X-t + 32 1185.3-120 + 32 . 1Q1 F = 970.4 = ~9704 = U3L One thousand pounds of steam under the given conditions are there- * With the new value of r = 970.4 in place of 965.7 this figure becomes 33,478.8. BOILERS 95 fore equivalent to 1000 X 1.131 = 1131 pounds from and at 212 degrees F. The boiler horse power necessary to furnish steam for the 20-horse- power engine will be 1131 Boiler horse power = ^—^- = 32.8. Example : A 15,000 kilowatt steam turbine and auxiliaries require 14.7 pounds of steam per kilowatt-hour at rated load; steam pressure 200 pounds per square inch gauge; superheat 150 degrees F.; feed-water temperature, 179 degrees F. Required the boiler horse power necessary to furnish this quantity of steam. The heat furnished to the turbine and auxiliaries per kilowatt-hour is w \ X + Cpt s - (t - 32) } = 14.7 { 1199.2 + 0.57 X 150 - (179 - 32) } = 16,724 B.T.U. t> -i u 15 > 000 X 16 > 724 T^nn , \ Boiler horse power = 7Q = 7500 (approx.). Table 12 gives the required hourly evaporation per boiler horse power at various feed temperatures and steam pressures. The following table shows approximately the relation between boiler horse power and heating surface for different ratios of evaporation: EVAPORATION FROM AND AT 212 DEGREES F. PER SQUARE FOOT PER HOUR. 2 2.5 3.0 3.5 4 5 6 7 8 9 10 SQUARE FEET HEATING SURFACE REQUIRED PER HORSE POWER. 17.3 13.8 11.5 9.8 8.6 6.8 5.8 >.9 4.3 3.8 3.5 Builders of return tubular and vertical fire-tube boilers allow 11 to 12 square feet of heating surface per horse power; water-tube boilers are rated at 10 square feet per horse power, and Scotch-marine boilers at 8 square feet per horse power. 72. Grate Surface. — The amount of fuel which can be burned per hour limits the amount of water evaporated per unit of time and depends 96 STEAM POWER PLANT ENGINEERING c3 O O w -a c 3 1 A s 0) o o IN I*- CM CO i-l co CM 00 lO CM O •«* CO OS CM U3 OO OO t^ t^- OO NOCO<0 0> O) OO CN CO 00 00 00 00 05 (NNMfNN 05 05 OS O O CM CN CN CO CO o ^ ^ _ _ CO CO CO CO CO CM CM CO CO o C5 omffj^ai CN ■* CO O) i-H lO i-H OO lO CO ■* t^ OS CN »0 HHOHN 00-hM*NO CO CM CON 00 00 00 00 05 05 o o o CM CN CO CO CO OHHHN CO CO CO CO CO CM CM CO CO o CONNNM CO iO 00 O CO OS CO CO i-H OS •OOOrt^tO t~- t— CO t- OO OS CN >0 00 H O 05 »o OO 00 00 00 05 05 CM CM CM CM CM ©oooo CN CN CO CO CO O l-H l-H H CM CO CO CO CO CO CM CM CO CO o NHtDHN CO CO 00 i-H CO noNioco C0 05^^1> CN i-h ^h CN CO O CO CO 05 CM lO ** »0 OS CO 00 00 05 05 CM CM CM CN CN 00)000 CM CM CO OO CO rtrtrHrHN CO CO CO CO CO CM CM CO CO o CO ^h iO o *o i— I ^COOSiHTji t-» ^ iH 05 OO CO 05 CM •* t^ t^ CO CO t- OO O CO CO OS CM O 05 CO O) 00 00 00 05 05 CM CM CM CM CM ooooo CN CN CO CO CO rtrHrtlHIN CO CO CO CO CO CM CO CO CO o CN lOOJ^OSiO ■*C0 05i-it}* CN 05 CO ^ CO t^ OS CM »0 00 CN i-H t-h CN CO rHTJ- l-H 00 00 00 05 05 CM CM CM CM CM 05 o o o o CN CO CO CO CO l-H l-H ^H CM CM CO CO CO CO CO CM CO CO CO o o "*• 05 ^f 05 ^O lONOcqm CN OS CO lO CO OC O CO CO 05 CM CN CO -^ lO (NiOWiH'* l>- i-h CO00O5O5O5 CM CN CN CM CM 050000 CN CO CO CO CO i-H t-< t-h CM CM CO CO CO CO CO CN CO CO CO o OS 05 ^ OS ■"*' i— 1 lO OO O CO CO N^INOO) OO i-H •* t^ 05 OO OO O) O CN CN lO OO CM lO OO CN 00 00 05 05 05 CM CM CM CM CM 05 o o o o CN CO CO CO CO l-H ^H r-t CM CM CO CO CO CO CO CM CO CO CO o 00 ^OliOON (OOOiH-*© CO O 00 CO lO 05 CN "tf t^ O lO lO CO t^ O) CO CO O) CM lO 05 CO 00 00 05 05 05 CM CM CM CM CM 05 O O O i-H CM CO CO CO CO ^H i-H i-H CN CM CO CO CO CO CO. CM CO CO CO o O "0 i— I CO CO r- OS cm ■* t^. O t* »0 CO CM O CM lO OO i-h CN CN CO -^ CO '05 05 05 CO 00 00 05 05 05 CM CM CM CM CM O O O O i-H CO CO CO CO CO ^h i-h CN CN CM CO CO CO CO CO CN CO CO CO o CO NNNCOO NOMiOOO O CO CO 05 CM O O i-H CO lO lO OO r-i Tt< t- OO OO O •"*• OO 05 OS OS 05 CM CM CM CM CM O O O O i-i CO CO CO CO CO ^H i-h CM CM CN CO CO CO CO CO CO CO CO CO o »o •* 05 lO i-H 00 OO O CO CO OO tO CO i-H 05 05 i-H ^ t^ OS CM 00 O) O CN •* lO 00 CN lO OO t— OO l-H lO 00 05 05 05 05 CM CM CM CM CM O O O O i-h co co co co co i-H i-H CN CN CN CO CO CO CO CO CO CO CO CO o (N0C*ON 05 i— 1 ^ I— 05 hjinoooo CN lO OO O CO 00 O) i-H CO »0 CO O) CO CO OS CO O) CN CO O0 05 05 05 05 CN CM CM CM CM O O O t-H i-H CO CO CO CO CO i-H i-H CN CN CN CO CO CO CO CO CO CO CO CO o CO (NOO^ON O CM *0 00 O -^ CO i-H o o CO CO 05 CM »0 O i-H CO lO OO oo i-h -^ r~ o ^H O Tt< 00 OS 05 05 05 O CN CN CN CN CO O O O i-H rH CO CO CO CO CO i-H CN CN CN CO CO CO CO CO CO CO CN CO CO o - •* I-H 05 "ONOMiO MNCOON OO y-t "tf r>- o O)-HC0NH MNOMN CO OS o ■* OlfflOOO CN CN CO CO CO OhhhN co co co co co CN CN CO CO CO CO CO CO CO CO TJ4 -^ CO CO -aaduia emre I V™d O O O O O »OCDN 00O5 o o o o o O i-h CN CO ^ o o o o o lO©N00O) O CN O -H CM CM BOILERS 97 upon the extent and nature of the grate surface, the character of the fuel, and the draft. A liberal allowance of grate surface is usually desirable, particularly when the boilers are to be forced, since too small a grate increases the labor of handling and cleaning fires and results in poor economy. With good coal low in ash approximately equal results may be obtained with large grate surface and light draft and with small grate and strong draft, the amount of coal burned per hour being the same in both cases. Bituminous coal low in ash gives best results with high rates of combustion, provided the ratio of grate surface to heating surface is properly proportioned. Coals high in ash require a compara- tively large grate surface, particularly if the ash is easily fusible, tending to choke the grate. Where a strong draft is available a smaller grate may be used than with moderate draft, as a thicker bed of fuel can be carried. The relation between draft and rate of combustion for various sizes of coals is shown in Fig. 116, paragraph 127. A number of boiler tests made by Barrus (" Boiler Tests ") showed that the best economy with anthracite coal hand fired was obtained with an average ratio of grate surface to heating surface of 1 to 36 and at a rate of combustion of approximately 12 pounds of coal per square foot of grate surface per hour. In these tests a variation in grate and heating surface ratio of 1 to 36 up to 1 to 46 gave practically no difference in economy. With bituminous coal the tests showed that an average ratio of 1 to 45 gave the best results and at a rate of combustion of 24 pounds of coal per square foot of grate surface per hour. Tests made by Christie (Trans. A.S.M.E., 19-330) gave an average combustion of 13 pounds of anthracite per square foot of grate per hour for maximum efficiency and 24 pounds of bituminous. Current central station practice gives normal rates of combustion approximately as follows (lbs. per sq. ft. per hr.) : Anthracite 15-20 Eastern bituminous 20-24 Semi-bituminous 18-22 Western bituminous 30-35 Table 13 gives the relation between heating and grate surface in a number of recent boiler installations using different kinds of coal, and is illustrative of current practice. In proportioning the grate surface for a proposed installation the principal factor considered is the character of the fuel, a study being made of the various fuels available, and the one selected which gives the highest evaporation per dollar. The latter data may usually be obtained from records of plants using the same grade of fuel and grates similar to those intended for the proposed plant. 98 STEAM POWER PLANT ENGINEERING TABLE 13. RATIO OF HEATING SURFACE TO GRATE SURFACE IN RECENT BOILER INSTALLATIONS. Nature of Plants. Central stations Do Do Do Manufacturing plants Office buildings Central station* No. of Plants. 10 9 20 6 1 Type of Boiler. Hor water tube. ...do ...do ...do Return tubular. ..do Babcock & Wilcox. Type of Grate. Chain Roney Murphy. . . Miscel's . . . Hand fired Shaking grates. Roney Height of Chimney. 200 feet and over. ...do ...do...... ...do 150-175 Over 200 . . Over 200 . . Character of Fuel. 111. screen- ings, 15 to 20% ash. Bituminous. do. Anthracite. Anthracite. Bituminous Bituminous. Ratio of Heating to Grate Surface. 65 60 60 40 35 48 31 * Two stokers, one at front and one at rear of setting. {Power, Jan. 7, 1908, p. 25.) 73. Boiler and Furnace Efficiency. — The efficiency of the boiler, including the grate, is expressed by the ratio between the heat absorbed by the boiler per pound of dry coal fired and the calorific value of one pound of dry coal. The efficiency of the boiler alone is taken as the ratio between the heat absorbed per pound of combustible burned on the grate and the calorific value of one pound of combustible. The combustible burned on the grate is equal to the coal as fired minus moisture and the total refuse in the ash pit. The calculation of these efficiencies is illustrated by the following example: ANALYSIS OF COAL. Per Cent. Moisture 8 Ash 12 Combustible 80 100 Pounds. Water evaporated from and at 212° F. per pound of coal as fired . . . 8.281 Per Cent. Total refuse in ash pit 16 Percentage of ash in refuse 13 Combustible in ash 3 B.T.U. Heating value per pound of coal as fired 11,680 Heating value per pound of dry coal = 11,680 -*- 0.92 12,696 Heating value per pound of combustible = 11,680 -r- 0.80 14,600 8.281 + 0.92 = 9.001 = equivalent evaporation per pound of dry coal. 9.001 X 965.7 = 8,692 = heat absorbed per pound of dry coal. BOILERS 99 Efficiency of boiler and grate = .. ' ana = 68.49 per cent. Combustible burned on grate = 100 — (8 + 16) = 76 per cent. 8.281 -*- 0.76 = 10.896 = equivalent evaporation per pound of com- bustible burned on the grate. 10.896 X 965.7 = 10,522 = heat absorbed per pound of combustible. ■I f\ coo Efficiency of boiler = ' = 72.07 per cent. The efficiency of the grate alone might be expressed _„. . . Efficiency of boiler and grate Efficiency of grate = tW • ti — n — 5 > J Efficiency of boiler which is equivalent to .^~, . , , Combustible actually burned Efficiency of grate Combustible fired the numerator being the coal fired less moisture and the refuse from the ash pit, and the denominator the coal fired less moisture and the ash as determined from the proximate analysis. The efficiency of combustion is sometimes expressed in terms of the difference in temperature between fuel bed and flue gas: Tf — T Efficiency of furnace = ~ j~ » (14) If — la in which Tf = temperature of the furnace. T c = temperature of the flue gas. T a = temperature of the air. The efficiency of the furnace or combustion may also be stated (R. S. Hale, Trans. A.S.M.E., 20-769): S + F Efficiency of furnace = — = — > (15) in which S = B.T.U. absorbed by the boiler per pound of dry coal. F = B.T.U. lost in flue gases per pound of dry coal. H = Calorific value of 1 pound of dry coal. The heat absorbed by the boiler expressed in percentage of the heat available has been given the name true boiler efficiency by the U. S. Geological Survey and may be expressed rp rp True boiler efficiency = -=£ — ^> (15a) J- f J- s in which T s = temperature of the steam (saturated); other notation as in (14). 74. Boiler Performances. — Table 14 is compiled from a number of tests of different types of boilers with various types of grates and 100 STEAM POWER PLANT ENGINEERING jajioa: jo jfouapryg; •eiqiisnquuoo jo punoj J9d 1101^ j -od'BAg ^u^BAinba ifl WNNNO ■N050iNOONO)«00 00 N © CO -h ■<* © 00 O CO © ^H i-H O lO lO O N CO CO t-i o I> O o a osm a t> CO 00 t> CO ■* o o> co oo NHHiONHfflLOOSMNNN ^NiONlMiOCOWMiO »0 lO COHCCMM^(N'*HT|l(ONCq000000« coioiot>coo>-ii>i>cococoa)i>ooooco i-Hi-l H N rHi— ItHi-Ii-( H lO 1C ■* ^ ^(OlOCOH^COtO i> © © O O © £ _ PQ < £ O > o. c 3 tuD — ' O .2 2 o be C 3 3 O B !* o « 2 £3 S S a> 3 a o O .3 3 c3 •J9M0J 8SJ0H pai'B'jj jo sS-eiuaoisa; OJOOtOMOfiiOOO^ON H CO rHrH r-l tH ,H »-l i~l 5-3 O <3 aoioioioto^o OOOOHOOOOOffl ^H i-H r-4 CO •J9A\ot^OOOOCNCN(MOOOOOOOOOOOOOt^ ©0©iOiOO»Oi-Ht-I'-i-<,_I'-i>-hiOOCD©»0©>Oi-i H H H.H HNN^Tli^^iONNNNCOiO (N CM CO CO lO ooooooooooo QOOQPPflPQQfl sasss^ -°£3 Q Q *0^[ 90U9J9J9H N w CO CO ^ "O t^00©©©©CO©CO^t'-t'-ICMCO BOILERS 101 *(»OOCON>OffiiN N H 00 00 . 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CN CO ft CM p^ _,T ft • "* cs w o ho ^ "-• • cs ^ a a ho ^d _: cs rt •^ j co ^cm" h v cS co ft w S . d r ■* . o M g « : s » «*1** 5 ^ &3 ? CO co • co o 3 S fc S 102 STEAM POWER PLANT ENGINEERING TABLE 14a. PRINCIPAL DATA AND RESULTS OF TESTS ON BOILER NO. 6, UNIT NO. 10, FISK ST. STATION. COMMONWEALTH EDISON CO., CHICAGO. (B. & W. Boiler, " Standard " Setting.) Water-heating Surface, 5000 Sq. Ft. Superheating Surface, 914 Sq. Ft. Chain Grate Surface, 90 Sq. Ft. H.P. per Sq. Ft. Grate. Heat Total Super- Dry Coal Test Date, Horse Eff'y, Lost in Heating heat of per Sq. Ft. No. 1908. Power. PerCent. Refuse, Surface Steam, G. S. PerCent. per H.P. Deg. F. per Hour. 2 Mar. 9 873 67.4 9.70 2.8 6.76 197 41.2 4 " 10 873 69.0 9.52 2.8 6.89 195 39.1 6 " 11 852 67.3 9.47 2.8 6.93 189 38.9 8 " 16 836 65.3 9.29 6.4 7.06 174 39.5 10 a 17 870 68.8 9.67 5.0 6.78 180 39.3 14 " 19 920 66.2 10.22 9.2 6.42 187 43.7 16 " 23 900 69.5 10.00 4.0 6.56 181 40.5 18 " 24 916 69.1 10.18 5.5 6.44 190 41.6 20 " 26 912 69.2 10.13 4.4 6.48 179 41.2 22 " 27 906 67.7 10.07 4.1 6.52 194 42.5 24 " 30 925 69.8 10.28 2.8 6.38 179 41.6 26 " 31 894 69.4 9.93 5.2 6.60 170 40.6 28 Apr. 1 922 71.2 10.24 3.6 6.40 169 40.4 30 " 2 923 71.5 10.26 4.6 6.40 173 40.5 32 u 7 914 70.0 10.20 4.5 6.46 175 40.9 34 " 8 939 73.8 10.4 3.8 6.28 181 40.4 36 " 10 911 70.9 10.1 3.0 6.48 185 40.2 38 " 11 967 70.1 10.7 3.0 6.11 192 42.6 40 " 13 995 67.8 11.1 3.4 5.93 211 43.6 42 " 14 887 66.8 9.9 4.5 6.65 202 40.8 44 " 27 880 69.5 9.8 5.5 6.72 169 39.7 48 " 29 927 71.5 10.3 3.3 6.37 171 40.8 50 " 30 899 70.3 10.0 4.2 6.57 171 39.6 52 May 6 886 69.4 9.8 5.3 6.67 171 38.2 54 " 7 900 69.1 10.0 4.8 6.56 171 39.2 56 " 8 967 71.9 10.7 4.8 6.10 164 40.1 58 " 11 902 70.5 10.0 3.3 6.55 163 39.6 60 " 13 875 70.7 9.7 3.8 6.74 147 38.3 64 " 14 1102 72.0 12.2 4.8 5.35 180 43.2 BOILERS 103 TABLE 14a. PRINCIPAL DATA AND RESULTS OF TESTS ON BOILER NO. 6, UNIT NO. 10, FISK ST. STATION. COMMONWEALTH EDISON CO., CHICAGO. (B. & W. Boiler, "Standard" Setting.) Water-heating Surface, 5000 Sq. Ft. Superheating Surface, 914 Sq. Ft. Chain Grate Surface, 90 Sq. Ft. Draft Heat Lost B.T.U. per Pound Dry Coal. Ash in Dry Coal, Per Cent. Ash in Refuse, Per Cent. Uptake Temp. Deg. F. co 2 , Per Cent. up Stack (Dry Gas), Per Cent. Over Fire. In Uptake. .87 1.34 11,634 18.46 82.33 466 6.9 .78 1.25 11,759 16.81 81.36 461 6.7 .83 1.25 12,039 16.08 80.03 463 7.7 15*6 .94 1.34 11,993 15.91 67.42 477 7.6 16.8 .84 1.24 11,909 15.71 71.32 475 7.9 16.2 .99 1.41 11,768 16.04 63.78 479 8.5 15.4 .77 1.17 11,846 16.68 79.04 483 9.1 14.0 .81 1.25 11,800 16.39 71.98 484 8.3 15.8 .77 1.21 11,846 15.51 78.53 486 9.0 14.5 .78 1.22 11,659 17.59 80.58 494 9.2 14.6 .68 1.28 11,800 16.22 82.97 487 8.8 15.1 .70 1.24 11,752 16.18 76.84 484 8.8 15.1 .62 1.21 11,862 15.38 82.99 480 9.2 14.1 .58 1.40 11,800 16.02 78.37 480 9.1 14.4 .73 1.24 11,815 16.84 77.84 494 9.0 14.7 .72 1.25 11,659 18.06 82.27 504 8.9 15.3 .65 1.13 11,831 17.15 86.92 493 9.7 13.4 .70 1.24 12,002 16.05 84.39 502 9.0 15.1 .71 1.23 12,469 14.87 82.14 522 9.7 13.3. .63 1.09 12,049 15.17 78.12 500 9.5 13.3 .71 1.26 11,801 15.75 77.21 470 8.3 15.7 .68 1.23 11,769 18.59 84.04 472 8.7 14.2 .66 1.27 11,955 16.11 79.30 473 7.9 16.1 .62 1.20 12,360 13.63 74.59 476 8.8 14.5 .66 1.31 12,298 13.62 75.19 480 9.0 14.4 .66 1.29 12,423 13.37 75.61 474 9.4 13.3 .92 1.18 11,956 17.45 83.24 451 9.2 12.5 .76 0.98 11,971 17.45 80.99 443 10.0 11.2 .68 1.15 13,126 10.24 70.90 487 10.4 12.1 104 STEAM POWER PLANT ENGINEERING characters of fuel. Although some of the tests show a combined efficiency of boiler and grate as high as 85 per cent (Engr., Lond., March 21, 1902, p. 286), such a performance cannot be expected for continuous operation under the average conditions of practice. In pumping stations or in plants where there are no peak loads and the boiler may be operated under a practically constant set of conditions a continuous efficiency of 75 per cent has been realized with coal as fuel and 80 per cent with crude oil, though these figures are exceptional. In very large central stations, with the usual loads in the morning and evening, an average efficiency throughout the year of 65 per cent is possible, though a good figure is not far from 60 per cent. In large isolated stations with variable loads good practice gives an average of 60 per cent. Small stations though showing an efficiency as high as 75 per cent at times seldom average 50 per cent for the year. The usual discrepancy between efficiency as determined by special tests and everyday operation is due to the fact that the efficiency test is usually conducted under ideal con- ditions: the boiler surfaces are cleaned, the rate of combustion carefully adjusted for maximum economy, and special attention given to the firing, whereas in actual practice these refinements are seldom attempted. Much depends upon the efficiency of the boiler-room staff, the character of furnace and fuel, draft, and the load factor. From the commercial standpoint the performance is best expressed in terms of the " cost to evaporate 1000 pounds of water from and at 212," or the " pounds of water evaporated per $1 of coal." Table 15 gives the results of a number of tests, made at the Armour Glue Works, Chicago, 111., show- ing the cost of evaporating water with different grades of Illinois coal. The results were obtained from hand-fired Stirling boilers. Boiler Room Economies: Am. Elecn., Oct., 1901, p. 506, Sept., 1905, p. 472; Cas- sier's Mag., March, 1906, p. 373; Elec. World, March 4, 1905; Engr. U.S., May 1, 1905, p. 304, Jan. 1, 1907; Engr., June 4, 1907, p. 758; Eng. Rec, June 27, p. 685; Elecn., Lond., Aug. 5, 1904; Power, Aug., 1905, p. 484; Eng. Mag., Oct., 1901, March, 1903. Care and Management of Boilers : Engr. U.S., March 1, 1902, p. 142, Feb. 15, 1904, July 15, 1904, Jan. 1, 1907; Engineering, Feb. 18, 1898, p. 211, July 15, 1898, p. 84; Eng. Mag., Feb., 1901, p. 877, Oct., 1901, p. 91, March, 1903, p. 896; Power, Sept., 1904, p. 467, May, 1905, p. 267, Dec, 1905, p. 742, Sept., 1906, p. 550; Am. Elecn., Feb., April, Sept., 1904; Mech. Eng., July 25, 1903; Engr., Lond., April 15, 1904; Elec. Rev., July 13, 1907. 75. Effect of Capacity on Efficiency. — In general, as the horse power of a boiler increases above normal capacity the over-all efficiency will decrease, due to the fact that the furnace and gas passages are ordinarily proportioned to effect an evaporation of about 3.5 pounds of water from and at 212 degrees F. per square foot of heating surface per hour at rated load, the temperature of the escaping gases being from BOILERS 105 150 to 200 degrees above that of the steam. To increase the rate of evaporation more coal must be burned per unit of time and consequently a larger volume of gas is generated. The larger the volume of gas the higher will be its velocity, which finally reaches a point where heating surface is insufficient in extent to absorb the extra heat and as a con- sequence the flue gas escapes at a higher temperature, resulting in lower boiler and furnace efficiency. With properly proportioned grate, furnace and gas passages a boiler may be operated at 100% above stand- ard rating with little or no decrease in over-all efficiency. Fig. 40 shows a case in which the efficiency decreased with the increase in capacity, and Fig. 41a illustrates increased efficiency for the higher rates of driving. These curves are of value simply as illustrations of the behavior in specific cases, and are not applicable to all types of boilers. TABLE 15. RESULTS OF COAL TESTS AT ARMOUR GLUE WORKS, CHICAGO, AUG. 17, 1905. Date of Test. March 5, 1905 . . March 3, 1905 . . June 14, 1905 . . June 15, 1905 . . June 16, 1905 . . June 17, 1905 . . June 19, 1905 . . June 20, 1905 . . July 1, 1905. . . . July 6, 1905. . . . July 28, 1905. . . July 29, 1905. . . Aug. 5, 1905 . . . Aug. 7, 1905 . . . Aug. 8, 1905 . . . Aug. 9, 1905 . . . Aug. 11, 1905 .. Name and Kind of Coal. Williamson County Coal Co.'s, mine run Harden & Hafer, mine run Crerar-Clinch & Co., 2" screenings. ...do ...do Brackett Coal and Coke Co., lump. ...do do. Kelly ville Coal Co. mine run. Brackett C. & C. Co. Keeler mine run. Kellyville Coal Co. washed pea. ...do Dering Coal Co., mine run. Dering Coal Co., Sulli- van Co., screenings. Consolidated Indiana Coal Co., Sullivan Co., screenings. Screenings Ziegler, screenings Railroad Car Number. C.C.C.&St.L. No. 26368 S. I. No. 5735 I.C. No. 88362 I.C. No. 88362 I. C. No. 88362 C. &E. I., No. 8891. C. &E. I. No. 5002 C. & E. I. No. 5002 C. & E. I. No. 10030. C. & E. I. No. 12367 C. & E. I. No. 6211. C. &E. I. No. 6211 C. &E. I. No. 25125 E. &T. H. No. 5132. E. &T. H. No. 3239 E. &T. H. No. 6534 I.C. No. 81184 Cost per Ton Deliv- ered. Cost to Evaporate 1000 Pounds of Water. $1.90 $0.1531 1.70 0.1231 1.50 0.1293 1.50 1.50 1.65 0.1218 0.1175 0.122 1.65 0.1212 1.65 0.1352 1.595 0.1355 1.65 0.1236 1.50 0.1285 1.50 0.119 1.575 0.125 1.40 0.11 1.35 0.105 1.30 0.0973 1.50 0.1047 Pounds Water Evapo- rated per $1.00 of Coal. 6,532 8,123 7,734 8,210 8,511 8,197 8,251 7,396 7,380 8,091 v 7,782 8,403 8,000 9,091 9,524 10,277 9,551 * See, "The 1907, p. 677. Nature of True Boiler Efficiency," Jour. Wes. Soc. Engrs., Oct., 106 STEAM POWER PLANT ENGINEERING In nearly all stations the boilers must have sufficient overload capacity to take care of peak loads or to allow some of the boilers to be shut down for cleaning or repairs, since the installation of sufficient 70 a fy) 0> D > a =% ^ 1 let) 1 5 3 W ■o 55 5.5 g> 0) i- -.*«$ b.0gc! c# C^ ' 4 - 5 > | 4.0® 1 a 03 ft 50 3.5.Q- hi Jour.W.S.E.,Fet .1719(4. 0.45 0.5 0.2 0.25 0.3 0.35 0.4 Draft over Fire in Inches of "Water Fig. 40. Influence of Draft on the Efficiency and Capacity of a 350-Horse-power Babcock and Wilcox Boiler with Chain Grate. 12 3 11 10 — k i — , ^ 5\ ( "^X X N K t) 2 3 4 5 6 Lb. Water Evaporated per Sq! Ft. of Heating Surface per Hour Fig. 41. Effect of Rate of driving on Economy of a 150-Horse-power Stirling Boiler, Hand Fired. rated boiler capacity would be expensive and in many instances pro- hibitive in cost. In small stations, however, too large a boiler capacity frequently is to be preferred to an overloaded installation, since the BOILERS 107 75 « • • • • • • • 70 • • • ^^* i • • • • • ••x* • -*• • • a> O S 65 On • • • • • ••• *>* • < • • • • • • >^» • • • • • • a e eo H o • f • •" • • < • • • • a u 3 & 55 ■d a as • / 500 H.P, B. & W._Boiler 5000 Sq. Ft. Water Heating Surface 940 >» " Superheating Standard Chain Grate and Setting 90 Sq. Ft. Grate Surface ■2 o «50 ' various uraaes oi uoai Fisk Street Station, Common Wealth-Edison Co. Chicago 111. 45 •/ • 400 500 700 800 900 Boiler Horse Power 1000 1100 1200 Fig. 41a. Relation Between Efficiency and Capacity, 500 H.P. Boiler, Fisk Street Station, Commonwealth Edison Co., Chicago. ■ail- 8 >.io. c »10.i 3 a 10. 10, § 10 2 9.6 a a H 9.4 ■3 9.2 > I 9 600 H.P B.& W. BOILER h 6006 . -■-:-' - - : EQUIPPED WITH RONEY STOKERS AT THE 59th ST STATION OF THE INTERBOROUGH RAPID TRANSIT CO., N.Y. \ • 1 1 • \. 1 sir IGLE STO 1 » 60 1 / » * — x \ « "0 Sffi ;ier cy p ' — ^^_c ■s /( 1 c ) / / 30 Thickness of Fire, Inches Fig. 42. Effect of thickness of Fire on the Capacity and Efficiency of a 350-Horse-power Stirling Boiler, equipped with Chain Grate. power Stirling boiler equipped with chain grate at the power plant of the Armour Institute of Technology. The damper was left wide open throughout the test and the speed of the grate kept constant. Ratio of grate to heating surface, 1 to 42. Carterville washed coal No. 4 was used in all tests. The curves in Fig. 43 refer to the performance of a 150-horse-power water-tube boiler equipped with chain grate at the University of Illinois Engineering Experiment Station at Urbana. 110 STEAM POWER PLANT ENGINEERING The curves in Fig. 44 are plotted from a series of tests on a 500-horse- power Babcock & Wilcox boiler equipped with chain grate at the Fisk Street station of the Commonwealth Edison Company, Chicago, 111. In these tests the conditions of operation are not exactly comparable, but they serve to show the variation of economy with thickness of fire in each case. In general, with natural draft, fine sizes of coal necessi- tate thin fires, since they pack so closely as to greatly restrict the draft. Thin fires require closer attention to prevent holes being burned in Horse- Power Developed (34 Vz Lbs. of Water Evap. into Dry Steam F. & A. 212° per Hr.= l H.P.) o o o o o o o c L^ i/ / ^ Rate d Caps tcity o: * Boile r X / X I *P ^> y^o • •^ <^ f& o ^ m 7.5 c . -1 H 3 ° 6.5 a -all g g pn 6.0 §*& 8 M ** ft W 5.5 n o^^ • O ««&izi. 2£e__ o c ) « » • • 6-in. Fire • T 15 20 25 30 35 Dry Coal per Sq. Ft. of Grate Surface per Hr.-Lbs. 40 Fig. 43. Effect of Thickness of Fire on the Capacity and Efficiency of a 150-Horse-power Water-Tube Boiler. spots, and respond less readily to sudden demands for steam, but have the advantage of letting the air required pass through the grate, whereas thick fires often require air to be supplied above the grate to insure complete combustion. Thick fires require less attention and hence are preferred by firemen. Where sufficient draft is available thick fires are more efficient than thin ones, as the air excess is more readily controlled. BOILERS 111 77. Influence of Initial Temperature on Efficiency. — In general the higher the initial temperature of the furnace the greater will be the efficiency of the heating surface, since the heat transmitted varies almost directly with the difference of temperature between the water and the products of combustion. If the heating surface is properly distributed so that the final temperature of the escaping gas remains constant, the efficiency of the boiler and furnace will increase as the initial temperature increases, though not in direct proportion. This is on the assumption 1000 ». 800 400 200 00 40 ' ' ^ ^ / * ^> ■». >> «. 4 B< >— _i ?_ — « j A Boiler 14 Tubes High B " 9 " J )U VV S. fcj. 1.1 H cases the admission of air above ^ the grate through openings in the bridge wall or passages in the side walls frequently gives satisfactory results. When natural draft is not sufficient, as is usually the case under heavy load, steam jets or forced draft may be employed. For a description of such devices see paragraphs 148 to 150. 101. Cost of Stokers. — The following is the approximate cost of stokers suitable for a Babcock & Wilcox boiler of 350 horse power rated capacity with 45 square feet of grate surface ; height of chimney above grate, 1 75 feet ; coal burned, Illinois screenings. The cost of installation is not included. 1. Chain grate and appurtenances $1,500.00 2. Jones underfeed stoker 1,400.00 3. Hawley down-draft furnace 1,350.00 4. Burke smokeless furnace 1,000.00 5. Roney stoker 1,300.00 6. Murphy furnace and stoker 1,350.00 7. Wilkinson stoker 1,200.00 Parsons' Smokeless Furnace. See par. 149. Heinrich Smokeless Furnace. See par. 150. Steam jets. See par. 148. Hamler-Eddy Smoke Recorder. See par. 411a. Ringlemann Smoke Chart. See p. 765. Smoke Prevention: Bulletin No. 15, Univ. of 111., Vol. Geological Survey; Boiler Maker, May, 1909, Oct., 1909; Minn. Engr., Jan., 1910. Mechanical Stokers:Engr. U. S., Jan. 1, 1907, p. 83, Aug. 15, 1906, p. 540, July 2, 1906, p. 437;Cassier's Mag., Sept., 1906, p. 469; Power, Mar., 1906, p. 189, Aug., 1905, p. 487. Split Bridge Wall. 431 ; Bulletin No. 334, U. S. Cassier's Mag., Feb., 1907; CHAPTER V. SUPERHEATED STEAM; SUPERHEATERS. 102. General. — The steam engine fails to realize the efficiency of the ideal engine chiefly on account of cylinder condensation. The loss in heat due to this cause is seldom less than 10 per cent of the total supplied, and often as great as 40 per cent. If the steam is superheated before being admitted to the cylinder, condensation may be reduced or prevented entirely, as was recognized as early as sixty years ago, but the mechanical difficulties encountered prevented the practice until within the past few years. The principal advantages of superheated steam in connection with steam-engine work are: 1. At high temperatures it behaves like a gas and is therefore in a far more stable condition than in the saturated form. Considerable heat may be abstracted without producing liquefaction, whereas the slightest absorption of heat from saturated steam results in condensa- tion. If superheat is high enough to supply not only the heat absorbed by the cylinder walls but also the heat equivalent of the work done during expansion, then the steam will be dry and saturated at release. This is the condition of maximum efficiency in a single cylinder. (Ripper, " Steam Engine Theory," p. 155.) Greater superheat than this will result in a loss of energy unless the steam is exhausted into another cylinder. To obtain dry steam at release the steam at cut off must be superheated 100 to 300 degrees F. above saturation tempera- ture, depending upon the initial condition of the steam and the number of expansions, a higher degree of superheat being required for earlier cut off. A superheat of 200 to 275 degrees F. at admission is necessary to insure dry steam at release in the average single-cylinder engine cutting off at one-fourth stroke, boiler pressure 100 pounds gauge. In most cases superheat is only carried so far as to reduce initial conden- sation, the steam becoming saturated at cut off, thus permitting efficient lubrication. There will be a reduction of approximately 1 per cent in cylinder condensation for every 7.5 to 10 degrees of superheat. In compound and triple-expansion engines the steam is ordinarily super- heated between each stage as well as before admission to the high- pressure cylinder. 152 SUPERHEATED STEAM; SUPERHEATERS 153 2. A moderate amount of superheat produces a large increase in volume, the pressure remaining constant, and diminishes the weight of steam per stroke for a given amount of work. For example, the volume of 1 pound of saturated steam at 150 pounds pressure (gauge) is 2.75 cubic feet, and its temperature is 365.8 degrees F. The total heat of one pound of this steam above the freezing point is 1193.5 B.T.U. By adding 110 B.T.U. in the form of superheat its temperature will be increased to 565.8 degrees F. (superheated 200 degrees F.) and its volume to approximately 3.5 cubic feet (specific heat taken as 0.55).* Thus an increase of 9.2 per cent in the heat effects an increase of 22 per cent in the volume, which means a corresponding reduction in the steam admitted to the engine per stroke. These figures are purely theoretical, as no allowances have been made for condensation of the saturated steam or for reduction in temperature of the superheated steam. 3. Superheated steam has a much lower thermal conductivity than saturated steam, and therefore, less heat is absorbed per unit of time by the cylinder walls. General Discussion of Superheated Steam: Engr., Lond., Dec. 31, 1909; Eng., Lond., Sept. 13, 1901, Sept. 4, 1903, p. 237; Eng., U.S., Dec. 15, 1902, p. 821, Oct. 15, 1906, p. 687; Engr. Mag., Feb. 1903, p. 778, Sept., 1903, p. 897, Feb., 1904, p. 757, June, 1904, p. 436, March, 1905, p. 943, Nov., 1905, p. 271, May, 1906, p. 269; Eng. Rec, July 8, 1905, p. 28; June 30, 1906, p. 783, July 28, 1906, p. 86; Power, Aug., 1904, p. 463, Sept., 1904, p. 558, Oct., 1904, p. 762, Jan., 1905, p. 23, Feb., 18, 1908, Serial; Eng., Lond., Jan. 8, 1904, p. 42; West. Elec, Nov. 14, 1903, p. 369; Proc. A.S.M.E., May 14, 1908. 103. Economy of Superheat. — Many comparative tests of engines using saturated and superheated steam under varying conditions of pressure and temperature have been made during the past few years, showing in most cases a gain in favor of superheat due to the reduction in steam consumption, but in some cases the extra investment and cost of maintenance neutralize this gain, resulting in an actual loss when measured in dollars and cents per horse-power hour. As far as steam consumption per horse-power hour is concerned, superheating usually increases the economy five to fifteen per cent * The most satisfactory equation for determining the specific volume of super- heated steam is that given by Knoblauch, Linde, and Klebe (Peabody, " Steam and Entropy Tables," p. 22): pv = 0.5962 T - p(l+ 0.0014 p) / 150 > 3 Q Q > Q0Q _ .0833 V p = pressure, pounds per square inch absolute. T = absolute temperature of the steam, degrees F. v = specific volume of superheated steam, cubic feet. 154 STEAM POWER PLANT ENGINEERING and in some instances as much as forty, the latter figure referring to the more wasteful types of engines. A fair estimate of the average reduction in steam consumption per horse-power hour with moderate superheating, that is, 100 to 125 degrees F., based on continuous opera- tion of existing plants, is: Per Cent. 1. Slow running, full stroke, or throttling engines, including direct acting pumps 40 2. Simple engines, non-condensing, with medium piston speed, includ- ing compound direct acting pumps 20 3. Compound condensing Corliss engines 10 4. Triple-expansion engines 6 A prominent European builder of engines guarantees steam con- sumption with highly superheated steam as follows: Pounds per I.H.P. hour. Single-cylinder condensing engines 13.5 Single-cylinder non-condensing engines^ 15.5 Compound condensing engines 10 Triple-expansion condensing engines 8.75 In comparing the performances of engines using saturated and superheated steam it is advisable to base all results on the heat con- sumed per horse power rather than on the steam consumption, since the latter is apt to give a false idea of the relative economies. The real measure of economy is the cost of producing power, taking into consideration all charges, fixed and operating, and the next best is the coal consumption per I.H.P. hour, but as a means of comparing the engines only, the heat consumption per horse power per hour or per minute is very satisfactory. See paragraph 181 for the influence of superheat on the economy of reciprocating engines and paragraph 193 for the influence on steam turbines. Economy of Superheat : Eng. Mag., Dec, 1904, p. 757, April, 1905, Sept., 1903, p. 108; Trans. A.S.M.E., 22-899; Engr. Rec, July 8, 1905, p. 28; Power, Sept., 1904, p. 558, Oct., 1904, p. 598, Jan., 1905, p. 23; Cassier's, Nov., 1903, p. 18. 104. Limit of Superheat. — In this country steam temperatures exceeding 500 degrees F. are seldom employed, while in Europe few if any plants are installed without superheaters, and 600 degrees F. is a common temperature. Experience has shown that with engines of ordinary design, slide- valves and Corliss, the temperature at the throttle should not exceed 500 degrees F. This corresponds to a superheat of 160 degrees F. with SUPERHEATED STEAM; SUPERHEATERS 155 steam at 100 pounds gauge pressure, and 130 degrees F. at 150 pounds. This degree of superheat insures practically dry steam at cut off in the better grade of engines. Just how far superheating can be carried with a given engine of ordinary construction can be determined by experiment only, but a temperature of 500 degrees F. is probably an outside figure and 450 degrees F. a good average. Higher temperatures are apt to interfere with lubrication and sometimes cause warping of the valves. With temperatures below 450 degrees F. no difficulties are ordinarily met with. Metallic packing has been found to give the best results for both piston rods and valve stem. It is generally assumed that a greater quantity of oil is required for lubricating valves and cylinders in connection with superheated steam, but experience seems to show that such is not the case. (Proc. A.S.M.E., May 14, 1908.) Forced-feed lubricators are the most satis- factory for superheated steam engines, since they insure a positive and copious flow of oil directly to the valves or other parts requiring it.* With highly superheated steam involving temperatures of 600 degrees F. or more the poppet-valve type of engine is ordinarily employed, though balanced piston valves are not uncommon. The poppet valve is not distorted by heat and requires no lubrication. In Europe these engines have been brought to a high state of efficiency, but have not been generally adopted in this country owing, no doubt, to the higher cost. 105. Specific Heat of Superheated Steam.f — The total heat of super- heated steam is given as „■ , , ~ . n ~y. ti = A -r Cpt, (li) in which X = B.T.U. in one pound of saturated steam above 32 degrees F. C p = mean specific heat of the superheated steam at constant pressure. t = degree of superheat, degrees F. Regnault determined the mean specific heat at atmospheric pressure to be 0.48 between 127 degrees and 226 degrees C. of superheating, and until recently this has been assumed to apply to all pressures and tem- peratures. As early as 1876 Hirn concluded from experiments made with a throttling calorimeter that the specific heat of saturated steam increased with the pressures and decreased at any given pressure if the steam became superheated. Since then numerous investigators have promulgated theories pertaining to this subject which have been far from harmonious and none has been universally accepted. Some experiments appear to show that specific heat is independent of pressure * Effect of Superheated Steam on Cylinder Oils. Mech. Engr., Lond., July 31, 1908, p. 115. f See paragraph 113a. 156 STEAM POWER PLANT ENGINEERING and degree of superheat, while others indicate an increasing value as the pressure and degree of heat increase. Still others corroborate Hirn's theory. 0.60 0.55 0.50 0.46 / \ / \ / \ A Al LPr assures '. ^b. per A \ g q.Ii ,.A 3S. / \ \ 00 300 400 500 Temperature -Deg. JEan 600 roo Fig. 82. Specific Heat of Superheated Steam, Knoblauch and Jakob. -23A25. U 16 — "77 P! IV 40 GO 140 160 80 100 120 Pressure-Lb. Abs. Fig. 83. Specific Heat of Superheated Steam, A. R. Dodge. 180 200 The maximum figure ranges as high as 0.8 and the minimum 0.48 for a given pressure and degree of superheat. The curves in Fig. 82 are based upon the experiments of Knoblauch and Jakob ("Mitteilungen uber Forschungsarbeiten," etc., Heft 36, p. 109, and Stevens' Indicator, October, 1905); those in Fig. 83 upon the SUPERHEATED STEAM; SUPERHEATERS 157 experiments of A. R. Dodge (Trans. A.S.M.E., 1907); those in Fig. 84 are plotted from tests of Burgoon, Carpenter, and Thomas (Trans. A.S.M.E, 1907); and those in Fig. 85 are based upon the investigation of Professor Thomas (Trans. A.S.M.E., December, 1907). These curves \ Specific Heat of Steam *70 .68 .66 ^64 ■§ -62 W.60 § .58 t» .56 .54 52 \ s L Curves in Black :-C.E. Burgoon » " Light Lines :-Callendar Marked C C 600"* V <$$* 30ot kv &> i ,500' ^c> "S^ v> 400 . 200? '. >> ^C-v ^ -100 t » wo" 1 i c, f ,,— ~^~ ^o~ .50 50^ 3S =5= » -c- "c ~ = c = rC- .48 i — c-f 5 ii »- " c 40 80 400 120 160 200 240 280 320 3( Degrees, F. of Superheat Fig. 84. Specific Heat of Superheated Steam, C. E. Burgoon. differ both in theory and in value of c p , but until further experiments prove otherwise the values in Fig. 82 may be accepted as sufficiently accurate for all practical purposes. The values given by Knoblauch and Jakob have been accepted by authorities as the most reliable. Table 17a is based upon their results. Table 17b has been calculated by means of Linde's equation. (See footnote, page 153.) TABLE 17. VALUE OF c p AT ATMOSPHERIC PRESSURE BY VARIOUS AUTHORITIES. Superheated Steam Cooled by Water- Jacketed Calorimeter. Publication and Date. Temp. Deg. F. c p at Atmos. Pres. Variation of c p with Increasing Pressure. Increasing Temp. Regnault Ann. de Chimie et de Physique, Tome 23. Sibley Journal, 5-1904. Trans. A.S.M.E. . . . Varied Varied Varied 0.4805 0.4844 0.48 None Increases Increases None None Carpenter (Jones) Dodge 158 STEAM POWER PLANT ENGINEERING TABLE 17 — Continued. Throttling Calorimeter. Saturated Steam Expanded to Lower Pressure. Author. Publication and Date. Temp. Deg. F. c p at Atmos. Pres. Variation of c p with Increasing Pressure. Increasing Temp. Grindley Phil. Trans., Vol. 194. 239 0.4317 Increases Increases Increases Increases Increases Increases Increases Increases Hirn Griessmann Peake Zeit. V.D.Ing.,52, 1903. Proc. Royal Soc. A-509, 1905. Sibley Journal, May, 1904. Sibley Journal, May, 1904. Sibley Journal, May, 1904. 269 Varied Varied Varied Varied 0.506 0.43 0.463 0.4825 0.48 Increases Increases None None None Carpenter (Stew- art and Marble).. Carpenter (Hoxie and Wood). Carpenter (Sickles). Superheating Steam Electrically. Peake Carpenter (Berry) Carpenter (Thomas). Lorenz Knoblauch and Jakob. Proc. Royal Society A509, 1903. Trans. A.S.M.E., Eng. Mag., March 1907. Z.V.D.I.,No. 20... Engineering, L, Feb. 22, 1907. Varied 0.46 Varied 0.48 212 0.49 402 0.487 212 0.445 700 0.49 None Increases Increases Increases Decreases None None Decreases Decreases Decreases then increases From Combustion of Explosive Gases. Mallard and LeChatelier. Sarran and Vieille Langen Zeit. V. D. Ing. Tome 48. ...Do ...Do 212 212 212 0.46 0.464 0.463 None None None Increases Increases Increases From Calculation. Reeve Wor. Poly. Journal 215 265 236 0.39 0.4895 0.38 0.568 0.468 0.36 0.4805 0.513 0.493 0.479 Increases Increases Increases Hirn. Decreases London Engineer . . Zeuner Weyrauch Perry Zeit. V. D. Ing., Tome 48. Steam Engine Thermodynamics . . Rose Technic, 1905 Publication by au- thors, Berlin, 1905 212 212 Varied 284.4 212 356 Decreases Increases None Increases Increases Increases None Roentgen None Knoblauch, Linde, Decreases SUPERHEATED STEAM; SUPERHEATERS 159 3 < o w xi H 7J 3 1-3 Q ■o H a e8 «*1 fl H ■3 X S3 tf JQ a O M X o fa gg o H E < K x X o fa *^ M r u fa a fa 3 w TJ l?i 1 < w a X! fa "eg CD a 3 OQ O 8 CD &b - t>- O0 GO GO ^H ^^ ^* ^* ^r OOHNM to to to to to t^ CO OS tO t-h CO Tj< -OtOtO tO to tO to to O o CN tOCOCOM lOCOtOtON tOOJMlOrH NNOOXOi lO io >o>o >o GO CO CN OS CO Tt< IO CO CO t^ tO to to «o to t^ toa^oocN lOiOtOtON CO O CO CO CO t>- GO GO GO OS CO GO CN OS GO O i-i CO CO <* io»o»o »o o CO CN O GO CO iOtONNCO tO to to to to O CO OS "^ OS (M lomtotoN tOO^N'* t— GO GO GO OS OS CN "* ■* iO OtNCO<*iO IOU5 lO >0 lO CO CO CN O OS CO t^ GO OS OS tO tO to to to o CO OS •>* OS CN iCOtOtON NrHTjIOOlO t^ GO GO GO OS rHCNCO-*lO lO «o iO IO lO GO GO t^ CO to CO t- GO OS O IO to to tO CO o - lO05Tt*O)C0 lOiOtOtON NHiOOOtO t~- GO GO GO OS CN CO OS O CN i-H CN CO to CO IO lO IO tO lO CN CO CO CN CN NOOOJOth to «0 to CO CO o r0 lOiOtOtON N(NtOOOO I>- GO GO OS OS N^OOCNN rHCO'* tON to to to tO to O CO IO OS t-h OS O t-h CN •* to CO CO CO CO o OS 10 05^©M lOiOtOtON ■^ ^t ^ ^< -^ GO CO CO O OS t- GO GO OS OS OS t^ CN CO CN t-h CO IO CO GO tO to tO to to CO t-h -* GO CN OS t-h CN CO to tO CO CO CO CO o 00 to OS ^ OS "^t* lOiOtOtON ^* ^i ^r ^i ^H 00MNHO t^ GO GO OS O "^ "^ ^ ^f to t-h O CO CN OS CN ■* tO t- GO to to to to to to o to O CO O CN CO tO CO CO CO CO CO CO o ^Jr OS "^ OS ^r lOiOtOtON O0CONHH t>X00 03O Tft ^H ^ ^ lO CO CN O OS t^ CN ■* CO t^ OS tO to to IO to TfH t-H GO •* CO t-h CO ■* CO GO CO CO CO CO CO o co •^ © Tfri © -^t* ifltOtONN OS CO GO tM CO t^ GO GO OS O "^ "^ "^ ^ *o to CO to to «o CN -^ CO GO O tO to to to CO CO CO CO t-h CN N^toceo tOtOtOtON © to ^ © "^ o ^ ifl to to N N OS •* GO CN -* t— GO GO OS O ^* ^H T^H tH^ IO CN tO t^ OS t-h tO tO «o to CO '^tOCOO'* M«ONO■ CN CO to !>• OS CN tO to to tO CO tO CN tO CN O ■* f^ OS CN "^ COtOtONN o oo co © to © »o OtOtONN ^* ^* ^* ^* "^ O ""^ OS ^ t» GO GO GO OS O ^H ^* ^t ^H lO CO CO O CO CO CO tO GO O CO to to to CO CO NtOCNtOCO lOCCH^N tOCONNN o CN NOiOO"5 U5tOtONN OWO^OO GO GO OS OS O ^H ^< ^H ^1 IO to o to to to CO CO GO t-h t*t tO to to CO CO to O O o to t^ T-H -* t^ <^> CO l>- t^ t^ GO o NOWOif) lOtOtONN ^* ^* ^* ^t 1 ^^ OiOOiOO) GO GO OS OS O *^ ^* ^- ^* *0 o o o o o -* t^ OS CN CO to to to CO CO o o o o o OS CO t^ o to tONNOOOO t-h tO o to o t-h ,-h CN tO O iO o © CN CO CO ■* to to o to o to NONiON o to o »o O O CN tO t- O CN CN CN CN CO qouj aj'enbg jad spunoj 'ajnssajj a^rqosqy 160 STEAM POWER PLANT ENGINEERING s < w H 7J Q H H TJ g a 3 (H O W Pn fa P o 0Q -fa N U 3 fa O o > 1 fa tl 0) t_i CO a 3 W o CO p o © © N- © 00 OOCOMN © © ■* CN t- CN ©> © »0 r-H CN CN t~- © CO © r-H OS r-H CO OO © "<*< -<^ CO r-H OO r-i C75 r-H CN O0 CO CO CN CO CN CN CN CN »o o © tJ< © CO lO CO © "^ CO CN <* © t^ tC -*l © (M^^rtrH 00 CO oo © -«^ CO CO lO OO CO t-*- »0 "* CO CO lO -* © © CN © CO ■* r-H O CN CN CN CN CN o «5 OOONiOOO (NOc*CO(M r-H OS © CO »0 © CO -<*! 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Assume that the " heat of the liquid " in the exhaust steam is returned to the boiler and that the specific heat is 0.48 in one case and 0.8 in the other, these being the extreme values given by different experimenters for the given conditions. Specific heat B.T.U. in 1 pound of saturated steam above ideal hot- well temperature B.T.U. in superheat B.T.U. in 1 pound of superheated steam . . B.T.U. per I.H.P. per minute Case I. Case II. 1118.0 144 0.48 162. 1 1118.0 240 0.8 1262.0 1358.0 192.4 On the basis of the specific heat of 0.48 the heat consumption, 162.1 B.T.U. per I.H.P. per minute, is exceptional and corresponds to an equivalent saturated steam consumption of 9.6 pounds per I.H.P. hour, whereas the performance of 192.4 B.T.U. per I.H.P. per minute based upon the specific heat of 0.8 has been excelled by a number of actual engines using saturated steam. 106. Superheaters. — The installation of a superheater is equivalent to an increase in boiler capacity. The superheater may be independ- ently fired or can be arranged in connection with the boiler proper. The heating surface is usually of wrought iron, mild steel, cast iron, or cast steel. Engineers are not agreed as to which arrangement or which material gives the most economical returns. The recent symposium on superheated steam conducted by the research committee of the American Society of Mechanical Engineers (Trans. A.S.M.E., 1907) clearly indicated this lack of agreement. The requirements for a suc- cessful superheater are: 1. Security in operation, or minimum danger of overheating. 2. Economical use of heat applied. 3. No exposure of joints to the fire. 4. Provision for free expansion. 5. Disposition such that it may be cut out or repaired without interfering with the operation of the plant. 6. Ease of application to existing plants. SUPERHEATED STEAM; SUPERHEATERS 163 Nearly all superheaters depend upon carrying steam at a high velocity through small tubes in the form of return bends or coils and arranged to be heated by the hot gases in the boiler furnace or from some other source. The independently fired superheater has the following advantages: 1. The degree of superheat may be varied independently of the per- formance of the boiler. 2. It can be placed at any desirable point. 3. Repairs are readily made without shutting down boilers. Some of the disadvantages are: 1. It requires separate firing and extra attention. 2. Saturated steam can only be furnished in case of a breakdown to the superheater. 3. Extra piping is required. 4. Extra space required. Standard practice in this country advocates that the superheater be contained within the boiler setting. Of two hundred recent installa- tions, one hundred and eighty, or ninety per cent, were of this type. Fig. 86. Babcock and Wilcox Superheater. 107. Babcock & Wilcox Superheater. — Fig. 86 shows the applica- tion of superheating coils to a Babcock & Wilcox boiler, illustrating the indirectly fired type. The wrought-iron tubes are bent into U shape, the ends being connected into manifolds, the upper one receiving the saturated steam from the boiler and the lower one the superheated 164 STEAM POWER PLANT ENGINEERING steam after it has traversed the superheater tubes. A small pipe con- nects the lower manifold with the water space of the boiler by means of which the superheater may be cut out if desired, or flooded when starting up. Any steam formed in the superheater tubes is returned into the boiler drum through the collecting pipe, which, when the superheater is at work, conveys saturated steam into the upper manifold. When steam pressure has been attained the superheater is thrown into action by draining the water away from the manifolds and opening the super- heater stop valve. The tubes are free at one end and the manifolds are not rigidly connected with each other, thus avoiding expansion strains. With the proportion of superheating surface to boiler surface ordinarily adopted the steam is superheated from 100 to 150 degrees F, STEAM PIPE FEED PIPE SAFETY VALVE Fig. 87. Stirling Superheater. 108. Stirling Superheater. — This superheater consists of two drums, Fig. 88, connected by seamless drawn tubes two inches in diameter. It may take the place of the middle bank of tubes in the Stirling boiler as shown in Fig. 87, or be installed in the final pass of the gases in the back SUPERHEATED STEAM; SUPERHEATERS 165 of the boiler. The drums around the tubes are protected from intense heat by asbestos cement. A pipe connecting the front drum of the boiler with the lower drum of the superheater permits the coils to be flooded in starting up or when the superheater is not needed. In this Fig. 88. Arrangement of Tubes ; Stirling Superheater. case the superheater acts as additional boiler-heating surface. The upper drum is divided into three and the lower into two compartments. The tubes are arranged with alternately wide and narrow spacing, so that any tube may be removed without disturbing the rest. The flow of steam is indicated by arrows. 109. Foster Superheater. — Fig. 89 shows the application of a Foster superheater to a Babcock & Wilcox boiler. This device consists of cast-iron headers joined by a bank of straight parallel seamless drawn-steel tubes, each tube being encased in a series of annular flanges placed close to each other and forming an external cast-iron covering of large surface. The tubes are double, the inner tube serving to form a thin annular space through which the steam passes as indicated. Caps are provided at the end of each element for inspection and cleaning purposes. Foster superheaters are more costly than plain-tube super- heaters, but are longer lived and offer a much larger heating surface in proportion to the space occupied. Fig. 91 shows a Foster superheater arranged for independent firing. 166 STEAM POWER PLANT ENGINEERING The " Schwoerer " superheater, which is somewhat similar in external appearance to the Foster, differs from it considerably in detail, the heating surface being made up of suitable lengths of cast-iron pipe ribbed outside circumferentially and inside longitudinally. The ends of the pipes are flanged and connected by cast-iron U-bends. The intention is to provide ample heating surface internally and externally, with a compact apparatus. Fig. 89. Foster Superheater in Babcock & Wilcox Boiler. 110. Independently Fired Superheaters. — The Schmidt superheater, Fig. 90, consists of two nests of coils, A and D, of equal size and dimen- sions, connected to cast-iron headers and 7. Saturated steam enters the first nest of coils through C and passes into header 0. From the steam, which is now dried and partly superheated, flows through the cast-iron pipe E to header I, and thence through the second nest of coils into header adjoining 0, and through pipe R to the engine. In chamber D the steam and gases flow on the counter-current and in chamber A on the concurrent principle. This combination permits of a low flue temperature and high steam temperature without subjecting the tubes to an excess of heat as would be the case if the steam left the coils A at header I, where the furnace gases are the hottest. A steam SUPERHEATED STEAM; SUPERHEATERS 167 168 STEAM POWER PLANT ENGINEERING SUPERHEATED STEAM; SUPERHEATERS 169 temperature of 750 degrees F. and a flue temperature of 450 degrees F. are easily maintained with this apparatus. A mercury pyrometer T is fitted where the superheated steam enters the discharge pipe R. A thermometer cup L permits of checking the pyrometer by means of a nitrogen-filled thermometer. Each coil can be taken out separately and a new one put in without removing the others or dismantling the plant. Water produced by condensation while the superheater is inoperative collects in the bottom header N and escapes through a drain cock. If the steam supply should be suddenly shut off, the air door P is opened automatically by weight K. As soon as steam begins to flow it raises the weight through the opening of valve C and the door closes. The Schmidt superheater when arranged in the flue has practically the same construction as the independently fired. »!5 o z ua PS ^ u )- X win D5^i ^ ; ■ I t=fc* 1»>///>»}»://»i>> FEED WATER ECONOMIZER Fig. 92. Schmidt System of Combined Superheater, Feed Water Heater and Economizer. Fig. 92 shows a combination of Schmidt superheater, economizer, and feed-water heater which finds much favor with engineers on the Continent. 170 STEAM POWER PLANT ENGINEERING 111. Materials used in Construction of Superheaters. — Most super- heaters are constructed either of wrought iron, mild steel, cast iron, or cast steel, the latter material having the advantage of not being dam- aged by any temperature to which it is likely to be subjected, which does away with the necessity of damper mechanisms and simplifies the installation. On the other hand, cast-metal superheaters are usually ribbed after the fashion of an air-cooled gas engine, and are, therefore, very heavy and thick walled, necessitating a higher temperature for the same useful effect than in the case of the wrought-iron construction, but have the advantage of minimizing fluctuation of steam temperature which would otherwise be caused by a wide variation in temperature of furnace. One of the most successful cast-metal heaters is of European design and is constructed of a special alloy known as " Schwoerer " iron. Table 18 gives the yearly cost of repairs to piping and necessary brickwork for a number of installations equipped with cast-metal super- heaters of the " Schwoerer " type. Wrought iron and mild steel offer the advantage of lightness, ease of construction, and low first cost, but cannot be exposed to very high temperatures without injury, and consequently provision must be made for diverting the direction of the heated gases or for flooding the coils while the boiler is being warmed before steam is generated. Neither cast iron nor steel loses in tensile strength when subjected for a very short time to the temperature of superheated steam, but, on the contrary, may be stronger. Tests made by Professor Lanza (" Applied Mechanics," p. 489) showed that the tensile strength of steel dimin- ished from degrees F. to about 300 degrees F. and then increased, reaching a maximum between 500 and 650 degrees F. Cast iron and steel maintained their strength, with a tendency to increase, up to 900 degrees F. beyond which the strength is diminished. Ordinary cast-iron valves and fittings have shown permanent increase in dimensions under high superheat and in numerous instances have failed altogether, but sufficient data are not available to prove conclusively the unreliability of cast iron if the iron mixture is properly compounded and the necessary provision is made for expansion and contraction.* 112. Extent of Superheating Surface. — The required extent of super- heating surface for any proposed installation depends upon (1) the degree of superheat to be maintained; (2) the velocity of the steam through the superheater; (3) the character of the superheater; (4) the weight of the steam to be superheated; (5) the moisture in the wet steam; (6) the temperature of the gases entering and leaving the superheater; and (7) the conductivity of the material. * See Symposium on the "Effect of Superheated Steam on Cast Iron and Steel," Jour. Am. Soc. Mech. Engrs., Dec., 1909. . SUPERHEATED STEAM; SUPERHEATERS 171 9 k3 H S CO >> fa "i & a; cp 1! 1 s 2 1 co JS T) _g s CD £5 "2 t: s 3 CO C s * i CO a i a | a ~ Z £ a PC 1 CU 3 Is 31 .S "S 2 £ a> ffl CO d >» g n 5 i Average Daily Use. Hours. - , (24) Ud On account of the great variation in the values given for U, and the difficulty of determining d, t Xi and t 2 for different types of super- heaters, equation (24) is hardly applicable in practice. An empirical formula for determining the extent of superheating surface in connection with indirect superheaters which appears to con- form with practice is given by J. E. Bell (Trans. A.S.M.E., May, 1907) : 10 S (25) 2{T - t) - S in which x = square feet superheating surface per boiler horse power. S = superheat, degrees F. T = temperature of the products of combustion where the super- heater is located. t = temperature of the saturated steam. The value of T may be found from the equation 1 (T - t) ' 16 in which = 0.172 H + 0.294, (26) H = the per cent of boiler-heating surface between the point at which the temperature is T and the furnace. t as in (25). SUPERHEATED STEAM; SUPERHEATERS 173 83SVO as U3AO oassvd aovjuns onuvsh uxlvm so xn30 msa 174 STEAM POWER PLANT ENGINEERING Equation (26) is based upon the assumption that the heat trans- ferred from the gases to the water is directly proportional to the differ- ence in temperature; that the furnace temperature is 2,500 degrees F.; flue temperature 500 degrees F.; steam pressure 175 pounds per square inch gauge; one boiler horse power is equivalent to 10 square feet of water-heating surface. Example: What extent of heating surface is necessary to superheat saturated steam at 175 pounds gauge pressure, 200 degrees F., if the superheater is placed in the boiler setting where the gases have already traversed 40 per cent of the water-heating surface? Substitute H = 0.4 and t = 378 in equation (26), = 0.172 X 0.4 + 0.294 (T - 378)°' T = 950. Substitute T = 950 and S = 200 in equation (25), = 10 X 200 " 2 (950 - 378) - 200 = 2.12 square feet. The curve in Fig. 92a was plotted from equation (26), and gives a ready means of determining T and of observing the law governing heat absorption by the boiler between furnace and breeching. The abscissas represent the temperatures of the hot gases at different points in their path between furnace and breeching. The ordinates represent (1) the per cent of boiler-heating surface passed over by the hot gases, and (2) the per cent of the total heat generated which is absorbed by this heating surface. In the use of equation (26) the probability of error is greatest when considering a point near the furnace, since large quantities of heat are transmitted to the tubes by radiation from the fuel bed which are not taken account of. For most practicable purposes the assumption is sufficiently accurate. For the application of the curve in Fig. 92a to the design of direct and indirect superheaters for various degrees of superheat, see " Stir- ling," published by the Stirling Boiler Company, pp. 92-96. 113. Performance of Superheaters. — Published tests of both directly and indirectly fired superheaters cover such a wide range of conditions of installation and operation that general conclusions cannot be drawn, but it may be of interest to note briefly the performances in a few specific cases. SUPERHEATED STEAM; SUPERHEATERS 175 The curves in Figs. 93, 94, and 95 are plotted from tests of a Babcock & Wilcox boiler, with 5000 square feet of water-heating surface, equipped with superheating coils of 1000 square feet area, as illustrated in Fig. 62. The furnace with ordi- nary short ignition arch was pro- vided with chain grate of 75 square feet area. Fig. 93 shows the relation be- tween degrees of superheating and total horse power of boiler and superheater. Fig. 94 shows the relation be- tween the horse power produced Horse Power Produced in Boiler Fig. 94. Percentage of Horse-Power Pro- duced in the Superheater of that Devel- oped in the Boiler. ri 5 Percentage of Boiler H. Produced in Superheat © Ol O Of o 900 / 800 /: / 700 600 M S 500 £ '"■'■'1 ■••/ J 400 o / \/- 300 / / . / 200 / / 100 / / 50 100 150 200 Degrees in Superheat— F 250 120 110 u ioo 1 90 a> 1 8 ° & % 60 > 2 50 o / / / / I '.[■■ • £.: i r 6 ^ | 30 w ■ ■' 20 10 I egret 1 isofS X) super i ueat- 50 -F 2( X) Fig. 93. Relation of Degree of Superheat to Total Horse-Power Developed. Fig. 95. Relation of Degree of Superheat to Horse-Power of Superheater. in the boiler and the percentage of boiler horse power produced in the superheater. Fig. 95 shows the relation between the degree of superheat obtained and the horse power developed in the superheater. 176 STEAM POWER PLANT ENGINEERING Hi 3 SI 31 H P cc Em O CO OS (M 00 100 00 lO ©^ OS CN »0 OO^MC* OO t- t>. ■«* 0500 W OO <0 t- cd COCOCOO HOSN^MfOtOO lO S-i CO ^+1 © N«Q0lON00 cS CO OOH t-H "tf< iO O0 00 Pi d> t-h t*i t-i t^ o m T— 1 T— 1 43 • 43 43 *-l fc* <-! 43 43 43 03 a; 03 3 3 ofefeppp a fi C 03 0) O O r< ... 03 03 03 a? 3 fe £ O S> ^ b.K^WPQ 03 03 03 2 2 £ £ <£ bfi be . . . q^ n. n. 33 PhOh^o303... i-h Hh c ph cr cr cr E8o q> 03 S-i t-> 03 ^^ 03^ 03 bfi 03 • • ; 03 03 n 3 cd s- a. +; 1? 'o *o ^ 43 o o " ':Sc5§2g>§g 3 •ro , '-;03 Q 1bCM M ~'o3 03 ^ rfi fi 43 ^ O 03 u 4- XI a +- *c s. a r \ S- .fi £ cl^ 03i: CT 03^ 03 t w a, „ a - <^ m ^ O t a^ g 1 & £ 43 2^1 ^ C3 ^ i P 03 C3--J 03 a 03 O O 03 +1 fe OPQPtHGC 03 ai.S BX O 3 fi 3 2 "gj-i 1 ■&s& 8 G -£ fi W) .2* § fi * ^^ S 4. §43^ 03 +3 03 Xj U fi *o 0) T> O r_ ^ 03^ 55 j SOW 2 ogs t O 3 O fc rH o ^ a -, ? o fi ^ 03 C fi"fi « j b*0 03 „ •P ft 45 1/ c3 . o3 o 03 £ 03 — Wft ffl SUPERHEATED STEAM; SUPERHEATERS 177 O *-» i-i With Super heat- ing. »0 «? CO H njl w lO ©i r>. »o OS OS t>- CO cOi-inOcoW rH CO io CO M< OS ■<* i-i io co Hi 00 CO «M ©1 rH I© "* co cm CO 00 OS r-l CO CM With- out Super- heat- ing. O CM CO ... t^. CO CO ^+1 • 09 CO CO OS CM *© • 05 CM "* CO OS • ■>* i-H !>. • . CO OS IO -<* »o CO OS OO »o With Super heat- ing. O !>■ J>- O ©J s© »o OS ^ (M HOffiHo* 00 00 t-i i-f io O »• CO <>» rH i—i a» x* "tf HH-tlO IG oo »o CO r>» Tfi co o o With- out Super heat- ing. ^H io co co • OS co t-T ©4 -* os i-l O • rH HitOfli 'N . O CM CO QO t- is CD *» vt i-i O l>- "tf rJHI>- rH 5© t>. oo a 1,955.86 chwoere Patent 45.21 CM <* rH IO ©jl OS © t>i © oo i-l CM IO ^ 00 O rH i-i OS ■* -CO <* C « a With- out Super heat- ing. CO t>- OS CM 00 • CiD CM »o CM CO ©1 • I© ai rHCNUjOS •»« 3 . OS o io rH lO I>i With Super heat- ing. CO O 1>- CO ©CJ CO rH «HH SO *° CM CO rH OS «>• CM OS id SO OS a CO OC o: CM O0 00 OS CO iO rH If} tC> CO 1 OS IO CM ^ »o CO With- out Super- heat- ing. rH -^00 00 • OS O0 co OS 05 • OS c^r CM* NNOJ't •»« CM 00 O0 *© • IO rH «D • CO "tf O0 00 co to CM With Super heat- ing. «© J>- CO CO oo »o CM OS 00 iO *© «© CO 00 OS CO O CO 00 Tt 00 Tf oo^cc C rH -># rJH?© O0 rH ** -rJH CN »fl With- out Super- heat- ing. 00 I>- l>- CO IO CO OS ocT CM OS OS coO • i-H TtH TtlO • rH IO • CO IO rH CM ■"*! With Super heat- ing. CM lOONWO) CO 00 "^ lO CM* CO i-I rJH (N rji W ^H os OC rH CO O CO CM ,_, 00 OC rH rH TH ^ CO CO »o x: .^ 5rS -*> .•_ "* cd a a 2 Pn. tu p& a a a c cd a a o o ™ * r„ <^ 5 CS ir, 55 C3 ^ eed-water consumption of heating surface oal consumption per sq of grate area oiler pressure (absolute eed-water temperature emperature of gases of team temp, on leaving fficiency of boiler plant icrease in heat efficienc use of superheated ste •ecrease in grate area r( to use of superheated i •ecrease in heating surf due to use of superhea ) t a 'c c JS CD O cj a is » ^i ! cd a pC) a > o J ? 'Si > i i 1 £ t- a 'c s. t 5 p ! 1 3 1 E- i ff a 2 C p: rH O PC fX( HoqH rH « R 178 STEAM POWER PLANT ENGINEERING w Pm P m O H o 1 m » w ^ O ^ O CI W PQ g * H H 1 tz OO OS o Boiler ith Su rheate o hN(N O l>. tH CO CO CM OS © GO oo io t- eo r-i i—l CO * s. 00 5P Boiler withou Super- heater CT> 1-H CO CO i-I t- ©' 00 CO eo y ~* t — ■^ CO t-, 3 <» . T»H • CO ■ ~ : ' !-< CO lO t>. Boiler ith Su rheate »o "* CM OS QO QO CO CM "* » bi 3 A J lO CO • © © Boile witho Supe: heate lO CM* l-H • oj »o t> CO • CO : I sJ iO GO toiler ih Su heate CM 00 CM i-H gq 00 "* i-H o co oo a th NiOCO CM 1—1 m •- ^ T-H CO * a lO QO 5S "3 £ ^ CM -^ CO © i2 o c3 S ■s .fl a -g CO t-i io «> s O +3 3; cc i—i CM 1 ^ i-H . t>. iO • OO iH J^ -* • CO dfefa 3 S "8 sr sq. i e, Deg, .Deg. per ho per ho Per ce & l : .d ^-e&g o : _Q JL. E ft 2 cS t- oj 0) S ft 4- B a E- i PQ fea >W O I-H SUPERHEATED STEAM; SUPERHEATERS 179 Tables 19 to 21 are taken from the report of Otto Berner (" Zeit. d. Ver. Deut. Eng." and reprinted in Power August, 1904). Table 19 compares the heat efficiency of a steam plant equipped with directly and with separately fired superheaters, the former showing a much higher efficiency. Table 20 compares different boilers with and without flue super- heaters, showing the effect upon the temperature of the flue gases. The gain in heat efficiency of the entire plant due to the use of the super- heater is decisive in each case. TABLE 22. - (Engineer, U. S., May 1, 1904.) Time of start. Time of finish . Hours run Average steam pressure Average water pressure, triple expansion head in feet Average water pressure, compound, head in feet Average vacuum of suction for triple and compound, inches of mercury Total head on triple, feet of water Total head on compound, feet of water . . . Total double strokes, triple Total double strokes, compound Gallons pumped from piston displacement, total, triple Gallons pumped from piston displacement, total, compound Gallons pumped from piston displacement, total, triple combined Gallons, total, pumped as measured by weir Per cent slip Foot pounds, weir Total coal consumed Per cent refuse Total refuse Total feed water Duty per 100 pounds coal Duty per 1,000 pounds steam With Superheater. Without Superheater. 12 noon, Feb. 8 11 a.m., Feb. 11 12 noon, Feb. 9 11 a.m., Feb. 12 24 24 79.3 1b. 79.4 1b. 0.99 1.05 7.10 7.10 22.90 23.21 29.05 29.46 33.04 33.39 30,557 34,114 35,395 32,158 2,854,023 3,186,247 2,930,706 2,662,682 5,784,720 5,848,930 4,492,680 4,549,480 22.3 22.2 1,163,815,819 1,184,983,596 5,015 lb. 6,410 lb. 23.7 18.7 1,188 1,203 38,399 50,960 23,206,696 18,486,483 30,308,498 23,253,213 Per cent increase of work per 100 pounds coal 25.5 Per cent increase of work per 1,000 pounds steam 30.2 Per cent saving in coal per foot pound work 20 . 2 Per cent saving in feed water per foot pound work 23 . 2 Average temperature steam leaving superheater 527 . 4 deg. F. Average temperature steam entering superheater 320. 1 deg. F. Average degree superheat 207 . 3 deg. F. 180 STEAM POWER PLANT ENGINEERING Table 21 shows the gain in heat efficiency due to the use of super- heaters in a number of plants equipped with fire-tube boilers. Table 22 gives the results of tests on one of the return tubular boilers at the Spring Creek Pumping Station of the Brooklyn Waterworks (leb. 9, 1904) with and without a superheater. The superheater, of the Foster type, was installed between the rear wall of the setting and the tube sheet. 113a. Properties of Superheated Steam. — The following equations derived by Prof. Goodenough of the University of Illinois and based upon the experiments of Knoblauch and Jakob, give a comparatively simple method of determining the various properties of superheated steam if steam and entropy tables are not available. The results as obtained from these equations agree substantially with Marks' and Davis' Steam Tables and the 1909 Edition of Peabody's Steam Tables. T = absolute temperature of the superheated steam, deg. F. p = absolute steam pressure, lbs. per sq. in. X = total heat, B.T.U. per pound. u = intrinsic energy, ft. -lbs. n = entropy. C p = true specific heat. X=T (0.372 + 0.00005 T)-p(l + 0.00035) -^ + 882.4, in which log C = 9.42383. u = T (202.44 + 0.0389 T) - j^. (1 + 0.00025 p) + 686242, in which log C = 12.20551. n = 0.85657 log T + 0.0001 T - 0.25392 log p - p (1 + 0.00035 p) £ M - 0.4300, in which log C = 9.31469. C p = 0.372 + 0.0001 T - p (1 + 0.00035 p) ^- 5 , in which log C = 9.96790. The mean specific heat may be obtained by subtracting the total heat of the saturated steam from that of the superheated steam and dividing the difference by the degree of superheat. The specific volume may be determined from Linde's equation as stated at the bottom of page 153. CHAPTER VI. COAL AND ASH-HANDLING APPARATUS. 114. General. — The cost of coal and its delivery into the furnace are usually the largest items in the operating charges, hence large central stations are located, when practicable, adjacent to a railway line or water front, to minimize the cost of handling coal and ashes. Isolated stations in the business districts of large cities are usually unfavorably situated, so that the cost of handling coal and ashes is a large percentage of the total fuel cost. In large stations the amount of fuel and ash handled frequently warrants the expense of elaborate conveyor systems which would not be justified in smaller plants. In whatever way coal is supplied provision should be made for storing a quantity sufficient to operate the plant for some time in case the supply is inter- rupted, thereby guarding against an enforced shut-down. If adjacent to a railway line, a side track must be provided for switch- ing the cars. As bottom-dumping cars cannot be depended upon, provision should be made for unloading by hand. If coal is delivered by water, clam-shell drop buckets are ordinarily used for unloading the barges. If the power house is located at some distance from the railroad or water the coal is generally hauled by teams in two to five- ton loads. 115. Coal Storage. — In small stations the storage bins or coal bunkers may usually be located within the building, but in larger plants the quantity of coal consumed daily is frequently such that an immense space would be required to furnish storage capacity for even a short period of time. For example, one of the large central stations in Chicago burns an average of 30 tons of Illinois screenings per hour throughout the year. Allowing 45 cubic feet to the ton this would necessitate a space of 45 x 30 x 24 = 34,800 cubic feet to store coal for one day's operation. A ten-days run would require a coal pile 50 feet wide, 30 feet high, and 232 feet long. It is a good plan, if the location and character of the plant permit, to carry four or five days' supply within the plant and provide a separate building for the coal reserve. Such provision is made in the power plant of the New York Edison Company, which has a storage capacity of 150,000 tons in addition to that of the overhead bunkers. 181 182 STEAM POWER PLANT ENGINEERING Exposed coal piles are objectionable, because of freezing in winter, the crust sometimes freezing so hard as to necessitate the use of dynamite to break it; moreover, a slow depreciation in heat value takes place, especially with bituminous coal. This depreciation is more rapid in warm weather and in the tropics. Stored coal is oftentimes subject to spontaneous combustion, particularly when there is a large content of iron pyrites. Coal bunkers or hoppers are ordinarily placed on the same level with boiler-room floor or above the boiler setting. The former is the cheaper as far as first cost is concerned, but necessitates additional handling of the fuel before it can be fed to the stokers. In the overhead system the coal gravitates to the stoker through down spouts. Overhead bunkers are usually found where real estate is costly. They are gener- ally constructed of steel plates lined with concrete or of reenforced concrete. The bottoms slope at an angle of 35 to 45 degrees and empty into the coal chutes or down spouts. Fig. 99 shows the general appear- ance of a single overhead bunker and Fig. 441 that of a double bunker. In some bunkers the floors are made with very slight slopes, but it is not advisable to use a slope less than the angle of repose of the coal, as it may be necessary to shovel the coal over the spouts. Convenience in framing makes the 45-degree slope the more desirable. Separate bunkers for each boiler are preferred to continuous bunkers, since fire in the coal is more readily prevented from spreading. In the new power house of Swift & Co., Chicago, 111., the bunkers are of circular cross section instead of rectangular as is the usual practice. The capacity of the cylindrical hopper is considerably less than that of a rectangular hopper of the same width, but is much cheaper to construct. Ash bins are invariably lined with concrete or brickwork, since the corrosive action of the ashes would soon destroy the bare iron, and are usually located alongside the coal hopper, as in Figs. 97 and 99, so that they may be discharged by gravity. The angle of repose of most ashes is approximately 40 degrees, but the 45-degree angle is preferred on account of convenience in construction. Coal Storage: Power, April, 1907, p. 217, Aug., 1899, p. 3, Nov., 1904, p. 651; Eng. Rec, Sept. 23, 1905, p. 534, June 1, 1902, p. 532, July 4, 1903, p. 4; Eng. News, July 11, 1907, June 5, 1902, p. 463, April 3, 1903, p. 272; West. Elec., Oct. 28, 1905, p. 335; Trans. A.S.M.E., 23-473. Coal Storage under Water: Eng. News, Dec. 24, 1908; Eng. Min. Jour., Dec. 1, 1904, p. 975; Engineering, Sept. 4, 1903, p. 863. Calorific Value of Weathered Coals: Trans. A.S.M.E., 20-333; Bulletin No. 17, Univ. of 111., Aug. 26, 1907. Design of Coal and Ash Bins : Eng. News, July 21, 1904, p. 62; Eng. Rec, Sept. 1, 1900, p. 201; Power, Nov., 1899, p. 14, Nov. 1904, p. 651; Elec. Age, March, 1907, p. 141. COAL AND ASH-HANDLING APPARATUS 183 116. Coal Conveyors. — Coal is carried to the stokers in a variety of ways, depending upon the location of the plant, the type of stokers, and the personal tastes of the builder. Of the various methods the following are the most common : 1. Hand shoveling from coal pile to furnace. 2. Wheelbarrow or hand car and shovel. 3. Bucket conveyor. 4. Belt conveyors. 5. Hoist and hand cars. 6. Hoist and automatic cable cars. 7. Combinations of the above. Coal-Handling Plants for Power Houses : Am. Elec, June, 1900, p. 266, Oct., 1901, p. 486; Cassier's, April, 1905, p. 480; Elec. World., Dec, 1901, p. 463; Engr., U.S., July 1, 1904, p. 461; Jan. 1, 1905, p. 4; Eng. Rec, April 5, 1902, p. 322. Ash Handling : Elec. World., Oct. 5, 1901, p. 569; Eng. News, Oct. 19, 1905, p. 403; Eng. Rec, May 10," 1902, p. 435, Jan. 17, 1903, p. 85, Feb. 7, 1903, p. 153, Oct. 28, 1905, p. 482, Oct. 7, 1905, p. 396, Dec. 9, 1905, p. 655; Power, Oct., 1904, p. 507, July, 1904, p. 422; St. Ry. Jour., Jan. 5, 1901, p. 11. 117. Hand Shoveling. — Where possible the coal is dumped direct from the cars or wagons into bins located in front of the boilers. In such instances one man may handle the coal and ashes and attend to the water level of 200 horse power of boilers equipped with common hand-fired furnaces. With stoking and dumping grates 300 horse power may be controlled by one man and from 800 to 1000 horse power with chain-grate stokers. This refers of course to average good coal not too high in ash nor productive of much clinker. Sometimes the coal cannot be stored in front of the boilers but must be hauled by wheel- barrow, cart, or rail car. For distances over 100 feet and quantities over 20 tons per day the cost of handling the coal in this way may justify the installation of an automatic conveyor system. Hand-fired furnaces and manual handling of coal and ashes are usually associated with small plants of 500 horse power and under, but a number of large stations are operated in this way with apparent economy. A notable example is the new (1907) steam-power plant of the Wood Worsted Mill, Lawrence, Mass., in which 40 return tubular boilers are fired by hand. A tipcart with a capacity of one ton brings the coal a distance of 100 to 200 feet to the firing floor, and firemen shovel it on to the grate. Four men are stationed at the coal pile. One man drives two carts (one of which is being filled while the other is gone with its load), sixteen firemen attend to the furnaces, and two men dispose of the ashes. 184 STEAM POWER PLANT ENGINEERING Most large plants, however, are equipped with conveying machinery, not so much because of the possible reduction in cost of operation, taking into consideration all charges fixed and operating, as because of the large and often unreliable labor staff which it dispenses with. Hand shoveling is sometimes necessary even with modern unloading devices on account of the freezing of coal in the cars. This is par- ticularly true of washed coals, and it is not unusual to have an entire car load solidly frozen. In this case it has to be picked and shoveled by hand, or the unloading tracks must be equipped with steam pipes and outfits for thawing purposes. A good man is capable of shoveling 40 to 50 tons of coal in eight hours when unloading a car, provided it is only necessary to shovel the coal overboard. 118. Bucket Conveyors. — One of the most common methods of automatically handling the coal from car to bunker is by means of an endless chain of traveling buckets. Many of the largest central stations in this country are equipped with such systems. The details of opera- tion are best illustrated by a few examples. Fig. 96 gives a diagrammatic arrangement of the link-belt over- lapping pivoted bucket carrier, and Fig. 97 illustrates its application to a typical boiler plant. Coal is discharged from the railway cars into a track hopper and from there delivered by a " feeding apron " into a crusher which reduces it to such a size as can be conveniently handled by the stokers. It is then discharged into a short bucket conveyor, which carries it to the main system of buckets, and it is elevated to the proper level and discharged into the overhead bunkers. The discharge is effected by special tripping devices which engage the buckets and turn them over. The ashes are dumped from the ash pit through a series of chutes into the lower run of buckets, by which they are elevated and discharged into the ash hopper alongside the coal bunkers. From the ash hopper the ashes discharge by gravity directly into the railway cars below. The system is operated by means of two motors, one driving the crusher and the other the main bucket system. The buckets are made of either sheet steel or malleable iron. In Fig. 96 the coal is fed to the crusher by the " reciprocating feeder," which is usually placed directly under the track hopper. The feeder consists of a heavy steel plate mounted on rollers and having a recip- rocating movement effected by a crank mechanism from the carrier. The amount of coal delivered depends upon the distance the plate moves, and this can be varied by changing the throw of the eccentric. The number of strokes corresponds to the number of buckets. Any size coal can be readily handled. When the distance from track hopper to carrier is so great that the reciprocating feeder is not practicable a COAL AND ASH-HANDLING APPARATUS 185 186 STEAM POWER PLANT ENGINEERING I '3 cr COAL AND ASH-HANDLING APPARATUS 187 188 STEAM POWER PLANT ENGINEERING Fig. 99. Coal and Ash-Handling System in the Power House of the South Side Elevated Railway Company, Chicago. COAL AND ASH-HANDLING APPARATUS 189 continuous or " belt " feeder is used to supply the crusher with fuel. The " equalizing gear " is designed to impart a pulsating motion to the driving sprocket wheel which will counteract the natural pulsation to which long pitch chains are subject, producing violent increase of the normal strain at frequent intervals. This is accomplished by driving the spur wheel with an eccentric pinion, causing the pitch line to describe a series of undulations corresponding to the number of sprockets on the chain wheel. Figs. 99 and 100 show the general arrangement of crusher and " cross conveyor " in the old portion of the South Side Elevated Power House, Chicago. A coal and ash system similar to the one illustrated in Fig. 97 for a plant consisting of eight 350-horse-power boilers will cost in the neigh- borhood of $8,000, completely installed. This does not include the cost of coal and ash bunkers. The Hunt conveyor, Fig. 101, while usually called a " bucket " conveyor, is in fact a series of cars connected by a chain, each having a body hung on pivots and kept in an upright position by gravity. The chain is driven by pawls instead of by sprocket wheels. The " buckets " are upright in all positions of the chain, consequently the chain can be driven in any direction. The change of direction of the chain is accom- plished by guiding the carriers over curved tracks. The chain moves slowly, and the capacity is governed by the size of the buckets. The ordinary size buckets carry two cubic feet of coal and move at a rate of fifteen buckets a minute, carrying about 40 tons per hour. Two methods of filling the buckets are employed, the " measuring " and the " spout filler." In the former each bucket is separately filled with a predetermined amount by a suitable " measuring feeder." In the latter the material is spouted in a continuous stream, necessitating the use of overlapping buckets to prevent spilling of the material. Fig. 102 shows an application of the Hunt system to the old plant of the Baltimore United Railways and Electric Company. Fig. 103 gives a sectional elevation of the coal and ash-handling machinery at the power plant of the Commercial National Bank Build- ing, Chicago. Underneath the sidewalk on the Clark Street side of the building is a coal-storage bin of 600 tons capacity, served with a bucket conveyor. One leg of the conveyor reaches down to a level below the track of the Illinois Tunnel Company. By this arrangement coal can be delivered either by cars in the tunnel or by wagons from the street. In taking coal from storage a gate at the lower extremity of the hopper is opened and the coal filling the buckets is elevated and tripped into any one of the screw conveyors leading from bucket con- veyor to boiler hopper. The ashes are shoveled from the ash pits into 190 STEAM POWER PLANT ENGINEERING i COAL AND ASH-HANDLING APPARATUS 191 j^^TT^ Fig. 101. Driving Mechanism of Hunt Conveyor. Fig. 102. Coal and Ash-Handling System at the Old Power House of the Baltimore United Railways and Electric Company. 192 STEAM POWER PLANT ENGINEERING cars running in a cross tunnel under the boiler floor, and by these cars are transferred to a dump at one side of the boiler room and discharged into Illinois Tunnel Company's cars for removal. Fig. 103. Bucket and Screw Conveyor at Commercial National Bank Building, Chicago, Illinois. 119. Belt Conveyors. — The Robins belt conveyor, Fig. 104, consists essentially of a thick belt of the required width driven by suitable pulleys and carried upon idlers so arranged that the belt becomes trough-shaped in cross section. The belt is constructed of woven cotton duck covered with a special compound on the carry- ing side. The belt is thicker at the middle than at the edges, since the wear is greatest in a line along the center. The idlers are carried by iron or wooden framework, and are spaced from 3 feet to 6 feet between centers on the troughing side according to the width of belt and the weight of the load. On the return side these distances range from 8 to. 12 feet. Ow Fig. 104. Guide Pulleys, Robins Belt Conveyor. COAL AND ASH-HANDLING APPARATUS 193 High-speed rotary brushes with interchangeable steel bristles prevent wet, sticky material from clinging to the belt. Automatic tripping devices placed at the proper points cause the material to be discharged where it is needed. The trippers consist essentially of two pulleys, one above and slightly in advance of the other, the belt running over the upper and under the lower one, the course of the belt resembling the letter S. The material is discharged into chutes on the first down- ward turn of the belt. The trippers may be movable or fixed, single or in series. Movable trippers are used when it is desired to discharge the load evenly along the entire length, as, for instance, in a continuous row of bins, while fixed trippers are employed where the load is to be discharged at certain and somewhat separated points. The movable^ trippers are made in two forms, " hand-driven " and "automatic." In the former they are moved from point to point by means of a hand crank. The " automatic " tripper is propelled by the conveying belt through the medium of gearing. • It reverses its direction automatically at either end of the run, and travels back and forth continuously dis- tributing its load. It can be stopped, reversed, or made stationary at will. The most notable installations of this system are at the 96th Street station and the Kingsbridge station of the Metropolitan Street Railway Company, New York City. 120. Elevating Tower, Hand-Car Distribution. — Fig. 105 illus- trates the coal and ash-handling installation at the Aurora and Elgin Interurban Railroad power house, Batavia, 111. Coal is delivered to the plant by railroad cars which dump directly into coal hoppers located inside a steel structure running the entire length of the building and spanned by two railroad tracks. There are 18 hoppers constructed of 17-inch brick walls fitted with steel -plate bottoms. Subdividing the storage space in this manner makes it possible to carry different grades of coal, prevents the spreading of fire, and affords a simple con- struction for the support of the railroad tracks. The basement of the boiler room extends underneath the hoppers, and two lines of narrow- gauge tracks are imbedded in the concrete floor. Turntables at the center facilitate the switching of cars to the elevators which rise through the boiler room close to the chimney. The cars, of one ton capacity each, are of special construction, with roller-bearing axles and a com- bined ratchet lift and friction dump. The filled cars are pushed from underneath the hoppers to two elevators which lift them to the line of tracks supported overhead across the boiler fronts. They are then pushed to the hoppers suspended above the boiler setting and the coal is dumped. These hoppers have a capacity of six tons each. From the hoppers the coal is fed to the stoker by an ordinary down spout. 194 STEAM POWER PLANT ENGINEERING The ashes fall from the stokers into an ash pit, from which they may be discharged into ash cars. The ash cars are elevated to a set of tracks running at right angles to the main tracks, and are transferred to ash bins located directly over the coal bins. Coal and ashes are weighed GRADE TRACK TO ELEVATOR Fig. 105. Coal and Ash-Handling System at the Power House of the Aurora and Elgin Interurban Railway, Batavia, 111. in the small cars. There are ten boilers in this plant and four men are required to handle the coal and ashes. The entire coal and ash-handling system cost about $10,000, and. the cost of handling the coal and ashes is approximately 4 cents per ton. This does not include wages of firemen or water tenders. COAL AND ASH-HANDLING APPARATUS 195 121. Overhead Storage, Bucket Hoist. — Fig. 106 gives a general view of the coal-handling plant of the Depot Street power house of the Cincinnati Traction Company. This installation is a good example of an application of the " overhead storage gravity feed " system to an existing plant without interfering in any way with its operation. The system consists essentially of a receiving pit below the car tracks from which the coal is hoisted to a series of overhead bins. The coal storage is outside the boiler house in an independent structure. The bins are of steel framework with concrete floors, and are sufficiently elevated Fig. PIT CAPACITY SO TONS Coal and Ash-Handling System at the Depot Street Power House of the Cincinnati Traction Company. to spout coal easily to the stoker magazine. The total capacity of the overhead bins is about 1,600 tons. The four bins or receiving pits have a capacity of 50 tons each, or approximately one car load, and are so situated that all four may be filled simultaneously without shifting the train. The coal-handling apparatus consists of a one-ton self-filling bucket operated on a three-motor electric crane running on rails at the top of the storage bins. The coal is hoisted from the receiving pit through suitable shafts in the bin structure and dumped into the over- head hoppers. The maximum capacity of the hoist is 50 tons per hour. The labor required to handle the coal from car to bins is performed by one man working five hours per day and an assistant engaged a small part of the time to dump cars, clean hoppers, etc. The average daily coal consumption is approximately 200 tons. The total cost of the 196 STEAM POWER PLANT ENGINEERING equipment was about $18,000 for the bins complete and $4,500 for the coal-handling crane. The cost of handling the coal and ashes is approxi- mately 1.5 cents per ton of coal. Including all charges fixed and operat- ing the total cost of handling the coal is about 3.5 cents per ton. This does not include wages of firemen or water tenders. 122. Elevating Tower, Cable-Car Distribution. — The coal and ash- handling system of the new turbine power plant of the Detroit Edison Company, Fig. 107, is a typical example of a large station equipped with elevating tower and cable-car distributers instead of the usual bucket conveyor. The system consists essentially of a lofty steel tower in which are housed at various levels a track receiving hopper, crushing rolls and screens, weighing hopper, hoisting apparatus, etc., and a small cable railway for delivery to the bunkers. The railroad coal cars enter the tower on an elevated trestle 18 feet above grade, below which is a track receiving hopper. A two-ton " tub hoist " is filled with coal from the bottom of the receiving hopper and elevated to a 20-ton bin at the top, 120 feet above ground level. This bin has a grille bottom at one side and under the outlet a heavy duty coal crusher, thus allowing the fine coal to screen through directly while all the larger lumps are automatically -delivered to the crusher. The hopper beneath this delivers to the revolving screen, which sorts the slack into one bin below and the nut coal into the other. From the two bins the small cable cars are filled for dumping into the desired bunkers over the boiler rooms. The cars are arranged for automatic dumping by means of adjustable trips which may be located at any point. The object of separating the nut coal and slack is to burn the latter during light or medium loads, keeping the former for heavy loads and " peak " overloads. The down spouts are double, with a valve in each branch operated from the floor, so that either grade of fuel may be drawn out at any time and in any proportion desired. The entire system has a capacity of from 50 to 75 tons of coal per hour and is driven by steam engines, with the exception of the revolving screen which is motor driven. The ash-handling system consists of brick-lined concrete hop- pers underneath each pair of stokers which discharge their contents by gravity into the small cars operated on the track system in the boiler-house basement. When handling 275 tons per day of 24 hours the cost of operation is approximately 12.5 cents per ton from coal car to ash car, including wages of firemen and water tenders. 123. "Vacuum" Ash Conveyor. — Fig. 108 gives a diagrammatic arrangement of a recently patented ash-conveying system depending upon the velocity of a column of air for moving the ashes. The system COAL AND ASH-HANDLING APPARATUS 197 ^Os. RECEIVING BIN Fig. 107. Coal and Ash-Handling System at the Power House of the Detroit Edison Company, 198 STEAM POWER PLANT ENGINEERING is simple in operation and low in first cost. One end of special cast- iron header F leads to the ash pits of the various boilers by means of branch tubes, and the other end is connected with a sealed separating chamber A. Each branch pipe is fitted with simple circular openings directly underneath each ash-pit door for admitting ashes and which are kept covered except when in operation. Exhauster E creates a partial vacuum in chamber A and draws in air at a high velocity from the opening in the ends of the branch pipes. Ashes raked into the Fig. 108. Diagrammatic Arrangement of the "Vacuum" Ash-Handling System. pipes through the openings are caught by the rapidly moving column of air and forced into chamber A. The ashes fall to the bottom and are fed into the main ash pit by a slowly revolving ash valve B. Air and dust are withdrawn from the top of the separator chamber through pipe G and discharged to the stack or to waste. A spray is introduced into pipe F to reduce dust. The process is a continuous one, and the ashes may be completely removed from the ash bin without interfering with the operation of the exhauster. In a later construction the ash COAL AND ASH-HANDLING APPARATUS 199 200 STEAM POWER PLANT ENGINEERING COAL AND ASH-HANDLING APPARATUS 201 bin and separating chamber are included in one chamber, thus doing away with the revolving ash valve and the small motor operating it. In this latter design the bin is never completely empty, a certain depth of ashes being maintained to seal the bottom at all times. At the Armour Glue Works, Chicago, 111., this system is applied to a boiler plant of thirteen boilers, aggregating 4,800 horse power, and cost, completely installed, $5,600. As originally installed the separating chamber had a volume of about 35 cubic feet and the suction intake was placed 58 feet above the ash-pit level. The revolving ash valve made about 13 r.p.m., and was driven by a one-horse-power motor. In the present installation the separating chamber and motor-operated ash valve are dispensed with and the discharge pipes lead directly into the main ash bin, which has a capacity of 60,000 pounds of wet ashes and is constructed of five-sixteenths -inch sheet iron. The exhauster (a 30-foot Root blower) has a capacity of about 8,000 cubic feet per minute at 265 r.p.m., and is driven by a 75-horse-power motor. Under normal conditions of operation the motor requires 50 horse power when deliver- ing 250 pounds of ash per minute, and the vacuum on the suction side of the exhauster is 3.3 inches of mercury. The pipe from the ash bins to the separating chamber is 10 inches in diameter and is constructed of extra heavy chilled cast-iron pipe. The piping from the separating chamber to exhauster and to stack is 22 inches in diameter and is con- structed of number 16 and number 20 galvanized iron. The ashes are raked by hand from the ash pits to the suction openings of the branch pipes, and are handled dry, the dust being taken along with the ashes. Elbows are soon worn out by the abrasive action of the ashes, and tees are used instead, since the accumulation in the " dead " end receives the impact and takes up the' wear. The cost of handling the ashes in this installation is approximately 7 cents per ton. 124. Cost of Handling Coal and Ashes. — In large stations where a number of men are employed to handle coal and ashes only it is a simple matter to divide the cost of handling into the various stages, thus : 1. Cost of unloading cars or barges. 2. Cost of conveying coal to bunkers. 3. Cost of feeding coal to furnace. 4. Cost of removing ashes. These costs are usually expressed in cents or dollars per ton of coal burned, or in terms of cents or dollars per horse power hour or kilo- watt hour of main prime mover output. Item number 3 is oftentimes included under " boiler-room attendance " and items 1, 3, and 4 under " coal and ash handling." Not infrequently all four items are included under " attendance." So much depends upon the character of stokers 202 STEAM POWER PLANT ENGINEERING and furnace, size of boilers, and the like, that general figures on the cost of handling the coal and ashes are of little value unless accompanied by a description of the equipment. For the sake of general comparison the most satisfactory method of expressing the cost is in dollars per ton of coal from coal car to ash car. This includes wages of coal and ash passers, repair men, and boiler tenders. In small stations the coal and ash handling is done by the boiler tenders, in which case it is impracticable to separate the items mentioned above, and the cost is ordinarily included under attendance. An average figure for handling coal by barrow and shovel is not far from 1.6 cents per ton per yard up to the distance of five yards, then about 0.1 cent per ton per yard for each additional yard. With automatic conveyors the operating cost, not including wages of firemen and water tenders, varies with the size of plant and the type of conveyor, and ranges anywhere from a fraction of a cent per ton to four or five cents per ton. The larger the plant and the greater the amount of coal handled the lower will be the cost per ton. In comparing the relative costs of manual and automatic handling, fixed charges of at least 15 per cent of the first cost of the mechanical equipment should be charged against the latter in addition to the cost of operation. In large central stations equipped with stokers and conveyors and consuming 200 tons or more of coal in twenty-four hours, the cost of handling the coal from coal car to ash car, including wages of firemen and water tenders, will range between 10 cents and 18 cents a ton. 125. Coal Hoppers. — Fig. 109 shows a front and side elevation of a typical set of stationary weighing hoppers as applied to the boilers of the Quincy Point power plant of the Old Colony Street Railway Company, Quincy Point, Mass. Each battery of boilers is provided with an independent set of hoppers. The bottoms of the overhead coal bunkers lead into the small hoppers A, A. The operation of any single weighing hopper is as follows: Coal is fed from the overhead bunkers to weighing hopper H by means of valve V. The weight of coal in the weighing hopper is transmitted by a system of levers and knife edges to the inclosed scale beam / and noted in the usual way. The weighed charge of coal is then admitted to the down spout S by means of valves similar to those at V. Although separate weighing hoppers for each battery, as illustrated in Fig. 109, offer many advantages, they are quite costly and it is not unusual to install one or more large weighing hoppers mounted on overhead traveling carriages so that one may supply a number of boilers (Fig. 110). At the Armour Glue Works, Chicago, the coal supply is stored in one large overhead bunker of 1000 tons capacity. A five- COAL AND ASH-HANDLING APPARATUS 203 Fig. 109. Stationary Coal Weighing Hoppers. Fig. 110. Traveling Coal Hoppers. 204 STEAM POWER PLANT ENGINEERING ton motor-driven traveling hopper receives its supply from this central bunker and delivers it to the various boilers. One man operates the Fig. 111. Common Slide Coal Valve. Fig. 112. Simplex Coal Valve. traveling hopper, tends to the coal valves, and supplies all boilers with coal. Weighing hoppers are sometimes made automatic; that is, the opening and closing of valves, feeding of coal, and recording of weight are auto- matically performed by the weight of the coal itself. The scale is set for discharges of a certain weight and continues to discharge this amount auto- matically. In the few plants which are equipped with auto- matic weighing hoppers the capacity of the hopper is approximately 100 pounds per discharge. These hoppers are necessarily more complicated and more costly than the ordinary weighing hoppers, and it is a question whether the advantages offset the extra first cost and main- tenance charges. A small automatic hopper of 100 pounds discharge capacity costs approximately $400 as against $250 for the ordinary weighing device. Fig. 113. Duplex Coal Valve. COAL AND ASH-HANDLING APPARATUS 205 126. Coal Valves. — Figs. Ill to 115 illustrate the principles of a few well-known coal valves. They may be conveniently grouped into two classes according to the location of the coal pocket: (1) those drawing the coal from overhead bunkers and (2) those drawing from the side of a bin. In the first class come the simple slide valve, the simplex and duplex rotating valve. In the latter are the flap valve and the rotating valve. They are made in various sizes and designs, but those illustrated are examples of the most common types. The simple slide valve, Fig. Ill, is applicable only to small size coal and to small spouts, since coarse or lump coal may get in the way and prevent proper closing. The simplex valve, Fig. 112, consists of a rotating jaw actuated by a lever. There are no rubbing surfaces, and the jaws cut through the material without jamming. The duplex Fig. 114. Common Coal Valve. Flap Fig. 115. " Seaton " Coal Valve. valve, Fig. 113, consists of two rotating jaws con- nected to a common actuating lever. The jaws move simultane- ously, so that even a par- tially open valve delivers the coal centrally. When closing the valve the flow is gradually stopped by the decreasing width of the opening and there is but little resistance to The largest valve can easily be operated the movement of the jaws, by hand. The flap valve, Fig. 114, is the simplest form for drawing coal from a side bin. It consists merely of an iron flap hinged to the bottom of 206 STEAM POWER PLANT ENGINEERING the chute. The valve is lowered to let the coal run over its top and is raised to stop the flow. It cannot be clogged or get jammed in closing. The flap is raised and lowered by a simple lever. For very large bins, where the valves are to be opened and closed frequently, the " Seaton " valve, Fig. 115, is usually preferred. This valve consists of two jaws EE', and TT' pivoted to suitable framework at and actuated by lever A. The valve is shown fully closed. Raising lever A causes the cut-off blade EE / to rotate about and permits the coal to flow through the space between the edge of the jaw E and the end of the chute. The rate of flow is regulated by the width of this opening. The cut-off blade does not reach a stop, hence there is no possibility of a lump of coal getting in the way and preventing the prompt closing of the valve. Coal and Ash-Handling Installations : Commonwealth Edison, Chicago, Power, Dec, 1906, p. 718. Boston Elevated, Elec. World, Sept. 7, 1901, p. 396. Inter- borough Rapid Transit Co., New York, Elec. World, Feb. 4, 1905, p. 264; Engr., U.S., May 15, 1904, p. 337; Eng. News, Jan. 14, 1904, p. 41. Waterside Station, New York, Edison Co., Eng. Rec, Sept. 9, 1905, p. 287. Detroit Edison Co., Eng. Rec, Oct., 1905, p. 396. Brooklyn Rapid Transit Co., St. Ry. Jour., Sept. 23, 1905, p. 435. N. Y. C. and H. R. R., St. Ry. Jour., Nov. 11, 1905, p. 876. Aurora, Elgin and Chicago Ry., Eng. Rec, Feb. 7, 1903, p. 153. Missouri River Power Co., Eng. News, Oct. 19, 1905, p. 403. Brooklyn Edison Co., Gold St. Station, Elec World, June 15, 1907. Hoisting and Conveying Machinery: Pro. A.S.M.E., June, 1908. CHAPTER VII. CHIMNEYS. 127. Chimney Draft. — Draft produced by a chimney depends upon so many conditions and involves such a large number of variables that empirical methods of proportioning, based upon actual performances, are more to be relied upon than theoretical calculations. Draft is due to the difference in the weight of the column of hot light gases in the stack and that of the cooler and heavier surrounding atmos- phere, the latter tending to flow into the base and thereby force the lighter gases out the top of the stack. The commonly accepted theory of chimney draft is based upon Peclet's hypothesis that the flow through the furnace flues and chimney may be represented by the equation h== £Tl( l+ G + — )> < 27 > 64.4 \ m ) in which h = the head of fluid producing the flow, feet. u = velocity of the gases in the chimney, feet per second. G = a coefficient to represent the resistance to the passage of air through the coal. I = total length of the path of the gases, feet. m — area of cross section divided by the perimeter. / = a coefficient depending upon the nature of the surfaces over which the gases pass. From experiments on chimneys and boilers Peclet gives in connection with this theory the following values of coefficients G and / : G - 12, / = 0.012, on the basis of 20 to 24 pounds of coal burned per square foot of grate surface per hour. On account of the variation in practice of the factors u, f, and G and the difficulty of determining them engineers prefer to use the modified formulas given further on. The difference of pressure, or intensity of draft may be expressed theoretically, ignoring friction, as follows : 207 208 STEAM POWER PLANT ENGINEERING Let H — height of chimney in feet. T = absolute temperature of the freezing point, degrees F. T x — absolute temperature of the gases in the chimney. T 2 = absolute temperature of the outside air. P = average atmospheric pressure. P 2 = observed atmospheric pressure. W = weight of a cubic foot of air at 32 degrees F. and pressure P. W x = weight of a cubic foot of chimney gas at 32 degrees F. and pressure P. Then the weight of a cubic foot of hot gas in the chimney will be W * %■'■£[ (28) and the weight of a cubic foot of cold air outside will be w 7? • # ■ (29) The weight of a column of hot gas H feet high and one foot square will be W X H^. |L. (30) Similarly the weight of the cold-air column will be WH ^-Y 2 (31) and the difference in pressure or the intensity of draft will be where D is in pounds per square foot. By making P = P 2 = 14.7, T = 493, W = 0.0807, W. x = 0.084, and D x = pressure in inches of water (D t = 0.192 D), equation (32) assumes the familiar form JW( 7 -f-?f). 03) By assuming W = W x = 0.081 and P = 14.7, equation (32) may be written A=0.52tfP 2 (i-I-) (34) This latter form is ordinarily used where the atmospheric pressure differs considerably from that at sea level, as at high altitudes. Table 23 gives the density of air and chimney gases at various temperatures. CHIMNEYS 209 Example : Required the maximum theoretical draft obtainable from a chimney 150 feet high, atmospheric pressure 14.7 pounds per square inch, temperature outside air 60 degrees F., temperature chimney gases 550 degrees F. Here H = 150, T 2 = 461 + 60 = 521, T x = 461 + 550 =1011. Substituting these values in equation (34), D x = 150 (— - — ) = 1.02 inches of water, which is about 25 per cent greater than the draft actually obtained, and represents the maximum possible under the given conditions, neglecting the resistance offered by the chimney and the pressure TABLE 23. DENSITY AND SPECIFIC VOLUME OF AIR AND CHIMNEY GASES AT VARIOUS TEMPERATURES. Air. Chimney Gases. t 5 V d t d t d t d 11.581 .935 .086353 200 .06334 430 .04695 660 .03730 5 11.706 .945 .085424 210 .06239 440 .04643 670 .03697 10 11.832 .955 .084513 220 .06147 450 .04592 680 .03665 15 11.931 .965 .083623 230 .06058 460 .04542 690 .03633 20 12.085 .976 .082750 240 .05971 470 .04493 700 .03602 25 12.211 .986 .081895 250 .05887 480 .04445 710 .03571 30 12.337 .996 .081058 260 .05805 490 .04398 720 .03540 32 12.387 1.000 .080728 270 .05726 500 .04353 730 .03511 35 12.463 1.006 .080238 280 .05648 510 .04308 740 .03481 40 12.589 1.016 .079434 290 .05573 520 .04264 750 .03453 45 12.715 1.026 .078646 300 .05499 530 .04221 760 .03424 50 12.841 1.037 .077874 310 .05428 540 .04178 770 .03396 55 12.967 1.047 .077117 320 .05358 550 .04137 780 .03369 60 13.093 1.057 .076374 330 .05290 560 .04096 790 .03342 62 13.144 1.061 .076081 340 .05224 570 .04056 800 .03316 65 13.220 1.067 .075645 350 .05159 580 .04017 900 .03072 70 13.346 1.077 .074930 360 .05096 590 .03979 1000 .02861 75 13.472 1.087 .074229 370 .05035 600 .03942 1100 .02678 80 13.598 1.098 .073541 380 .04975 610 .03905 1200 .02516 85 13.724 1.108 .072865 390 .04916 620 .03869 1300 .02373 90 13.851 1.118 .072201 400 .04859 630 .03833 1400 .02245 95 13.976 1.128 .071550 410 .04803 640 .03798 1500 .02131 100 14.102 1.138 .070910 420 .04749 650 .03764 1800 .01848 110 14.354 1.159 .069665 2000 .01698 d = density, pounds per cubic foot. t = temperature, degrees F. s = specific volume, cubic feet per pound. v = comparative volume, volume at 32° = 1. Density of chimney gas taken 0.085 pound per cubic foot at 32° F. and 29.92 inches of mercury. (Rankine, " Steam Engine," gives the density at 32° F. as varying from 0.084 to 0.087.) 210 STEAM POWER PLANT ENGINEERING TABLE 24. THEORETICAL DRAFT PRESSURE IN INCHES OF WATER. CHIMNEY 100 FEET HIGH. 1 Temp. Temperature of the External Air - - Barometer, 14.7 Pounds per Square Inch. 2 in the Chim- ney. 0° 10° 20° 30° 40° 50° 60° 70° 80° 90° 100° 200 .453 .419 .384 .353 .321 .292 .263 .234 .209 .182 .157 220 .488 .453 .419 .388 .355 .326 .298 .269 .244 .217 .192 240 .520 .488 .451 .421 .388 .359 .330 .301 .276 .250 .225 260 .555 .528 .484 .453 .420 .392 .363 .334 .309 .282 .257 280 .584 .549 .515 .482 .451 .422 .394 .365 .340 .313 .288 300 .611 .576 .541 .511 .478 .449 .420 .392 .367 .340 .315 320 .637 .603 .568 .538 .505 .476 .447 .419 .394 .367 .342 340 .662 .638 .593 .563 .530 .501 .472 .443 .419 .392 .367 360 .687 .653 .618 .588 .555 .526 .497 .468 .444 .417 .392 380 .710 .676 .641 .611 .578 .549 .520 .492 .467 .440 .415 400 .732 .697 .662 .632 .598 .570 .541 .513 .488 .461 .436 420 .753 .718 .684 .653 ..620 .591 .563 .534 .509 .482 .457 440 .774 .739 .705 .674 .641 .612 .584 .555 .530 .503 .478 460 .793 .758 .724 .694 .660 .632 .603 .574 .549 .522 .497 480 .810 .776 .741 .710 .678 .649 .620 .591 .566 .540 .515 500 .829 .791 .760 .730 .697 .669 .639 .610 .586 .559- .534 550 .863 .828 .795 .762 .731 .700 .671 .644 .618 .593 .585 600 .908 .873 .839 .807 .776 .746 .717 .690 .663 .638 .613 1. For any other height multiply the tabular figure by ^-r t where H is the height in feet. p. 2. For any other pressure multiply the tabular figure by , where P is the barometric pres- sure in pounds per square inch. required to impart velocity to the gases. Table 24 has been computed from formula (34), and gives the maximum theoretical draft in a chim- ney 100 feet high for different flue-gas temperatures. The intensity of draft required to produce best results depends upon the kind and condition of fuel, the thickness of fire, character of grate, and resistance of the breeching, tubes, baffles, dampers, etc. As stated above, the loss of draft in the chimney proper approximates 20 per cent of the total, that in the breeching is taken as 0.1 inch per 100 feet of flue, and 0.05 inch for each right-angle bend; the loss in the boiler varies from 0.3 to 0.6 inch, depending upon the type;* the loss in the furnace varies between wide limits, and depends upon the kind of fuel and the rate of combustion. The curves in Fig. 116 compiled by the Stirling Company and published in their book " Stirling'' give the furnace drafts necessary to burn various kinds of fuels at different combustion rates, and give an idea of the influence of the character of the fuel and the rate of combustion. * Specific figures may be obtained from the manufacturers. CHIMNEYS 211 jl f|| |^:S ^P : M lljp isi:ij| . 1 Ik n\\\\\\\\\± 44U-N ff + yT fi-!-r iTF'^xi SEE 1 1 1 1 (j iii||r ±ff5gffi '•tNfW fnTTT^^ffi 111 ^ffi ^gl Sjg^^ II |ll || ilfig r^k l^i ^IPsil g j?7 ffitflr^S ffijVteSl |||J| fr^f 2 fc (baiVM jo 63H0ND uid hsv QNY aovNianj N33Mxaa a3aino3a xjvaa jo souod 212 STEAM POWER PLANT ENGINEERING Example : Determine the probable draft necessary to burn 30 pounds bituminous run of mine per hour per square foot of grate when the out- side air is 60 degrees F., the temperature of the chimney gases 550 degrees, and the flue is 100 feet long, with two right-angle bends. The losses will be divided approximately as follows : Inch. Loss in furnace (from curves in Fig. 116) 0. 17 Loss in boiler (average) 0. 40 Loss in flue, 100 feet at 0.10 per 100 . 10 Loss in turns, 2 X 0.05 .10 6777 Since the loss in the chimney alone approximates 20 per cent of the total, 0.77 -T- .80 = 0.96 will be the theoretical draft necessary. From equation (33), Substituting for the given values of D lf T lf and T 2 in above equation, °- 96 - H fer ~ ion] ' From which, H = 142, height of stack necessary to produce a draft of 0.17 inch in the furnace. Table 25 gives the results of a test of a 100-foot unlined steel chimney, showing the variation in draft at different points in the stack. Theory of Chimney Draft : Power, Oct., 1896, p. 18, Dec, 1898, p. 20, March, 1906, Feb., 1900, p. 12; Engr. U.S., Jan. 15, 1903, May 15, 1902, p. 313; Trans. A.S.M.E., 11-451, 762, 772, 974, 984; Elec. Rev., Lond., Oct. 14, 1904. 128. Chimney Formulas. — Rational methods of determining the height and area of chimneys being cumbersome and unwieldy and of doubtful value for practical use, the various empirical formulas outlined in Table 26 are quite commonly used. They give good results within the limits of the assumptions upon which they are based, but otherwise may lead to absurd results, their applicability depending largely upon the available data covering the various losses with the particular kind, quality, and condition of coal, and conditions of operation. Occasionally practical and local considerations fix the height of the stack irrespective of theoretical deductions. The logical procedure is to determine first the height of chimney necessary to produce the draft at the desired maximum rate of combustion, and then to proportion the area by such formulas as (2), (4), or (5), to suit the quantity of fuel to be burned. CHIMNEYS 213 The following heights have been found to give good results in plants of moderate size: Feet. With free-burning bituminous coal 80 With anthracite, medium and large sizes 100 With slow-burning bituminous 120 With anthracite pea 130 With anthracite buckwheat 150 With anthracite slack 175 TABLE 25. CHIMNEY DRAFT. Test of a 100-Foot Unlined Steel Chimney 3 Feet in Diameter at Massachusetts Institute of Technology. (Peabody & Miller, "Steam Boilers," p. 121.) Over the grate At the bridge wall Half-way between bridge and back end of boiler At the back end of boiler In uptake near boiler In stack 34 feet above grate In stack 51 feet above grate In stack 68 feet above grate In stack 85 feet above grate Draft, Inches of Water. Maximum. Minimum 24 382 0.410 0.354 0.572 0.440 0.334 0.216 0.122 0.218 0.372 0.374 0.334 0.543 0.414 0.312 0.168 0.086 Temperature, Fah- renheit. Maximum. Minimum 403 396 380 370 345 389 374 368 354 314 The chimney serves two 80-horse-power boilers. During test one was banked and the combustion at the grate of the working boiler was 19.8 pounds per square foot of grate surface per hour. Coal burned per hour 590 pounds. For plants of 800 horse power or more the height of stack should never be less than 150 feet, regardless of the kind of coal used. Referring to Table 26, formulas (1), (2), (6), (7), and (9) are based upon a fuel consumption of 13 to 15 pounds of anthracite and 22 to 26 pounds of bituminous coal per square foot of grate area per hour. In formulas (3), (4) and (9), the diameter is dependent solely upon the quantity of coal burned per hour and the height is determined mainly by the rate of combustion per square foot of grate. The results accord well with practice. With western coals formula (3) gives results rather too large and the constant should be 120 instead of 180. Formula (5) is 214 STEAM POWER PLANT ENGINEERING perhaps the most used and has met with much approval. It is based on the assumptions that 1. The draft of the chimney varies as the square root of the height. 2. The retardation of the ascending gases by friction may be con- sidered due to a diminution of the area of the chimney or to a lining of the chimney by a layer of gas which has no velocity and the thickness of which is assumed to be 2 inches. Thus, for square chimneys, E=&-^- -A-\V1, (35) and for round chimneys, E =l( V -*-£) =^- 0-591 x/A (36) For simplifying calculations the coefficient of V A may be taken as 0.6 for both square and round chimneys, and the formula becomes E = A- 0.6 VI (37) 3. The horse-power capacity varies as the effective area E. 4. A chimney should be proportioned so as to be capable of giving sufficient draft to permit the boiler to develop much more than its rated power in case of emergencies or to permit the combustion of 5 pounds of fuel per rated horse power per hour. 5. Since the power of the chimney varies directly as the effective area E and as the square root of the height H, the formula for horse power for a given size of chimney will take the form H.P. = CE Vlf, (38) in which C is a constant, found by Mr. Kent to be 3.33, obtained by plotting the results from numerous examples in practice. The formula then assumes the form H.P. = 3.33 eVh ■ (39) H.P. = 3.33 (A - 0.6 VA) VW, (40) or from which H ««£)■: m Table 27 has been computed from equation 5, Table 26. Many engineers simply adopt the following proportions : Internal area of chimney at top, one-seventh grate area for bitumi- nous coal. 1* I&3 ft Oh ft > ft, > O CO Tfl « O o II II II II "< 1 ' ^ ^ ft} I, <£ o„ o- C ^ * C <3 ? ^ fti^J &q > ft3 6 QQ a CO 03 "0 J3 <: Q i— i CM co ^ G G CD CD a « -3 M '«_ 1 o ^ ^ (M ^1 S ^"|^ C ^ 2" » CO o ^ 5 |l Cft O M,? 1 l o s *— • >" II tCJ tq I&3 > 1*3 > o Up > .5 QO CO oo 525 co CD A CD bfl CO += c CO 03 - Oh CO CD >> £ 0Q &,,§ cn <3 ta «> rjS • *« cs . r MM!*, ■e © ©*« ti fi b c L" » O C8 r-.E-rt ersec mney . cbin rnala gas; a.-- £ *> © .9 a flS » © a> 3 © rH 3 3 © -fcS ?8 o o3 oj © Lb. fuel Ratio gi Weight Temper Temper ii H sdErfKK* >> © a a .d CJ ^ m"S 11 •s ^1 » ••«£ ed sta stack gas; t per °.8 p-. o S^ 2 "3 ©,©. a— C flW-rf ^ is for u for li ature tical gbt. oo n S w •- •£ . ^ag*** — S.a ^ II S H to *3 «^ o o 13? 3 . a O - «H "* fl © ^2 s 8 •~a (H © -h t-i cm cm cm co co co co -^ ■<*< to «o <© *>. t~- oo oo os os o o »-h cm W s ^J o o CO ^ o to +» o co o W5 lO 00 -^ CM CO t^CO^lO O HiOHO) OS CM Tt4 »0 O CO CO O CO t~ CM i— I i— I cm" (n"^^^ co^'io'co"' io o c© co o oo c© »o «o oo coco HNioa oiooi-H lO 00 rtTflNO "<*< 00 CO CO i-i h c^Tc^Tcn'co' co'co'rt^io" CO »C CO 00 OO OS OS OONU5H lO OO CO O OS O CO OO lO lO CO CM ■«* t~- © CM CO OS CM CO tJ4 CO i-Tt-Ti-T cN^Cm"^^ Co'cO*-^"^ t-i i-4 © t>. CONOS-H O 00 O CO 00 00 O CO OS CO »0 t- C*0(N OS i-4 "^ CO OOrHTJIN tH^MO i-Ti-h".-^ i-T^CnTcM* CO^Co'-^ia" lO 00 00")Oh © N © CM OtOOlH OS -*f i-h O i-h CO N(NO© © CM CM © ION ffi-HMiO NOCCiO OS CN^OS^t^ i-Ti^i-h" i-TcN^CN^CsT CN^CO^Co""^ CO CO i— i CM OS CO CM OO OS CO © OS «5©NM t-i CM »0 OS -<*4 CN i— I i— I MNMO 00 00 CO lO CO-**4»OCO OOOCNrt* CO 00 i— I CO CO OS CO CO i-Ti-Ti-h" i-h^i-TcnTc^T ^ OS © t— I i— I t— ( r— I t— i CM ©HO) i— i iO (N CO OOCOQOt)4 »0 OS CM 1>- CO !>• OS CM N CO t-I t-i t-i CM (NCOThlO N00ON OS OS CM OS 00 00O«5 Tf4 lO T-HT+tOOi-l lO"*Tt4CO OS CO t-4 i-i t-h CM CM CO Tf4 i© CO00 CO00 CO-fCOOO lOONO cooo »-i-«cMt^o •* co cm co HHHIN CM CO "tf »0 OS ""tf CM CO N CO CO CO t-H t-i cq^cooo © co co os co t-i 1-4 t-4 i-l i-l CM CO NrHOOOO © U0 CM CO CO W*«5N © CM »© 00 i-4 T— I T— I T-H 1—4 CM V/k90-V=3 ■B8JV GAIlOaga ""Eh ^CQ 5l t-4 CN O CO CO »o OS CO CO t^ O0 O0 t>- CN i-H "^4 O0 N- U0 CM HNN© O0'^4 CO o 00 iO O i-l i-l CM CO CO ■*«5NM OS CM lO OS CO 00 CO 00 CM CM CO CO x*4 O CO CO ^lOiOCO o oo »o CO NNCfiH CHIMNEYS 21 7 Internal area of chimney at top, one-ninth grate area for anthracite coal. Example : Determine the area and diameter of a stack for a 2000- horse-power plant to operate under the following conditions: Rated load 2000 horse power; maximum overload 40 per cent of rated; flue 150 feet long, with one right- angle bend; average rate of combustion 20 pounds of bituminous coal per square foot of grate surface per hour; atmospheric temperature 60 degrees F.; flue-gas temperature at over- load 600 degrees F.; coal burned per boiler horse power, 4 pounds. With modern types of steam engines or turbines an overload of 40 per cent has little effect on the economy of the prime mover, and the boiler efficiency is but slightly reduced, but an additional allowance of 25 per cent should be made in estimating the overload combustion rate. The maximum rate of combustion then will be pounds per square foot of grate surface per hour. The draft required at the point where the flue enters the chimney, considering the various losses, will be found as follows : Inch. Furnace (see curves. Fig. 116) 0.3 Boiler ..0.4 Flue, 150 feet at 0.1 inch per 100 feet 0. 15 Turns, 1 at 0.05 0.05 CUT From formula (33), •>-' nor less than J inch for larger sizes. CHIMNEYS 221 i Orf%AT£ .&?ICH ■ SACKING FLOOft LEV£l Fig. 117. Steel Chimney at the Power House of the South Side Elevated Railroad, Chicago. 222 STEAM POWER PLANT ENGINEERING It is customary to make the courses about 5 feet in height for con- venience in erection. Table 29 gives the dimensions of self-supporting steel stacks as made by the Riter Conley Company of Pittsburg, who use the following empirical formula in determining the thickness of the shell 8 t M 0.8ZV (43) in which S x = stress per lineal inch of section considered, M = wind moment in inch-pounds, and Dj = diameter of the shaft in inches. Allowing 8000 pounds per square inch as the safe stress for single- riveted joints and 10,000 for double-riveted joints, the required thick- ness is found by dividing S ± by 8000 or 10,000. Example : Determine the thickness of plate at a section 150 feet from the top of a cylindrical steel stack 12 feet in diameter and 200 feet high. Horizontal seams to be double riveted. The total wind pressure on the section is 150 X 12 X 25 = 45,000 pounds. The moment arm is -HP X 12 = 900 inches. D x = 144 inches; S = 8000 pounds per square inch. TABLE 29. STEEL STACKS. — SIZES OF RITER CONLEY COMPANY, PITTSBURG. Diameter of Flue. Total Height. Total Weight. How Made. Ft. In. 5 6 7 8 6 10 12 11 6 12 Ft. 165 160 150 200 200 225 255 Lb. 67,000 79,000 94,000 150,000 175,000 232,000 256,000 40 ft. of & in., 45 ft. of J in., 50 ft. of & in., 30 ft. of 1 in. 30 f t. of & in., 50 ft. of J in., 50 ft. of & in., 30 ft. of 1 in. 60 ft. of \ in., 60 ft. of & in., 30 ft. of f in. 90 ft. of \ in., 60 ft. of & in., 50 ft. of f in. 35 ft. of I in., 35 ft. of & in., 35 ft. of & in., 35 ft. of H in., 35 ft. of f in., 25 ft. of $f in. 40 ft. of \ in., 40 ft. of & in., 40 ft. of & in., 40 ft. of H in., 40 ft. of f in., 25 ft. of ^ in. 75 ft. of | in., 65 ft. of & in., 55 ft. of f in., 35 ft: of ^ in., 25 ft. of | in. CHIMNEYS 223 Substituting these values in equation (42), 45,000 X 1800 _ 8000 x ZJA /144 4 - DA 32 V 144 D a = 143.36. Now t = Dl ~ D » 2 = 144 - 143.36 2 = 0.32 inch. The nearest commercial size lies between nine thirty-seconds and five- sixteenths. The Riter Conley formula gives for this section s = M 45,000 X 900 1 0.8 D* 0.8 X 144 2 = 2440 pounds. , S, 2440 QnK . r t = T^h. = ^^ = -305 inch. 8000 8000 134. Riveting. — The diameter of rivets should always be greater than the thickness of the plate but never less than one-half inch. The pitch should be approximately 2\ times the diameter of the rivet, and always less than 16 times the thickness of the plate. Single-riveted joints are ordinarily used on all sections except the base, where the joint should be double riveted with rivets staggered, although in very large stacks all horizontal seams are double riveted to give greater stiffness to the shaft. 135. Stability of Steel Chimneys. — The wind being ordinarily the only force tending to overturn the stack, and the chimney being rigidly bolted to the foundation, a condition of stability requires that (W c + Wf) — be equal to or greater than P t — - + h ). (44) in which W c = weight of the chimney in pounds. Wf = weight of the foundation. P = total wind pressure in pounds. D, H, and h, in feet, as indicated in the figure. 224 STEAM POWER PLANT ENGINEERING Expressed graphically: Lay off GP, Fig. 118, equal to the total wind pressure in direction and amount and acting at the center of pressure of the shaft; lay off GW equal to the weight of the stack and foundation; find the resultant GR and produce it to intersect the base line as at R'; if R / falls within the inner third of the base the stack is stable, provided, of course, that the chimney is properly designed and constructed. Therefore the heavier the combined weight of the chimney and its foundation the more stable the structure. (See also para- graph 140.) D in Fig. 118 varies from one-tenth to one-fifteenth H, depending upon the character of the subsoil. For the ordinary concrete foundation, Christie ("Chimney Design and Theory," p. 57) gives as an average value for D H 2 d D = 26,000 + 10. (46) Fig. 118. Detailed Description of Steel Chimneys : Eng. Rec, Feb. 15, 1902, p. 146, July 5, 1902, p. 2, April 23, 1904, p. 523, Sept. 10, 1904, p. 314; Power, Dec, 1905, p. 716, April, 1905, p. 231, Jan., 1902, p. 6; Engr. U.S., Sept. 1, 1904, p. 591, June 15, 1905, p. 403; Eng. News, July 20, 1905, p. 64. 136. Brick Chimneys. — By far the greater number of power-plant chimneys are of brick construction and usually of circular section, though octagonal, hexagonal, and square sections are quite common. The round chimney requires the least weight for stability, and the others in the order mentioned. Taking the total wind pressure on the flat surface of a square stack as unity, the effective pressure for the same projected area will be 0.75 for the hexagonal, 0.6 for the octagonal, and 0.5 for the round. Brick chimneys may be divided into two general classes : 1. Single shell, Fig. 119, and 2. Double shell, Fig. 120. The double shell is the most common and consists of an outer shaft of brickwork and an inner core or lining extending part way or throughout the entire length of the shaft. CHIMNEYS 225 h-*VH k*fH H I & /6-S" &> 1 e& - /ff-S' ; 2 tor-* - /6*-S m 3 i *3'-~ ^ /S'-S* 4 i IS'h- /G'~S" 5 ■ „- «. /6'-S" s s 6 k F TOP OF FOUNDATIONAL! TOTAL HEIGHT ABOVE FOUNDATION 2.00FF SECTION ON A-A SECTION ON BS Fig. 119. Custodis Radial Brick Chimney. 226 STEAM POWER PLANT ENGINEERING BBB BOB Fig. 119a. Custodis Radial Brick. The single shell is the general construction where carefully burned and selected brick not easily affected by the heat are used. As the inner core or lining is independent of the outer shell and has no part in the strength of the chimney, the rules for determining the thickness of the walls are practically the same for both single and double shell. 137. Thickness of Walls. — The thickness of the wall should be such as to require minimum weight of material for the proper degree of stability, due consideration being paid to the practical requirements of construction. The thickness does not vary uniformly, but decreases from bottom to top by a series of steps or courses as in Fig. 121. In general, the thickness at any section should be such that the resultant stress of wind and weight of shaft will not put the masonry in tension on the windward side or in excessive compression on the leeward side. For circular chimneys using common red brick for the outer shell the following approximate method gives results in conformity with average practice: * = 4 + 0.05 d + 0.0005 H, (47) where t = thickness in inches of the upper course, neglecting ornamenta- tion, and should, of course, be made equal to the nearest dimension of the brick in use. Ordinary red bricks measure 8^x4x2. d = clear inside diameter at the top, inches. H= height of stack, inches. Beginning at the top with this thickness, add one-half brick, or 4 inches, for each 25 or 30 feet from the top downwards, using a batter of 1 in 30 to 1 in 36. The minimum value of t for stacks built with inside scaffolding should be 7 inches for radial brick and 8^ inches for common brick, as a thinner wall will not support the scaffold. Radial brick for chim- neys are made in several sizes, so that the thickness of the walls when they are used increases by about 2 inches at the offsets. For specially molded radial brick or for circular shells reenforced as in Fig. 120 the length of the different courses may be much less than CHIMNEYS 227 Fig. 120. Brick Chimney at the Power Plant of the Armour Institute of Technology. 228 STEAM POWER PLANT ENGINEERING stated above. The external form of the top is a matter of appearance, and may be designed to suit the taste, but should be protected by a cast- iron or tile cap and provided with lightning rods. Ladders for reach- ing the top of the chimney are generally located inside of brick stacks and outside of steel ones. Professor Lang's rule (Eng. Rec, July 20, 1901, p. 53) for determining the length of the different courses is (Fig. 121) - o ( 2(X+6(H + 0.1056G 0.453 p - 18 ,), + 2.5 1 + 656 tan a - 0.007 H (48) in which h = length of the course under consideration. C = constant = 1 for a circular, 0.97 for an octagonal, and 0.83 for a square chimney. i = increase in thickness for each succeeding section in feet. G = weight per cubic foot of brickwork. p = wind pressure, pounds per square foot. a = angle of the internal batter. All other notations as indicated in Fig. 121. For chimneys over 100 feet in height he recommends that 100 be used instead of the actual height, since the critical point will be in one of the lower sections and not at the base. If a value of h is obtained which is not con- tained an even number of times in H, it may be slightly increased or decreased so as to effect this result. To determine the stresses at any section the shaft is treated as a cantilever uniformly loaded with a maximum wind pressure of 25 pounds per square foot. If the tension on the windward side subtracted from the compression leaves a positive remainder, the chimney will be stable; if the remainder is negative, the masonry will be in tension, which it withstands but feebly. The sum of the com- pressive stresses on the leeward side due to wind pressure and weight must be less than the crushing strength of the masonry. The practice, however, of assuming a fixed value for allowable pressure irrespective of the height of the stack gives dimensions that are too low for small CHIMNEYS 229 stacks and too high for large stacks. According to Professor Lang, compressive stress on the leeward side in pounds per square inch with single chimneys should not exceed p = 71 + 0.65 L, (49) where p = pressure in pounds per square inch. L = distance in feet from top of chimney to the section in question. With double shell p = 85 + 0.65 L. (50) The tension on the windward side should not exceed, For single shell: p = (18.5 + 0.056 L). (51) For double shell: p = (21.3 + 0.056 L). (52) Example: Determine the maximum stress in the outer fibers of the brickwork at the base of section 8 of the chimney illustrated in Fig. 119 when the wind is blowing 100 miles an hour. Assume the weight of the brickwork 120 pounds per cubic foot. A wind velocity of 100 miles per hour is estimated to exert a pressure of 50 pounds per square foot on a flat surface and approximately 25 pounds per square foot of projected area on a cylindrical surface. The height of the chimney to section 8 is 131.4 feet. The projected area as computed from the figure is 1800 square feet. Hence p, the total wind pressure, is 1800 X 25 = 45,000 pounds. The volume of brick- work above section 9 may be calculated, and is 6150 cubic feet, hence the weight W = 6150 X 120 = 738,000 pounds. The area of the joint at this section is 75.3 square feet, therefore the pressure due to the weight of the superimposed brickwork is 738,000 divided by 75.3 = 9800 pounds per square foot. To find the stress due to the wind pressure, substitute the proper values in equation (42) : Ph = S I =0.0983 ( Dl *Z D4 ) S. Here P — 45,000 as computed above. h = 55 feet. (Found by laying out the section and locating the center of gravity.) D x = 16.2. D = 12.9. Whence 45,000 X 55 = 0.0983 16 ' 2 * ~ * 2 ' 9 * S. 16.2 From which S = 9907 pounds per square foot . 230 STEAM POWER PLANT ENGINEERING The net stress on any part of the section is the resultant of that due to the weight of the stack and that caused by the wind, the net stress on the windward side being 9907— 9800 =107 pounds per square foot, which is evidently a tensile stress and should never exceed the value given by formula (51): p = (18.5 + 0.056 L) = (18.5 + 0.056 X 131.4) = 25.8 pounds per square inch = 3715 pounds per square foot. The net compressive stress on the leeward side is 9800 + 9907 = 19,707 pounds per square foot, which should not exceed that given by formula (49) : p = 71 + 0.65 L = 71 + 0.65 X 131.4 = 156.4 pounds per square inch = 22,521 pounds per square foot. (See also analysis of steel-concrete chimney, paragraph 142.) 138. Core and Lining. — The core or lining of a brick chimney is commonly carried to the top of the shaft, though it sometimes extends only part of the distance. The inside diameter is generally uniform, the offsets being made on the outside. The core and outer shell should be independent to prevent injury due to expansion of the core. The rules for the thickness of lining in steel chimneys apply also to brick chimneys. The batters for the inner and outer shells should be such as to allow at least 2 inches clearance between the two shafts at the top, and the top should be protected by an iron ring or by a projecting ledge from the outer shell. 139. Materials for Brick Chimneys. — Brick for the external shaft should be hard burned, of high specific gravity, and laid with lime mortar strengthened with cement. Lime mortar itself is more resist- ant to heat, but hardens slowly and may cause distortion in newly erected stacks, and hence should be used only when a long time is taken in building. Mortar of cement and sand alone is not to be recommended, since it does not resist heat well and is attacked by carbon dioxide, particularly in the presence of moisture. A mortar consisting of 1 part by volume of cement, 2 of lime, and 6 of sand CHIMNEYS 231 may be used for the upper brickwork, 1, 2\, and 8 respectively for the lower part, and 1,1, and 4 respectively for the cap. The harder the brick the more cement is necessary, as lime does not cling so well to hard, smooth surfaces. The inner core may be constructed of second-class fire brick, since the temperature seldom exceeds 600 degrees F. Lime mortar is invariably used for the core. 140. Stability of Brick Chimneys. — When there is no wind blowing and the chimney is built sym- metrically about a vertical axis the pressure due to weight is uniformly distributed over the x bearing surfaces, and the center Z 1 w r *- D 7 A (A) of pressure lies in the line XX, Fig. 122. But when the wind blows the pressure exerted tends to tilt the shaft as a whole- column in the direction of the current, and the pressure de- creases from the windward to the leeward side of the base, until, with a sufficiently high velocity of wind, it may become zero, in which case the center of pressure moves a distance q FlG - 122 - towards the leeward side of the base. As soon as the pressure at A becomes zero the joint begins to open (assuming no adhesion between chimney and base) and the shaft is evidently in the condition of least stability. The distance q through which the center of pressure has moved is called the radius of the statical moment. For any column it may be shown that (B) q = -j- . (Rankine, " Applied Mechanics," p. 229.) (53) Ae in which / = moment of inertia of the section. A = area of the section. e = distance from the center of the shaft to the outer edge of the joint. D Thus for circular section, For square section } a 8 D 9 "T 232 STEAM POWER PLANT ENGINEERING D 2 + d 2 For annular circular ring, q = For hollow square, q = 8D D 2 + d 2 6D The relationship between weight of shaft and wind pressure for the condition of least stability is Ph = Wq, (54) in which P = total wind pressure, pounds. h = distance in feet from the base line of the section under con- sideration to center of gravity of that section. W = weight of shaft in pounds above the assumed base line. q = radius of the statical moment. The condition of least stability for round chimneys requires, there- fore, that D 2 4- rl 2 Pk = W 8D (55) For many purposes it is sufficiently accurate to assume D = d, and equation (55) becomes Ph = W — for round chimneys. (56) Ph = W — for square chimneys. (57) o The rule commonly used in Germany, and which is finding much favor with engineers in the United States, gives for the condition of least stability w(^R +jA= Ph. (Eng. Rec., July 27, 1901, p. 82.) (58) Notations as in Fig. 122, all dimensions in feet. This permits of a lighter chimney than equation (55), and the maxi- mum wind pressure may be assumed to put the joint on the wind- ward side in tension or even to permit a slight opening of same. A rule of thumb for stability is to make the diameter of the base one- tenth of the height for a round chimney; for any other shape to make the diameter of the inscribed circle of the base one-tenth of the height. The factor of stability is the quotient obtained by dividing the value of q from formula (54) by that from (53). If less than unity, the chimney is in tension at the outer fiber on the windward side, and must be redesigned unless the tension is less than that allowed by equation (51). Calculations for stability should be made for various sections. CHIMNEYS 233 Example: Analyze the chimney illustrated in Fig. 119 for stability at say section 8, the following data referring to the portion above the base line of this section. From the drawing : Projected area of the stack, 1800 square feet. Volume of brickwork, 6150 cubic feet. Outside diameter of base, 16.2 feet. Inside diameter of base, 12.9 feet. Center of pressure to base line, 55 feet. Total height above base line, 131.4 feet. Maximum total wind pressure : P = 1800 X 25 = 45,000 pounds. Weight of shaft : W = 6150 X 120 = 738,000 pounds. For stability, according to equation (55), Substituting the proper values : Ph = 45,000 X 55_=_2,475,000 foot-pounds. W D * + d ' = 738,000 I 16 - 2 ' + J 2 ' 92 ) = 2,441,000. 8 D ' \ 8X 16.2 / While Ph is slightly greater than W — — — — , for practical purposes 8 D the shaft at this section would be called stable under maximum allow- able wind pressure. For stability, according to equation (58), Ph < W(iR + Jr). Ph = 2,475,000, as determined above. 6\45\ 4 ) = 4,177,000. Ph is therefore considerably less than W(i R + J r), and the con- dition imposed in equation (58) is more than fulfilled. Detailed Description of Brick Chimneys : At Paris Exposition, Eng. Rec, Feb. 17, 1900, p. 155; Boston Elevated Ry., Eng. Rec, Dec. 22, 1900, p. 593; Plymouth Cordage Co., Eng. Rec, May 18, 1901, p. 466; Boston, Edison Co., Eng. Rec, Oct. 10, 1903, p. 438; Smelting Works, Freiburg, Germany, Power, Aug., 1900, p. 13; Ter- minal R.R. Assn. of St. Louis, Mo., Cassier's Mag., Jan., 1898, p. 261; Cambridge Electric Light Co., Engr. U.S., May 15, 1904, p. 331; Interborough Rapid Transit Co., Engr. U.S., Nov. 1, 1904, p. 737; Metropolitan St. Ry. Co., Power, March, 1899, p. 1. W(i R + J r) = 738,000 ( y + 234 STEAM POWER PLANT ENGINEERING 141. Custodis Radial Brick Chimney. — Fig. 119 gives the details of a 200 x 10 foot radial brick chimney constructed of special molded radial brick, formed to suit the circular and radial lines of each section, thus permitting them to be laid with thin, even mortar joints. The blocks are much larger than common brick and the number of joints is proportionately reduced. They are molded with vertical perforations, as shown in Fig. 119a, which permits thorough burning, thereby increas- ing the density and strength and at the same time reducing the weight of the block. In laying, the mortar is worked into the perforations about one-half inch. The first 60 feet above the base are octagonal in section, with 36-inch walls, and the balance of circular section, with walls tapering gradually from 22 inches to 7 J inches in thickness. A radial brick lining extends 60 feet from the base as indicated. The chimney was designed to furnish draft for a 3500-horse-power boiler plant and cost, erected, $8,800. The entire weight of the chimney exclusive of foundation is 870 tons. Radial brick chimneys without the inner lining are likely to be unduly affected by heat. The tallest chimney in the world (1907), located at Great Falls, Mont., is of the Custodis type, and is used for leading off the gases from the smelter plant of the Boston and Montana Consolidated Copper and Silver Mining Company. The height above the top of the foundation is 506 feet, and the internal diameter at the top 50 feet. The chimney and foundation cost approximately $200,000. Custodis Chimney Details: Eng. Rec, Oct. 1, 1904, p. 385; Power, May, 1900, p. 12. 142. Steel-Concrete Chimneys. — The use of concrete reenforced with iron or steel for the construction of chimneys is rapidly increasing. The advantages claimed for this class of stack are: 1. Light weight of the whole structure, being but one-third as great as an equivalent common brick chimney. The space occupied is much less than with either brick or steel stack, on account of the thinness of walls at the base and the absence of any flare or bell. 2. Total absence of joints, the entire structure including foundation being a monolith. 3. Great resisting power against tension and compression. 4. Rapidity of construction. May be erected at an average rate of six feet per day. 5. Adaptability of the material to any form. This type of chimney being comparatively new, little data concern- ing depreciation are available, but some which have been in use ten years show little or no deterioration. CHIMNEYS 235 GRADE Fig. 123. Weber Reenforced Concrete Chimney. 236 STEAM POWER PLANT ENGINEERING Fig. 123 gives the details of a Weber steel-concrete chimney erected at Portland, Oregon, for the Portland General Electric Company. The entire structure, foundation and shaft, is a. monolith, 238 feet in total height and 12 feet internal diameter, weighing only 889 tons. It occupies but 168 square feet of ground space at the grade level. The weight not including foundation is 470 tons. The stack was erected complete in 58 working days, and cost approximately $13,000. The cement used was German Portland mixed with select bank sand in proportion of one to three, gravel or crushed stone being used only in the foundation below the ground. The mortar was used medium dry and tamped in the form around the steel reenforcement. The shaft is of the double-shell type, with inner core extending 70 feet above the grade. The core is but 4 inches in thickness at the base, and the outer shell 8 inches. Both inner and outer shell are reenforced with vertical T bars, lJxlJx T \ inch, of low-carbon Bessemer steel, spaced at the base 24 inches between centers in the inner core and 4 inches in the outer shell, and increasing in spacing to the top, where the dis- tance between the bars is 12 inches. The horizontal rings are 1 x 1 x J T's spaced 18 inches between centers in the core and 36 inches in the outer shell. The steel bars vary from 16 to 30 feet in length, and where they meet lengthwise are lapped not less than 24 inches. The use of differ- ent lengths of steel prevents the laps from concentrating in any given section. The tallest chimney of this type (1907) was erected for the Butte Reduction Works at Butte, Mont. Its height is 350 feet and inside Fig. 124. diameter 18 feet. The following strain sheet gives the Weber Company's analysis of the chimney illustrated in Fig. 123, and is based on a wind pressure of 50 pounds per square foot. Notations as in Fig. 124. 19.75 Sq.Ft. CHIMNEYS 237 Weights. Wf = weight of foundation = (k 2 K + l * + l * h\ 150 = /30 2 - 2 + 3 ° 2 + 15 ' 3^ 150 = 523,200 pounds. 150 = weight per cubic foot of concrete. W e = earth weight on foundation = {lih 6 — (volume of foundation)} 100 = (7200 cubic feet - 3995 cubic feet) 100 = 320,500 pounds. 100 = weight per cubic foot of earth. W = weight of shaft = {A t (h A + h 3 ) + A 2 (h, + h a ) + AJi 5 } 150 = {38.5 (72 +3) + 13 (72 + 3) + 19.75-158} 150 = 934,950 pounds. W t = total weight = W f + We + W = 1,778,650 pounds (889 tons). Section at Grade G r . I. Wi = weight of outer shell and single shell above section = (AJit + A 3 hr ) 150 = (28.5 • 72 + 19.75 • 158) 150 = 775,806 pounds. II. r = radius of statical moment 14.66 = 3.35 feet. III. P = wind pressure on chimney = 14.66 X 72 X 25 + 13 X 158 X 25 = 77,738 pounds. 238 STEAM POWER PLANT ENGINEERING M = wind moment on section 5( 2 : ^ + (/)Af)^ + |) 14.66 X 72 X y X y +( 13 x 158 X |°-)( 72 + 158> 2 > = 8,703,818 foot-pounds. IV. N = statical moment = rWi. = 3.35 X 775,806 = 2,598,950 foot-pounds. V. B = bending moment =-- M -N = 8,703,818 - 2,598,950 = 6,104,868 foot-pounds. VI. — = section modulus e = 0.0982 ( D *~ dl *) = o 0982 ( U M X 12 * ~ 13 - 33 X 12 * \ \ 14.66 / = 169,703. VII. z = tension per square inch sectional area = 12B ^ - e = 12 X 6,104,868 -^ 169,703 = 432.5 pounds. VIII. Z = total tension = 144 A,z = 144 X 28.5 X 432.5 = 1,825,015 pounds. IX. s = area steel required = — • (a = sectional strain on steel) a = 16,000 pounds per square inch = 114.2 square inches. CHIMNEYS 239 X. K = number of bars = — ■ (x = 0.45 square inches = area of one bar) x = 252 bars. For Stability. XI. L = length of one side of base. = 8,703,818 6 1,778,650 = 29.4 feet. Section 42' 0" above Grade. I. W t = 596,250 pounds. II. r = 3.35 feet. III. P = 62,295 pounds. IV. N = 1,997,438 foot-pc V. B = 3,763,889 foot-pc VI. I e = 169,703. VII. z = 222 pounds. VIII. Z = 911,088 pounds. IX. s = 57 square inches, X. K = 127 bars. M= 5,761,325 foot-pounds. Section at Offset. I. W t = 468,000 pounds. II. r = 3 feet. III. P = 51,350 pounds. M= 4,056,650 foot-pounds. IV. N = 1,404,000 foot-pounds. V. B = 2,652,650 foot-pounds. VI. I = 102,041. e VII. z = 311 pounds. VIII. Z = 786,000 pounds. IX. s =55.5 square inches. X. K = 123 bars. 240 STEAM POWER PLANT ENGINEERING Section 50' 0" from Top. I. W t = 148,125 pounds. II. r = 3 feet. III. P = 16,250 pounds. M= 406,250 foot-pounds. IV. N = 444,365 foot-pounds. Since the statical moment N is greater than the wind moment M, there is no bending moment B, so no steel is required, the chimney above this section standing of its own weight. However, thirty-two bars are continued to the top. Detailed Description of Concrete-Steel Chimneys : Weber Chimney, Power, July, 1907, p. 476; Burt Portland Cement Co., Eng. News, Dec. 29, 1904, p. 579; Leiter Coal Mines, Eng. Rec, March 21, 1904, p. 661, Power, Dec, 1904, p. 787; Pacific Electric Ry. Co., Los Angeles, Cal., Eng. News, April 2, 1903, p. 308; Laclede Fire Brick Co., Eng. News, April 2, 1903, p. 310; Tall Concrete Chimneys, Eng. News, July 20, 1905, p. 57, Aug. 3, 1905, p. 120, Feb. 15, 1906, p. 165, Oct. 11, 1906, p. 387; Butte Reduction Works, Eng. Rec, Feb. 3, 1906, p. 124; New Types, Eng. Rec, Dec. 15, 1906, p. 670. 143. Breeching. — The area of the flue or breeching leading from the boilers to the chimney is generally made equal to or a little larger than the internal area of the chimney, 20 per cent greater being an average figure. The flue may be carried over the boilers or back of the setting, or even under the fire-room floor, but in any case should be as short as possible and free from abrupt turns. Short right-angled turns reduce the draft approximately 0.05 inch for each turn, and a convenient rule is to allow 0.1 inch loss for each 100 feet of flue if of circular cross section and constructed of steel, and double this amount for brick flues of square section. The cross section of the flue need not be the same throughout its entire length, but may be tapered and proportioned to the number of boilers. Where two flues enter the stack on opposite sides, a diaphragm is inserted as indicated in Fig. 119. Flues should be covered with heat-insulating material. 144. Chimney Foundations. — On account of the concentration of weight on a small area the foundation of a chimney should be carefully designed. In most cities the building laws limit the maximum loads allowed for various soils and materials, and although they vary con- siderably the average is approximately as follows: Material. Safe Load, Lb. per Sq. Ft. Hard-burned brick masonry, cement mortar, 1 to 2 20,000-30,000 Hard-burned brick masonry, cement mortar, 1 to 4 18,000-24,000 Hard-burned brick masonry, lime mortar 10,000-16,000 Concrete, 1 to 8 8,000-10,000 CHIMNEYS 241 Kind of Soil. Safe Load, Tons per Sq. Ft. Quicksands and marshy soils 0.5 Soft wet clay 1.0 Clay and sand 15 feet or more in thickness 1.5 Pure clay 15 feet or more in thickness 2.0 Pure dry sand 15 feet or more in thickness 2.0 Firm dry loam or clay 3.0- 4.0 Gravel well packed and confined 6.0- 8.0 Rock broken but well compacted 10.0-15.0 Solid bed rock Up to -g- of its ultimate crushing strength. Tons per Pile. Piles in made ground 2.0 Piles driven to rock or hardpan 25.0 Chimney foundations as a rule are constructed of concrete except where the low sustaining nature of the soil necessitates the use of piles or a grillage of timber or steel. For masonry chimneys the foundation is designed to give the necessary support to the shaft without particular reference to its mass or distribution, as the shape of the foundation has virtually no effect on its stability as a column. In steel and reenforced concrete chimneys the shape and weight of the foundation are a function of the desired factor of stability, since the shaft is securely anchored to the foundation and the two form practically one mass. The founda- tion should be designed to fulfill the conditions in formula (46) in addition to the requirements for mere support. Table 30 gives the least diameter and depth of foundation for steel chimneys of various diameters and heights. 145. Chimney Efficiencies. — The chimney as a mover of air has a very low thermodynamic efficiency. Compared with that of a fan its performance is very poor, and mechanical-draft concerns sometimes use this as an argument. Example: A chimney 200 feet high and 10 feet in diameter furnishes draft for a battery of boilers rated at 3500 horse power. Average outside temperature 60 degrees F. ; temperature of flue gases 500 degrees F.; calorific value of the fuel 14,000 B.T.U. per pound. Compare the thermal efficiency of the chimney as a mover of air with that of a forced draft apparatus of equivalent capacity. From Table 24 we find that a chimney 200 feet high, with tempera- tures as stated above, will furnish a theoretical draft of 1.27 inches, equivalent to a pressure of 6.6 pounds per square foot. Neglecting friction the height H of a column of external air which would produce this pressure is H=(^- r ^h, (59) 242 STEAM POWER PLANT ENGINEERING in which h = height of the chimney in feet. d = density of the hot gases in the stack. d 1 = density of the outside air. Substitute in (59), d, = 0.0763, d = 0.0435, and h = 200. TABLE 30. SIZES OF FOUNDATION FOR STEEL CHIMNEYS. Diameter, Feet. Height, Feet. Least Diameter of Foundation. Least Depth of Foundation. 3 100 15' 9" 6' 0" 4 100 16' 4" 6' 0" 4 125 18' 5" 7' 0" 5 150 20' 4* 9' 0" 5 200 23' 8" 10' 0" 6 150 21' 10" 8' 0" 6 200 25' 0" 10' 0" 7 150 22' r 9' 0" 7 250 29' 8" 12' 0" 9 150 23' 8" 10' 0" 9 275 33' 6" 12' 0" 11 250 24' 8" 10' 0" 11 350 36' 0" 14' 0" H = * 0763 0.0763 / 85.9 feet. The theoretical velocity of the air entering the base of the chimney under this head is v = V2gh = V2 X 32.2 X 85.9 = 74.5 feet per second. The weight of the gas escaping per second = 74.5 X area of the stack X 0.0763 = 446 pounds. The displacement of this volume of gas is the result of heating it from 60 to 500 degrees F. Taking the specific heat of the gas as 0.2375, the heat necessary to displace 456 pounds per second is Heat required = 446 X 0.2375 X (500 - 60) = 46,500 B.T.U. per second CHIMNEYS 243 The work actually performed is that of overcoming a total resistance of 6.6 X 78.5 = 518 pounds (78.5 = internal area of the chimney) through a space of 74.5 feet; i.e., Work done = 74.5 X 518 = 38,591 foot-pounds per second, = 49.7 B.T.U. per second. Efficiency = 4 = .00107, or about T V of 1 per cent. If a fan be substituted for the chimney and we allow say 8 per cent for the efficiency of engine and boiler, 40 per cent for the fan, and 25 per cent for friction, the combined efficiency will be 0.08 X 0.40 X 0.75 = 0.024, or 2.4 per cent. 024 The fan then will be — '■ = 22.4 times more efficient than the 0.00107 chimney as a mover of air. 146. Cost of Chimneys. — Christie ("Chimney Design and Theory") gives the following costs of chimneys 150 feet high and 8 feet internal diameter : Common red brick approximate cost $8,500.00 Radial brick do do 6,800.00 Steel, self-supporting, full lined do do 8,300.00 Steel, self-supporting, half lined do do 7,800.00 Steel, self-supporting, unlined do do 5,820.00 Steel, guyed do do 4,000.00 The following approximate costs of various sizes of a well-known radial brick chimney give an idea of the variation in cost due to in- crease in diameter and height : Size of Chimney. Size of Chimney. Cost. Cost. Height. Diameter. Height. Diameter. Feet. Feet. Feet. Feet. 75 4 $1,350.00 175 8 $7,050.00 75 6 1,950.00 175 10 7,925.00 75 8 2,650.00 175 12 8,950.00 75 10 3,725.00 175 14 9,725.00 125 6 3,500.00 200 8 9,250.00 125 8 4,250.00 200 10 10,500.00 125 10 4,675.00 200 12 11,100.00 125 12 5,125.00 200 14 12,500.00 150 8 6,150.00 250 10 16,500.00 150 10 7,125.00 250 12 18,250.00 150 12 7,750.00 250 14 21,500.00 150 14 8,275.00 250 16 24,250.00 244 STEAM POWER PLANT ENGINEERING TABLE 31. PROPORTIONS OF CHIMNEYS FOR FACTORY STEAM BOILERS, COLLECTED FROM PRACTICE. (Hutton.) Height of Internal Dimensions. Ratio of Thickness of Walls. Chimney above the Ground in Feet. Bottom to Top. Internal Area. Size of Base at the Ground Line. Siz* . of Top. Thickness at Base in Inches at Ground Thickness at the Top in Inches. Line. 40 2' 6" 1' 9" sq. 2.04 18 9 60 2' 11" 2' 0"sq. 2.12 18 9 70 3' 4" 2' 3"sq. 2.13 23 9 80 3' 8" 2' 6" sq. 2.18 28 9 90 4' 0* 2' 9" sq. 2.27 28 9 100 4' 8" 3' 0" diam. 2.40 28 9 110 4' 10" 3' 3" diam. 2.33 28 9 120 5' 6" 3' 6" diam. 2.40 28 9 135 6' 0" 4' 0" diam. 2.30 28 9 150 4' 6" 3' 0" diam. 2.25 28 14 155 6' 0* 4' 6" diam. 1.78 56 14 160 9' 0" 5' 0" sq. 3.24 36 14 170 V 6" 5' 0" diam. 2.25 36 14 180 6' 4" 4' 6" diam. 2.00 54 14 200 5' 3" 3' 6" diam. 2.28 36 14 225 16' 0" 6' 6"sq. 4.00 36 14 250 19' 0" 13' 0" diam. 2.13 40 14 300 14' 0" 9' 0" diam. 2.42 48 14 450 21' 6* 10' 2" diam. 4.35 59 14 CHAPTER VIII. MECHANICAL DRAFT. 147. General. — The intensity of natural draft in a chimney depends mainly upon the height of the stack and the temperature of the chim- ney gases, and the chimney should be designed to meet the maximum requirements, permitting the damper to be partly shut at times. There is usually no practicable means of increasing the draft after the maxi- mum has been reached. Again, chimney draft is peculiarly susceptible to atmospheric influence and may be seriously impaired bv adverse winds and air currents. Notwithstanding these apparent limitations, by far the greater number of steam power plants depend upon chim- neys for draft, and for obvious reasons as will be discussed later. In many cases artificial draft has a great advantage and under certain conditions is indispensable; it is very flexible and readily adjusted to effect various rates of combustion, irrespective of climatic influ- ences, and permits any degree of overload without undue expenditure of energy. Artificial draft may be broadly classified under two heads 1. The vacuum or induced draft and 2. The plenum or forced-draft method. In the former a partial vacuum is produced above the fire by suitable apparatus, and the effect is substantially that of natural draft. In the forced-draft system pressure is produced in the ash pit, the air being forced through the grate. In both systems the artificial draft is usually produced by either 1. Steam jets or 2. Centrifugal fans or exhausters. 148. Steam Jets. — Fig. 125 shows an application of a ring jet to the base of a stack. The apparatus is very simple, inexpensive, and easily applied. It consists essentially of a ring or a series of concen- tric rings of 1-inch or lj-inch pipe, perforated on the upper side with T V or J- inch holes, and placed in the base of the stack, so that the jets are discharged upward, thus creating a draft independent of the temperature of the flue gases. The steam connection to the jet is generally made direct to the boiler and not to the steam main, though the jet is often produced by exhaust steam. 245 246 STEAM POWER PLANT ENGINEERING Fig. 126 illustrates a Bloomsburg jet, which involves to some extent the principle of the ejector. The increase in draft produced by these devices as ordinarily installed Fig. 125. Ring Steam Jet. Fig. 126. Bloomsburg Jet. is not great, although in locomotive practice where the entire exhaust is discharged up the stack an intense draft is obtained. Fig. 127 shows the application of a " McClaves argand blower." Fig. 127. McClaves Argand Blower. The steam is discharged below the grate through a perforated hollow ring, as indicated, drawing the air through the funnel by inspiration. This creates a powerful draft by forming an air pressure in the ash pit, MECHANICAL DRAFT 247 and is an especially useful system of forcing fires for boilers which need forcing for short periods only. Steam jets are very uneconomical, since a large amount of steam is required to produce good results. Table 32, based on experiments at the New York Navy Yard to determine the best form of steam jet for producing draft in launch boilers, shows steam consumptions of 8.3 to 21.2 per cent of the total steam made. Table 33 gives the steam consumption of a number of types of steam jet blowers as determined by A. J. Whitham. The best performance is 4.6 per cent and the poorest 11.1 per cent of the total boiler steam generated. Steam jets below the grate are said to prevent clinker from forming where fine anthracite coals are used, and thus to assist in keeping the fire free and open. Steam jets arranged above the grate and discharging either from the side walls, front wall, or bridge wall, oftentimes assist complete com- bustion by stirring up the volatile gases and air and insuring a thorough mixture, thus affording one of the simplest and frequently a very efficient means of furthering smokeless combustion. The action is of course purely mechanical, the steam in itself not being a sup- porter of combustion; hence if the air supply is deficient the steam jet is of no avail unless arranged to carry sufficient air along with the steam. Fig. 128 illustrates such an application to a hollow bridge wall. The top of the wall is fitted with a small cast-iron column M, partially GRATE Fig. 128. Application of Steam Jets to Hollow Bridge Wall. imbedded in the brickwork. A series of 1-inch holes " 00" drilled near the top of the casting, furnish exits for the steam and air. A steam jet in one end of the column induces air into the iron chamber and forces it across the fire in fine streams. Excessive air dilution is avoided by partially closing the ash-pit doors and by regulating the intensity of the jets. An installation of this type is especially effective in connection with coal having a tendency to fuse and seal the air pas- 248 STEAM POWER PLANT ENGINEERING sages in the grate. Two Stirling boilers at the Armour Institute of Technology equipped with this device gave practically smokeless combustion at all normal loads, though at heavy overloads it was sometimes necessary to slightly open the fire doors. Without the use of the jets smoke could not be prevented even at light loads. Analysis of the flue gases showed but a slight decrease in the percentage of C0 2 . TABLE 32. RESULTS OF EXPERIMENTS UPON STEAM JETS AT NEW YORK NAVY YARD.* Pounds of Water Evaporated per Hour. Index of Jet. A B C D E In boiler making steam In boiler supplying jets Per cent of steam used bv iet 463.8 97.5 21.2 580.0 120 20.7 361.25 30 8.3 528.5 63.2 12.0 545.00 76.25 19.0 * Annual Report of the Chief of the Bureau of Steam Engineering, U. S. Navy, 1890. TABLE 33. CONSUMPTION OF STEAM BLASTS COMPARED, t Coal. Name of Blower. Per Cent of Air Openings in Grate. Pounds of Dry Coal burned per Hour per Square Foot of Grate. Per Cent of Total Steam Generated in the Boilers that is required to operate the Steam Blasts. Rice Young 11 11 7 11 11 26 11 11 7 7 25.8 17.9 27.0 27.3 16.7 31.4 16.4 26.1 32.5 45.4 11.1 Do ...do 7.0 Do Wilkinson Young ....do 10.8 Buckwheat Do 10.8 4.6 Do ...do 8.9 Do McClave ....do . 6.7 Do 9.3 Do Wilkinson ....do 7.8 Do 10.2 t Trans. A.S.M.E., Vol. XVII. — Whitham. 149. Parson Smokeless Furnace. — The Parson forced-draft system for smokeless combustion, applied to a return tubular boiler as illus- trated in Fig. 129, comprises a specially designed grate G, depending upon a steam jet blower A for draft. Part of the steam is admitted MECHANICAL DRAFT 249 below the grate and part over the fire through the hollow bridge wall H. The supply of air above the grate is regulated by means of damper F. The steam to blower A is automatically adjusted by regulator N, which is actuated by the steam pressure. The steam to the jet is superheated by passing the supply pipe through the setting as indi- Fig. 129. Parson's Smokeless Furnace: cated. The bridge wall H is provided with an extension platform M for holding the unburned fuel when cleaning the fire. 150. Heinrich Smokeless Furnace. — Figs. 130 and 131 show the appli- cation of the Heinrich system of forced draft to a return tubular boiler. Hot air is taken from the boiler room above the boilers by a steam jet blower at A and forced into the superheating chamber below the combustion chamber. From this chamber part of the air is drawn by the auxiliary blowers C and forced through tuyeres above the grate, the rest passing through an opening beneath the bridge wall into the ash pit and up through the bed of fuel. Steam for the blower A and the auxiliaries C is supplied through an automatic regulator R, which opens when the steam pressure falls below the required value. The manufacturers (Heinrich Manufacturing Company, Milwaukee, Wis.) sell this apparatus with a guarantee of 15 per cent saving in fuel over natural draft, common grate bars, and hand firing. 151. Fan Draft. — Fig. 132 shows a typical installation of a centrif- ugal fan on the forced-draft or plenum principle, the fan creating a pressure in the ash pit and forcing air through the fuel. The most approved method is to pass the air through the bridge wall, thence toward the front of the grate, though it may enter through an under- ground duct or through the side of the setting. Forced draft is usually adopted in old plants where increased demands for power require that the boilers be forced far above their rating to save the heavy expense 250 STEAM POWER PLANT ENGINEERING of new boilers, or in plants burning refuse, anthracite culm or screen- ings, which require an intense draft for efficient combustion. Forced draft is also well adapted for underfeed stokers of the retort type, hollow blast grates, and the closed fire hole system. The air supply may be taken from an air chamber built around the breeching, thereby supplying the heated air to the fan and effecting a lower temperature in the breeching and a higher temperature in the furnace. The objection is sometimes raised against forced draft that the gases tend Fig. 130. Heinrich Smokeless Furnace (Sectional Elevation). Fig. 131. Heinrich Smokeless Furnace (Sectional Plan). to pass outward through the fire door when the fire is cleaned or re- plenished, since the pressure in the furnace is greater than atmospheric. This objection may usually be overcome by suitable dampers in the blast pipe which are closed on opening the fire doors. With a boiler plant of 1000 horse power or more the cost of a forced-draft fan, engine, and stack will approximate 20 to 30 per cent of the outlay of an equivalent brick chimney. The power consumption will depend MECHANICAL DRAFT 251 upon the character and efficiency of the motor or engine and will range from 1 to 5 per cent of the total capacity. Induced draft as illustrated in Fig. 133 is perhaps the most com- mon substitute for natural draft and is extensively used in street rail- way and lighting plants which have high peak loads, being ordinarily installed in connection with fuel economizers. The suction side of the fan is connected with the uptake or breeching of the boiler or bat- teries of boilers and the products of combustion usually exhausted through a stub stack. The illustration shows a typical installation in which two fans of the duplex type are placed above the boiler setting. The fan ducts are generally designed with a by-pass direct to the stack to be used in case of accident or when mechanical draft is not required. Fig. 132. Typical Forced-Draft System. Since the fan handles hot gases it must, under the ordinary con- ditions of practice, have a capacity approximately double that of a forced-draft fan delivering cold air, but the gases being of lower density the power required per cubic foot moved is less. With forced draft about 300 cubic feet of air are required per pound of coal; with induced draft the fan must handle twice this volume if the gases are exhausted at 500 degrees F. or 450 cubic feet if exhausted at 300 degrees F., a temperature to be expected in connection with economizers, The advantages of induced draft over forced draft are very pro- nounced. The pressure in the furnace is less than atmospheric, there- fore it is not necessary to shut off the draft in cleaning fires or ash pit, and the fire burns more evenly over the entire grate area and requires less attention than with forced draft. An induced-draft plant costs considerably more than forced draft on account of the larger fan 252 STEAM POWER PLANT ENGINEERING required, but the operating expenses are but little greater. With a boiler plant of 1000 horse power or more the cost of a single induced- draft fan, engine, stack, etc., will approximate 40 to 50 per cent of the outlay required for a brick chimney of equivalent capacity, and the double-fan outfit will approximate 50 to 60 per cent. The double-fan Fig. 133. Typical Induced-Draft System. system is particularly adapted to plants which operate continuously and where even a temporary break-down is a serious inconvenience. Advantages of Mechanical Draft: Am. Elecn., June, 1898, p. 244, Feb., 1902, p. 63; Eng. Rev., Sept., 1901, p. 4; Eng. Mag., April, 1901, p. 81, March, 1900, p. 931; Elec. Rev., Lond., Feb. 3, 1899, p. 186; Cassier's Mag., Nov., 1898, p. 48, Jan., 1905, p. 252, March, 1906; St. Ry. Rev., July 15, 1901, p. 415; Engr. U.S., July 16, 1906, p. 475; Elec. Eng., Aug. 11, 1905, p. 193. Application of Mechanical Draft to Stationary Boilers : Power, Dec, 1900, p. 30; West. Elecn., Feb. 16, 1901, p. 118; Jour. West. Soc. Engrs., March 19, 1902, p. 271; St. Ry. Rev., July, 1899, p. 463; Cassier's Mag., Nov., 1898, p. 48; Elec. Rev., July 27, 1898, p. 52; Engr. U.S., Jan. 1, 1907. 152. Performance of Fans. — The first satisfactory theory of centrif- ugal fans was promulgated by Daniel Murgue in 1872. He proved MECHANICAL DRAFT 253 that theoretically the maximum pressure created by a perfect fan is equivalent to twice the head which would produce a velocity equal to that of the periphery. Thus tf =^, (60) in which H = maximum difference in pressure in feet of air, u = peripheral velocity in feet per second, and g = acceleration of gravity 32.2. A and B, Fig. 134, represent Pitot tubes inserted in the discharge pipe of a centrifugal blower, A being bent to face the current, while B is at right angles to it. A receives the full impact of the stream, and H JL /^\ cf (A) (B) ORIFICE CLOSED Fig. 134. the manometer indicates the total pressure, static and velocity, while B registers the static pressure only. With the discharge orifice closed, as in Fig. 134", the velocity becomes zero, and the water depression in both manometers will be the same, due to the static pressure, which, according to Murgue's theory, will be a maximum and, ignoring fric- u 2 tion or eddy currents = — • Example : Determine the maximum pressure, in inches of water, which a perfect fan would exert with discharge orifice closed; diameter of fan 6 feet; r.p.m. 318. The peripheral velocity is u = 2 nrn = 6.28 X 3 X 318 = 6000 feet per minute. = 100 feet per second. Substituting in Murgue's formula, u = 100 and g = 32.2, H = 155! = 310 feet, 32.2 254 STEAM POWER PLANT ENGINEERING i.e., the pressure created by the fan would be equivalent to the weight of a column of air 310 feet high, or, assuming an air temperature of 75 degrees F., an equivalent head in inches of water of 310X0.074495 144 X 0.0361 = 4.45 inches. (0.074495 = density of air at 75 degrees F. and 0.0361 = pressure pro- duced by one inch of water in pounds per square inch.) If the discharge orifice be opened to its maximum (Fig. 135) the static pressure indicated by manometer B becomes zero, since there is no /rS\ n /^n\ n %# a (A) %# (B) ORIFICE WIDE OPEN Fig. 135. resistance due to the air flow, while the water in A stands at a height H the exact equivalent of the velocity head in accordance with the hydraulic formula, v = V2gH, in which v is the velocity of the air in feet per second. If the orifice be partially closed, say 50 per cent, as in Fig. 136, B indi- cates the static pressure, while A gives the dynamic or total pressure due to both velocity and resistance. The difference between A and B is therefore the pressure due to velocity alone. By connecting the two manometers as indicated in Fig. 136 C the velocity pressure is given directly. Pressure. — According to Murgue's theory the maximum pressure which may be developed by a blower or exhauster varies with the square of the speed and may be expressed Cdu 2 V = 9 in which p = pressure, pounds per square foot. d = density of the air, pounds per cubic foot. u = peripheral velocity, feet per second. C = a coefficient obtained by experiment. MECHANICAL DRAFT 255 Tables 34 and 35 give the relationship between pressure and speed for various sizes of forced and induced-draft fans. Fig. 139 shows the relationship between pressure and speed in a 45-inch Buffalo blower as tested at the Armour Institute of Technology. Velocity of Discharge. — The maximum velocity of the air leav- ing the tips of the blades varies directly as the peripheral speed, V = Ku, (63) in which V = velocity of the air discharged, feet per second. K = a coefficient obtained by experiment. u = peripheral velocity, feet per second. ^ a A) /~\ /^\ /?n\ %& (B) o ^ (O ORIFICE PARTLY CLOSED Fig. 136. For practical purposes the velocity of discharge with outlet wide open may be assumed to be that of the periphery. Capacity. — The relationship between capacity and speed, capacity and discharge opening for a 45-inch pressure blower is given in Figs. 139 and 140. As will be noted, the capacity varies almost directly with the speed of the wheel and the area of discharge as expressed by the equation in which Q B A D N Q = BttADN, cubic feet discharge per minute, coefficient determined from experiment, area discharge opening, square feet, diameter of the wheel, r.p.m. of the wheel. (63a) 256 STEAM POWER PLANT ENGINEERING Power. — The power required to drive a fan is proportional to the cube of the speed, Horse Power = XAN 3 , (64) in which X = a coefficient determined by experiment. A = area discharge outlet, square feet. N = r.p.m. The marked increase in power required for even a moderate increase in speed should be borne in mind in selecting a fan. It is as a rule more economical to err in selecting too large a fan than one which must be forced above its rated capacity. In practice the size of fan is proportioned upon experience rather than theory, the usual procedure necessitating the use of curves based upon the performance of fans of the type under consideration. The curves in Fig. 137 were computed by Mr. F. R. Still of the American Blower Company, and give the performance of steel-plate fans as manufactured by this company. These curves apply to this type and make of fan only, though the difference is not very great for any type of centrifugal fan. The " ratio of opening " refers to the actual percentage of opening compared with the total discharge. The " ratio of effect " is the relative effect produced by restricting the discharge. The abbreviations are as follows : D.P. = dynamic or total pressure. P.V.P. = pressure created by a column of air moving at the same velocity as the periphery. S.P. = static pressure. V.P. = velocity pressure = D.P — S.P. Suppose a fan with an unrestricted inlet and outlet delivers 25,000 cubic feet of air per minute against a head (D.P.) of 0.33 inch with a peripheral velocity requiring 6.16 horse power. It appears from the curves that if the discharge outlet is restricted to 50 per cent of the full area, only 12,500 cubic feet will be delivered; the pressure will be increased to 1.03 inches, and the power required drops to 4.84 horse power. If the outlet be still further reduced to 20 per cent of the full opening the capacity will drop to 5000 cubic feet, the pressure will increase to 1.15 inches, and the power will be decreased to 3.45 horse power. With a discharge area of 60 per cent, the mechanical efficiency is a maximum, and equal to about 43 per cent. With orifice closed the horse power required to drive the fan is about 37 per cent of that required when discharging the maximum volume of air. MECHANICAL DRAFT 257 iW /' N. 0. ?• & ^ DO £/ ■S 80 % r> f/ \\ S^ & ^ t™ A' l\ yC "^fefo o / s>- 5 60 / b f7 % I igL o V '3 50 0} y » sP/ &- O / 4^ <% Ac VfK- *g 40 / ^ 9 \ ^« te « 30 / .4! ^ oi \ / •,o d/ / * >/ 4 >/ / ^ 10 f 1 10 20 30 40 50 60 70 80 90 100 110 120 130 140 Ratio of Effect Per Cent Fig. 137. Performance of Steel Plate Fans. i £>L -o 9S rt *k ?'' o o a; tetv sS_ & ^> ^ £^ s 0) *y ^ o a 1! / 4 3 ° a— S"s 3 S* CO ( \ / i >ot a^ o g / -f o 13 ^ / .$) x T - 2 I 45 In Buffalo Blower Speed Constant 1500 R.P.M. Discharge Area Variable \ | 8 i u CO g-io o s / 9 MS p. \ V ft / ,*v \ «e* \ M / -^1 jVOC^t \ Q \ c 1 A rea of Dis 2 :ha rge Op eni 3 ag, Sq aar eF 4 eet 5 0.6 Fig. 138. 258 STEAM POWER PLANT ENGINEERING Curve " K " in Fig. 137 was determined from the empirical formula (based upon Murgue's theorem) A-KQ in which A = area of the inlet orifice, square feet. Q = volume of gas, thousands of cubic feet per minute. P = draft at the inlet in inches of water. K = constant determined by experiment. 500 600 700 800 900 1000 1100 .Revolutions per Minute Fig. 139. (65) -2.0 45 In. Buffalo Blower Discharge Area Constant Speed Variable -1.8 -1.7 -1.6 -1.5 -55- -50- -45- 6 - 2 £ — DO \z& 6 a o a- 2 , 6 -5 |"24f § 1 £? 1,3 -1.2 -1.1 -30- 6 -a- s £& LCie icy ^ || 1 -g-16-o- I-J4S xi | w -g-12-,0- » ! s -0.9 -0.8 -0.7 -0.fi -0.5 S /4» <&/ ■4.0- -3.5 -3.0 -2.5 .2 a -a-- 0Q i. -> „«*< >>V Q -0.3 ^ -1.5 1.0- 0.5 ■ .« 1C f ress ure 1200 1300 1400 1500 The curves in Figs. 138 and 139 are plotted from tests made at the Armour Institute of Technology on a 45-inch Buffalo pressure blower, and are characteristic of this type of fan. Theory of Fans : Power, May, 1907, p. 287; Engr., Oct. 9, 1903, p. 512; Mach., Aug., 1898; Sib. Jour, of Eng., Nov., 1902 ; Heat and Vent., Jan. 15, 1897, July, 1899 ; Prac. Engr., Jan. 16, 1903; Mech. Engr., April 18, 1903; Eng. Rec, Oct. 11, 1902. Pressure Fans vs. Exhaust Fans: Bulletin Ana. Inst. Min. Engrs., Feb., 1909. 153. Determination of Size of Fan. — The following analysis, based upon a paper on Mechanical Draft by F. R. Still of the American MECHANICAL DRAFT 259 Blower Company, gives a good idea of the usual procedure in deter- mining the size of fan for an induced draft installation. (Jour. West. Soc. Engr., May, 1902.) Example : Determine the size of induced fan and the approximate power required to drive it, for a boiler plant rated at 1000 horse power; temperature of flue gases 500 degrees F.; heat value of coal 14,000 B.T.U. per pound; ash 5 per cent; draft required, 1 inch of water pressure. Assuming a boiler efficiency of 70 per cent, the evaporation will be 14 I 000 X Q 70 = 1(U5 pounds of water from and at 212 degrees F. per pound of coal. Since one boiler horse power is equivalent to the evaporation of 34.5 pounds of water per hour from and at 212 degrees F., the, evapora- tion per hour will be 34.5 X 1000 = 34,500 pounds, and the coal burned per hour, -2^522 = 3400 pounds. 10.15 Allowing 18 pounds of flue gas per pound of combustible, 5 per cent for ash and 5 per cent for leaks, the fan will have to handle, at 500 degrees F., approximately 20 X 3400 = 68,000 pounds of gas per hour, or 26,000 cubic feet per minute. It is customary, when little is known about a plant in which a fan is to be installed, to assume that the resistance is equivalent to restricting the discharge outlet 25 per cent. Hence in this problem the various factors are referred to a " ratio of opening " of 75 per cent (see Fig. 137). From formula 65, the area of the inlet should be A ,gg, 0-«5X26 , 126 g fe Vp l which corresponds to a diameter of 48 inches. (K = 0.485 is taken from the curves in Fig. 127.) The area of the inlet should not exceed 40 per cent of the area of the side of the wheel; the latter, then, will be — -j- =31.5 square feet, which corresponds to a diameter of 76 inches (6.3 feet). Referring to Fig. 137, the ratio of dynamic pressure to peripheral velocity pressure (D.P. to P.V.P. at 75 per cent opening) is 0.73. The peripheral velocity pressure will be — - = 1.37 inches of water. 260 STEAM POWER PLANT ENGINEERING The peripheral velocity is U = V2 gH' = 8.03 V#', where H' is the peripheral velocity pressure expressed in feet of gas, or, x> X 62 5 Since W = _£. ! — , where p = inches water, 0.0478 X 12' F U = 87.5 Vl.37. = 102.5 feet per second. = 6150 feet per minute. The maximum effective discharge area which an inclosed fan of this type may have, and still maintain the pressure equivalent of the per- ipheral velocity, is usually called the " blast area." With a larger area the pressure will be reduced, but with a smaller area will remain sub- stantially constant. The velocity of the discharge is practically that of the tips of the blades, whence the blast area is equal to ' = 4.23 square feet, which, with this type of fan is found to be about J the projected rectangle of the wheel, therefore, The projected rectangle = 4.23 X 3 = 12.7 square feet. The proper width of periphery is found by dividing this area by the wheel diameter, thus, width of blades = ^- = 2.02 feet = 24.2 inches, 6.3 and speed of fan = — - =311 r.p.m. P 3.14X6.3 F w _ Volume of gas (cu. ft. per min.) X Pressure (lb. per sq. ft.) 33,000 X efficiency of fan W = 26 > 0Q0 x 5 - 2 = io.2 brake horse power. 33,000 X 0.4 (5.2 = pressure in pounds per square foot equivalent to one inch of water, and 0.4 is the mechanical efficiency for 75 per cent opening as taken from curve in Fig. 137.) Assuming a steam consumption of 70 pounds per brake horse power for a small, simple non-condensing high-speed engine, the steam con- sumed per hour will be 10.2 X 70 = 714 pounds per hour, or 2.3 per cent of the total steam capacity of the boilers. Table 34 gives the capacity and horse power required for various sizes of forced-draft fans, and Table 35 gives similar data for induced- draft fans. MECHANICAL DRAFT 261 154. Chimney vs. Mechanical Draft. — The choice of chimney or mechanical draft depends largely upon local conditions. Many power plants with tall stacks are provided with forced-draft apparatus to be used in emergencies, but as a general rule where ordinances require high chimneys mechanical draft is not considered. In a few isolated cases stokers of the forced-draft type are used in connection with chimneys as high as 250 feet, but such installations are rare, and not to be recommended. Where there are no limitations to the height of stack, mechanical draft offers many advantages over chimney draft, especially for rail- road work and large lighting plants. With certain types of grates and for low-grade fuels and anthracite culm or dust, it is indispensable. Again, where a fair quality of fuel is obtainable the size of plant may determine the choice. First Cost: In small plants of say 100 to 150 horse power the cost of a guyed steel chimney, 75 feet in height or less, would be consider- ably less than that of a mechanical-draft system, and once erected cost practically nothing for operation, while the power required to operate a fan in so small a plant would amount to 5 per cent or more of the total steaming capacity. TABLE 34. CAPACITIES OF FORCED-DRAFT FANS. {Power.) For Forced Draf ,, Temperature of Air 60°. Cubic Feet of Air De- Pressure in Inches of Water. . Diam- eter of livered to Furnace 0.5 0.75 1.00 1.25 1.50 2.00 2.50 per S P4 a Ph' § Ph' a Ph° § Ph' 3 Ph 3 Ph' Minute. Ph 510 1.6 Ph Ph 560 1.8 Ph 600 1.9 Ph Ph~ 640 2.1 Ph Ph" 710 w 2.3 Ph Ph 780 H 2.5 Ph PS* 850 w 2' 6" 4,200 2.7 3' 5,800 430 2.2 460 2.4 490 2.6 530 2.8 590 3.1 640 3.4 710 3.8 3' 6* 7,800 360 3.0 400 3.3 420 3.5 450 3.8 500 4.2 550 4.6 610 5.1 4' 10,000 320 3.9 350 4.2 370 4.4 400 4.9 440 5.4 480 5.9 530 6.5 4' 6" 12,400 290 4.8 310 5.2 330 5.6 360 6.0 400 6.7 430 7.3 470 8.0 5' 15,200 250 5.9 270 6.4 290 6.8 310 7.4 350 8.2 380 8.9 420 9.8 5' 6* 18,200 230 7.0 250 7.7 270 8.2 300 8.8 330 9.8 360 10.6 390 11.8 6' 21,400 210 8.3 230 9.1 250 9.6 260 10.4 290 11.5 320 12.5 350 13.9 7' 28,800 180 11.2 200 12.2 210 13.0 230 14.0 250 15.5 280 16.8 300 18.7 8' 37,200 160 14.4 170 15.7 190 16.7 200 18.1 220 20.1 240 21.8 270 22.5 9' 46,800 140 18.1 160 19.8 170 21.1 180 22.7 200 25.3 220 27.4 240 30.3 10' 57,400 130 22.2 140 24.3 150 25.8 160 27.9 180 3.1 200 33.6 210 37.2 Discharge velocity 2000 feet per minute. 262 STEAM POWER PLANT ENGINEERING TABLE 35. CAPACITIES OF INDUCED-DRAFT FANS. {Power.) For Induced Draft, Temp, of Flue Gases 500°. Cubic Feet of Air at 60°Temp. Drawn into Fur- nace per Minute. Pressure in Inches of Water. Diam- eter of 0.5 0.75 1.00 1.25 1.50 2.00 2.50 Fan. 3 Oh' 688 580 486 432 390 337 310 283 243 216 189 175 Oh' M 2.2 3.0 4.0 5.3 6.5 8.0 9.5 11.2 15.1 19.4 24.4 30.0 3 Oh' pi 756 621 540 472 418 364 337 310 270 230 216 190 Ph* w 2.4 3.2 4.5 5.7 7.0 8.6 10.4 12.3 16.5 21.2 26.7 32.8 Oh Ph 810 661 567 500 445 391 364 337 283 256 230 202 Ph' w 2:6 3.5 4.7 6.1 7.5 9.2 11.1 13.0 17.5 22.5 28.5 34.8 Ph' Ph- 864 715 607 540 486 418 405 351 310 270 243 216 pj W 2.8 3.8 5.1 6.6 8.1 10.0 11.9 14.0 18.9 24.4 30.6 37.6 3 Ph pi 958 796 675 594 540 472 445 391 337 297 270 243 Ph W 3.1 4.2 5.7 7.3 9.0 11.1 13.2 15.5 20.9 27.1 34.1 41.8 d Ph Ph 1053 864 742 648 580 513 486 432 378 324 297 270 Ph' a 3.4 4.6 6.2 8.0 9.8 12.0 14.3 16.9 22.6 29.4 37.0 45.3 Ph' Ph 1147 958 823 715 634 567 526 472 405 364 324 283 Ph" U 2' 6* 3' 3' 6" 4' 4' 6" 5' 5' 6" 6' V 8' 9' 10' 3,000 4,200 5,700 7,300 9,300 11,100 13,300 15,600 21,000 27,100 34,200 41,900 3.6 5.1 6.9 8.8 10.8 13.2 15.9 18.7 25.2 30.4 40.9 50.2 A tall, self-supporting chimney for larger plants, however, is very costly as compared with a fan system of equal capacity. For example, a brick chimney 175 feet high and 10 feet in diameter, foundation and all, capable of furnishing the necessary draft for a 3000-horse-power plant, will cost about $10,000. A two-fan induced system of equiv- alent capacity will cost in the neighborhood of $5000, a one-fan system $3500, and a forced-draft system $2500. See Fig. 140. With interest at 5 per cent, depreciation 5 per cent, taxes 1 per cent, and insurance one-half per cent, the annual fixed charges will be $575, $402.50, $287.50, respectively, for the fan equipment. Depreciation and Maintenance : The depreciation of a well-designed masonry or concrete stack is very low, and 2 per cent is a liberal factor. Maintenance is practically negligible, as it requires no attention what- ever for years. A steel stack, however, must be kept well painted or corrosion will take place rapidly. The depreciation and maintenance charges on a mechanical-draft system will range from 4 per cent to 10 per cent of the original outlay. Cost of Operation: Once erected, the comparative cost of operating a chimney is practically nothing; that is, of course, on the assumption that the chimney and fan exhaust equal volumes of gas per pound of MECHANICAL DRAFT 263 fuel and at the same temperature. A fan system requires for its opera- tion from one and one-half per cent to five per cent of the total steaming capacity of the plant, depending upon the type and character of the fan engine or motor, and the conditions of operation. Efficiency: With fan draft a very thick fire can be maintained on the grate, thus permitting a high rate of combustion, and minimum air per pound of fuel, both of which result in increased boiler efficiency. 15 j 14 13 12 1 10 / ^- « 9 +3 / o O 8 / / / *a " / "o V tf £ / 6 <¥/ S v / s D£ !§ •2 1 faij 5- / 3 / \<° & V jfg.1 V,- / / y 1 J >^ rvfl "0* ^ / / y •-- 1S5 V- V H / t / > / ^ -^ •p,3 D ra ft S s \ 1 - — — i ___ — - — -" - ""■ 1000 2000 3000 Horse Power Fig. 140. Comparative Costs of Chimneys and Mechanical Draft. (W. B. Snow.) The influence of the rate of combustion on air supply is illustrated in Fig. 141. For the same temperature of discharge each pound of air in excess of theoretical requirements results in a loss of about one per cent of the total heat in the fuel. (See Table 3.) With fan draft an average figure is 18 pounds of air per pound of bituminous coal against 24 pounds for the chimney, a saving of 5 per cent in favor of the fan. Again, a fan permits of a low temperature of the flue gases without 264 STEAM POWER PLANT ENGINEERING affecting the draft, while lowering the temperature in the chimney reduces the draft as shown in Table 24. From Table 4 we see that a reduction in flue gas temperature of 25 degrees F. will increase the boiler efficiency about one per cent. With an economizer the flue gases may be reduced to 350 degrees F., with a net saving of about 500 — 350 = 150, or 6 per cent of the total fuel. It is in this connection that the fan draft is peculiarly suitable. Of course, the chimney may be provided with an economizer, effecting the same reduction in tem- 300 =5 200 100 o o 10 20 40 50 X,b.Coal Burned Per Sq.Ft.Grate Per Hr. Fig. 141. Influence of Rate of Combustion on Air Supply. — Forced Draft. perature, but its height must be made sufficiently great to overcome the additional resistance of the economizer and the reduction in tem- perature of the chimney gases. Flexibility : With a fan the draft may be readily regulated for sudden increased or decreased requirements, independent of the boiler performance. Damp and muggy days appreciably affect the draft of a chimney, as do adverse air currents and high winds. Smoke: Smokeless combustion is more readily effected with arti- ficial draft than with natural draft, as a thicker fire can be carried, and the correct proportion of air can be more readily adjusted. Comparative Tests of Chimney and Mechanical Draft: Power, July, 1901, p. 22; Eng. Rec, July 25, 1903, p. 102; Eng., U. S., May 1, 1899, p. 105; Engr., U. S., April 15, 1907. 155. Balanced Draft. — Fig. 142 illustrates an application of the McLean " Balanced Draft " system to a water-tube boiler. The equipment consists of a blower, the speed of which is regulated MECHANICAL DRAFT 265 a CD 0Q 266 STEAM POWER PLANT ENGINEERING by the steam pressure, so that the draft in the fire box is main- tained at approximately atmospheric pressure. The chief claims for this system are (1) the velocity of the gases over the tubes is reduced, and short circuiting is prevented; (2) the correct proportion of air to fuel is readily maintained; (3) infiltration of air through the setting is impossible, as the pressures are " balanced"; (4) sudden changes in load are correctly taken care of. Tests of the apparatus at the Fuller Building, New York, gave excellent results (Trans. A.S.M.E., 26-641). CHAPTER IX. STEAM ENGINES. 156. Introductory. — The reciprocating steam engine is the most widely distributed and generally adopted prime mover in the power world although its field of usefulness has been greatly encroached upon in recent years by the steam turbine and the gas engine. The steam turbine has practically superseded the piston engine for large steam electric plants, while in other fields the gas engine offers many advan- tages, but the reciprocating steam engine is still an important heat engine and will probably continue to be a factor in the power world for years to come. The type of engine best suited for a given installation is the one which delivers the required power at the lowest cost, measured in dollars and cents, taking into consideration interest on the investment, operating expenses, maintenance and depreciation. 157. Ideal Engine. — The thermal efficiency of the steam engine is expressed by the ratio of the heat equivalent of the work done on the piston per unit of time, to the heat supplied. The degree of per- fection realized is ascertained by comparing the performance of the real engine with that of an ideally perfect engine, working between the same temperature limits. The theoretical limit of perfection is that defined by the Carnot cycle, the efficiency of which is represented by the equation E C = T '~ T \ (66) in which T x = the highest absolute temperature of the working fluid. T 2 = the lowest absolute temperature of the working fluid. The upper limit of temperature is that corresponding to boiler pressure, and the lower limit to that of the exhaust steam. Evidently the greater the temperature range the more nearly does the ideal efficiency approach unity, but with the present limits of temperature used in steam engines, it cannot exceed about 35 per cent. The nearest approach of any actual engine to the Carnot cycle is accomplished by the Nordburg system of progressive feed heating, 267 268 STEAM POWER PLANT ENGINEERING in which the feed water is successively heated from the receivers inter- mediate between each pair of cylinders. (Engineering News, May 4, 1899, p. 283.) Table 36 gives the Carnot efficiencies of condensing and non-condensing engines for ordinary ranges of steam pressures. The Carnot cycle is theoretically impossible for an engine using superheated steam at constant pressure, and, in general, it is not very closely simulated by engines using saturated steam. It is, therefore, i more instructive to select an ideal cycle which \ more nearly represents the performance of the \ actual engine. The diagram representing the \ operation of this perfect engine is shown in \. Fig. 142a, and is called the non-conducting or ^v Rankine cycle, ab represents the admission _^^ 6 - of dry steam from the boiler at pressure p x ', &; be is an adiabatic expansion to exhaust pressure p 2 ; cd represents the exhaust, and rfa is an adia- batic compression to the initial pressure. The heat necessary to raise the feed water from the temperature of exhaust, or ideal feed water temperature, to the temperature in the boiler and evaporate it into dry steam is #i - r x +. q ± - q 2 , (66) in which H t = quantity of heat supplied to the cylinder per pound of steam. r t = heat of vaporization at pressure p v q x = heat of the liquid at pressure p v q 2 = heat of the liquid at pressure p 2 . The heat, H 2 , exhausted from the cylinder and which must be with- drawn when it is condensed is H 2 = x 2 r 2 , (66a) in which x 2 = quality of the steam at pressure p 2 . r 2 = heat of vaporization at pressure p 2 . x 2 may be calculated by the aid of equation * 2 = | 2 (^ + 1 -0 2 )> . (66b) STEAM ENGINES 269 270 STEAM POWER PLANT ENGINEERING STEAM ENGINES 271 in which T 2 = absolute temperature of steam at pressure p 2 . T x = absolute temperature of steam at pressure p v 6 X = entropy of the liquid at pressure p v d 2 = entropy of the liquid at pressure p 2 . Other notations as above. The heat changed into work per pound of steam is H i ~ H 2 = r x + q x - q 2 - x 2 r 2 , (67) and the efficiency, E r , of the cycle is E = H >- H > = 'i + gi ~ v, - g, . ' (67a) H t r 1 + q 1 - q 2 The steam consumption W, or water rate, lbs. per h.p. hr. of the perfect engine, may be expressed W= , 2545 (67b) n + 2i ~ g 2 . If the steam entering the cylinder is wet and of quality x lf substitute x 1 r 1 in above equations for r v If the steam is superheated at admission but becomes moist at the lower pressures, which is the usual case, the efficiency may be expressed E r i + gi +c 1 t s -x 2 r 2 -q 2 t r i + Qi + Cits - q 2 in which c x = mean specific heat of the superheated steam at pressure p v t s = degree of superheat or difference in temperature between the superheated and saturated steam at pressure p v x 2 may be calculated by the aid of equation Jr Tc ~P + ^+e l = ^ + d 2 , (67d) in which t and T = thermometric and absolute temperatures of the superheated steam. c = true specific heat of superheated steam at temperature t. Other notations as above. 272 STEAM POWER PLANT ENGINEERING O d a o O I d d STEAM ENGINES 273 For many purposes equation (67d) may be expressed e, log.|j + £+*,- 3& +0,. (67e) For highly superheated steam in which the steam is still superheated at exhaust T? — r i "^" ?1 ~^" C J S ~ r 2 ~ ?2 ~ C 2^S C67f ^ r t + q t + c t t s - q 2 in which c 2 = mean specific heat of the superheated steam at exhaust. t/ = degree of superheat at exhaust. t s ' may be calculated by the aid of equation f T cdt ^r, n CW c'dt , r 2 ' Problems connected with the Rankine cycle may be conveniently solved by temperature-entropy tables to be found in connection with the usual steam tables or by the Mollier diagram as described in Appendix H. 158. Thermal Efficiency of the Actual Engine. — In calculating the thermal efficiency of the real engine the heat supplied is reckoned above the sensible heat of the exhaust, thus, _ Heat converted into useful work Heat supplied 42.42 B.T.U. supplied per I.H.P. per minute 42.42 w(x l r l +q v - q 2 )' in which (69) (70) w = the weight of steam supplied to the engine per indicated horse power per minute, or per brake horse power per minute, de- pending upon whether the efficiency is to be referred to the indicated or to the brake horse power of the engine. Other notations as in (66) and (67). The figure obtained by dividing the efficiency of the real engine by that of the ideal engine is called the efficiency ratio, and is a measure of the extent to which the theoretical possibilities are realized. 274 STEAM POWER PLANT ENGINEERING The efficiency ratio is calculated on the basis of the indicated horse power or the developed horse power: Eff. Ratio = 42.42 E r w (H 1 - H 2 ) (71) This ratio expressed in different terms has been referred to as a Poten- tial Efficiency" by C. V. Kerr (Trans. A.S.M.E., 25-920), and as " Cylinder Efficiency " by Professor Reeve. The commercial economy of an engine is measured by the cost of producing power, and does not necessarily depend upon its thermal efficiency. The performances of steam engines are frequently stated in terms of (a) pounds of steam utilized per horse-power hour, (6) pounds of coal per horse-power hour, (c) cost in cents per horse-power hour, (d) B.T.U. per horse-power hour. From a commercial stand- TABLE 36. STEAM-ENGINE EFFICIENCIES. (Saturated Steam.) Non-Condensing; Back Pressure 14.7 Condensing; Back Pressure 1 Pound Absolute. Absolute. Gauge Fress. Ratio Ratio Carnot Rankine Actual * b a Carnot Rankine Actual b a Cycle. Cycle (a). (b). Cycle. Cycle (a). (b). % % 25 7.5 7.3 5.5 76.0 22.6 21.0 11.6 55.0 50 11.2 10.7 8.5 80 25.7 23.5 13.5 60 75 13.7 13.0 10.4 80 27.8 25.3 15.9 61 100 15.7 14.8 12.0 81 29.5 26.7 20.2 76 125 17.3 16.3 13.5 83 30.8 27.8 20.3 74 150 18.7 17.5 14.3 82 32.0 28.8 21.6 75 175 19.8 18.5 14.8 80 32.9 29.6 21.9 74 200 20.8 19.3 15.2 79 33.7 30.2 22.6 75 225 21.6 19.9 15.5 78 34.5 30.6 22.6 74 250 22.4 23.0 23.6 20.5 21.0 . 21.4 35.1 35.6 36.0 31.0 31.3 31.5 275 300 * Best recorded performance of the actual engine, 1907 point the cost of producing power is the most important basis of com- parison, but the latter expression is most satisfactory for scientific purposes, since it gives a basis of comparing the performances of all types of heat engines. STEAM ENGINES 275 159. Mechanical Efficiency. — The power of an engine may be expressed in terms of indicated horse power, brake horse power, or pump horse power, according to the class of engine. The ratio of the brake to the indicated power is the mechanical efficiency of the engine, the ratio of the electric horse power to the indicated power is the mechanical efficiency of the engine and generator com- bined, and the ratio of the pump horse power to the indicated power of the steam engine is the mechanical efficiency of the engine and pump 45 n r X ^-< .— — ° C tf *f r<& * •*o o o/ % V .?v : / / / Mechanical Efficiencies of 75 KW.Generating Set Engine, Simple High Speed !Non Condensing // ( / / /o / 10 20 30 40 50 60 70 80 90 100 110 120 Per Cent of Rated Load Fig. 143. 140 150 combined. The percentage of work lost in friction is therefore the difference between 100 per cent and the mechanical efficiency. Table 37 shows the mechanical efficiencies for several types of en- gines, and Fig. 143 the combined efficiency of a direct-connected high- speed engine and generator. (See Engine Friction, par. 167.) Mechanical Efficiency: Peabody, Thermodynamics, p. 430; Spangler, Steam Engineering, p. 205; Ripper, Steam Engine, p. 275; Ewing, Steam Engine, p. 186. The following numerical example will illustrate the calculation of the various efficiencies mentioned: A simple high-speed engine uses 30 pounds of steam per I.H.P. hour; initial pressure 100 pounds per square inch, gauge; exhaust pressure, atmospheric; I.H.P., 120; D.H.P., 102. Steam assumed to be dry and saturated at throttle. Required: (1) the actual thermal efficiency; (2) the efficiency of the Carnot cycle; (3) the efficiency of the Rankine cycle; (4) the efficiency ratio; (5) the mechanical efficiency. 276 STEAM POWER PLANT ENGINEERING TABLE 37. MECHANICAL EFFICIENCIES OF ENGINES. Kind of Engine. Horse Power. Efficiency at Full Load. Simple : 1. High-speed, non-condensing 150 170 275 95 5 2 High-speed, condensing 96 3. Low-speed, non-condensing 94 4. Low-speed, condensing Compound : 5 High-speed, non-condensing 150 160 900 1000 865 712 94 6. High-speed, condensing 98 7. Low-speed, non-condensing 8. Low-speed, condensing 95 95 Triple: (combined efficiency of engine and pump) 9 Pumping engine 97.4 Quadruple: (combined efficiency of engine and pump) 10. Pumping engine 93 1. Buffalo Simple engine, 12 X 12, Elec. World, Sept., 1904, p. 147. 2. Reeves Simple engine, 15 X 14, Elec. World, Oct. 1, 1904, p. 587. 3. 24 X 48 Hamilton Corliss at Armour Inst, of Tech., 1898. 4. 5. Reeves Compound; Eng. Rec, July 1, 1905, p. 24. 6. Reeves Compound; Eng. Rec. 7. 21, 41 X 30 Cross Compound Ball & Wood, West Albany Station, N.Y.C. & H.R.R, 8. 20, 40 X 42 Rice & Sargent; A.S.M.E., 29-1276. 9. Allis Pumping Engine; Power, May, 1906, p. 299. 10. Nordburg Pumping Engine; Eng. News, May 4, 1899, p. 280. (1) The actual thermal efficiency is 42.42 E = W {x 1 r 1 + q l - q 2 ) 42.42 30/60 (879.6 + 308.9 - 180.3) = 0.084. That is, only 8.45 per cent of the heat supplied to the cylinder above the temperature of the exhaust steam is converted into work. The assumption that the exhaust steam may be used to heat the feed water to its own temperature justifies reckoning the heat supplied above the exhaust temperature. (2) The efficiency of the Carnot cycle is T t -T 2 E c 798.8 - 673 798.8 = 0.157, STEAM ENGINES 277 which means that if the engine were a perfect one employing the Car- not cycle between the same extremes of temperature as the actual engine, 15.7 per cent of the energy supplied would be converted into useful work. (3) On the basis of the Rankine cycle the ideal efficiency becomes V. — X i r i "*" 9l ^2 r 2 9.2 x Si + 9i ~ 92 879.6 + 308.9 - 0.885 X 969.7 - 180.3 879.6 + 308.9 - 180.3 = 0.149, the numerator representing the heat converted into work per pound of steam and the denominator the heat supplied. Thus if the engine be 300 275 250 Status of the Piston Engine 1907 Saturated Steam ? V %y 8 10 12 14 16 18 20 22 24 26 Thermal Efficiency, Per Cent Fig. 144. 32 34 assumed to have a non-conducting cylinder and work on the Rankine cycle, it would be capable of utilizing 14.9 per cent of the heat sup- plied. (4) The efficiency ratio is expressed: Efficiency ratio = — ~ = ^— - J E r 14.9 56.7 per cent, 278 STEAM POWER PLANT ENGINEERING which indicates the degree of perfection of the engine or the extent to which it realizes the maximum efficiency theoretically possible. The weight of steam which would have to be consumed per horse power per hour to actually attain the Rankine efficiency would be 30 X ^2^ = 30 x 0.567 = 17 pounds. T) TT "P 1 f)9 (5) The mechanical efficiency = ' ' ' = =0.85 = 85 per cent. The relation between the actual and the theoretical efficiencies based upon the Rankine cycle, and upon the best recorded perform- ances of modern engines, using saturated steam, is shown graphically in Fig. 144. (Also see Tables 39 and 40.) The theoretical curves are calculated upon the assumption of complete expansion, the back pressure being 14.7 pounds gauge for non-condensing and one pound absolute for condensing engines. The highest recorded (1907) efficiency ratio for a steam engine using saturated steam is 83 per cent non-condensing and 76 per cent condensing (see Table 39), and 78.5 per cent for a condensing engine using superheated steam (Table 43). 160. Heat Losses in the Steam Engine. — The principal losses which tend to lower the efficiency of the steam engine are due to (a) Presence of moisture in the steam at admission. (b) Leakage past valves and piston. (c) Cylinder condensation. (d) Clearance volume. (e) Incomplete expansion. (/) Wire drawing. (g) Friction. (h) Radiation. 161. Moisture. — The presence of moisture in the steam pipe is due to condensation caused by radiation or to priming at the boiler. Unless removed by some separating device between boiler and engine the amount of moisture entering the cylinder may be from 1 to 5 per cent of the total weight of steam, and the work done per pound of fluid is correspondingly reduced. This loss should not be charged against the engine, however, and its performance should be reckoned on the dry steam basis. Experiments reported by Professor R. C. Car- penter (Trans. A.S.M.E., 15 — 438) in which water in varying quantities was introduced into the steam pipe, causing the quality of the steam to range from 99 per cent to 57 per cent, showed that the consumption STEAM ENGINES 279 of dry steam per I.H.P. hour was practically constant, the water acting as an inert quantity. An efficient separator will remove practically all the entrained water. Influence of Moisture on Steam Economy: Trans. A.S.M.E., 18-699; Benjamin, Heat and Steam; Perry, Steam Engine, p. 353; Rankine, Steam Engine, p. 407; Ripper, Steam Engine, p. 33; Ewing, The Steam Engine, p. 139. 162. Leakage of Steam. — The loss due to leakage is a variable factor depending upon the design and condition of the engine, and is greater with saturated than with superheated steam. The usual method of measuring leakage past the valves and piston while the engine is at rest is likely to give erroneous results as demonstrated by Callender and Nicolson (Peabody, " Thermodynamics/' p. 351) in tests made on a high-speed automatic balanced valve engine and on a quadruple expan- sion engine with plain unbalanced slide valves. With the engines at rest they found that the leakage past valves and piston was insignificant, but when in operation the leakage from the steam chest into the exhaust was very considerable indeed. It was thought that a large proportion of the leakage was probably in the form of water formed by condensa- tion of steam on the seat uncovered by the valve. According to the report of the Steam Engine Research Committee (Eng. Lond., March 24, 1905, p. 298), leakage through a plain slide valve is independent of the speed of the sliding surfaces, and directly proportional to the difference in pressure on the two sides; with well- fitted valves the leakage is never less than 4 per cent of the volume of steam entering the cylinders, and is often greater than 20 per cent. Peabody, Thermodynamics, p. 350; Eng. Rec., May 22, 1897, p. 529; Barms, Engine Tests, p. 251. 163. Cylinder Condensation. — A large percentage of the steam admitted to the cylinder is condensed, due to the absorption of heat by the relatively cool cylinder walls. Condensation continues during expansion until the temperature of the steam falls below that of the metal, when the process is reversed and a part of the moisture is re- evaporated. Unless the cylinder is one of a compound series, the heat absorbed from the cylinder walls during exhaust does no useful work and is lost. The condensation up to the point of cut-off (initial con- densation) may amount to from 15 to 30 per cent, and is often as high as 50 per cent of the total weight admitted to the cylinder. The initial condensation becomes greater as the difference between initial and exhaust pressures is increased, and diminishes as the speed of the engine increases. Cylinder condensation and leakage are ordinarily 280 STEAM POWER PLANT ENGINEERING classified together, as there is no way of separating them accurately. They represent that part of the feed water which is not accounted for by the indicator diagram. Tests of 20 simple high-speed engines by G. H. Barms, Fig. 145, show some results obtained for various percentages of cut-off. Also see table compiled by C. H. Peabody, " Thermodynamics of the Steam 60 o 1 Condensation.and Leakage for Simple Engines using Saturated Steam* o 1 DO \o s a o o n s° 3 s ^ < J? 1 20 >r ^ 3 r O io E lgine Tests , Ban us, p. 254. 10 15 20 35 Percentage of Cut Off Fig. 145. 40 Engine," p. 336, showing analysis of the heat interchanges for a number of different types of steam engine. The various heat losses, including cylinder condensation and leakage, are best determined by transferring the indicator diagram to the tem- perature entropy or 0 chart. (See Appendix C.) This is useful for certain scientific investigations, but is unnecessary for commercial tests. Cylinder Condensation : Trans. A.S.M.E., I, 184, III, 215, IV, 88, VII, 375, XVIII, 950; Cotterill, Steam Engine, p. 331; Spangler, Steam Engineering, p. 228; Thurston, Manual of the Steam Engine, I, 271, 488, 585; Heck, Steam Engine, pp. 109, 113, 119; Ewing, Steam Engine, p. 148; Hutton, Heat Engines, p. 319; Peabody, Thermodynamics, pp. 241, 412; Reeve, Thermodynamics, p. 198; Wood, Thermodynamics, p. 212; Perry, Steam Engine, p. 78; Ripper, Steam Engine, p. 25. Initial Condensation: Cotterill, Steam Engine, p. 274; Golding, 6$ Diagram, p. 63; Marks, Steam Engine, p. 195; Peabody, Thermodynamics, p. 359; Popplewell, Heat Engine, pp. 323, 351; Reeve, Thermodynamics, pp. 156, 221; Ripper, Steam Engine, pp. Ill, 168. STEAM ENGINES 281 Condensation during Expansion: Trans. A.S.M.E., III, 286; Hutton, Heat Engines, pp. 223, 286; Pupin, Thermodynamics, p. 88; Rankine, Steam Engine, p. 385; Reeves, Thermodynamics, p. 221. Entropy : Power, Jan. 21, 1908; Baynes, Thermodynamics, p. 94; Benjamin, Heat and Steam, p. 37; Berry, Temperature-Entropy Diagram ; Boulvin, Entropy Diagram; Ewing, Steam Engine, p. 103; Golding, 0 Diagram; Hutton, Heat Engines, p. 276; Peabody, Thermodynamics, p. 97 ; Reeve, Thermodynamics, p. 39; Swinburne, Entropy; Wood, Thermodynamics, p. 136: Heck, Steam Engine, Chap. VI. 164. Clearance Volume. — The portion of the cylinder volume not swept through by the piston but which is nevertheless filled with steam when admission occurs is called the clearance volume. It is the space between the end of the piston when on dead center and the inside of the valves covering the ports. It varies from about 1 per cent of the piston displacement in very large engines with short steam passages to 10 per cent or more in small high-speed engines. When the steam retained in the clearance space is compressed to the initial pressure and expansion is carried down to the back pressure, the clear- ance has little effect upon the economy of the engine, but since expan- sion and compression are seldom complete in actual practice, the loss may be considerable. (Ripper, " Steam Engine," p. 103.) The shorter the cut-off the greater will be the ratio of the weight of cushion steam to that of the steam supplied and hence the greater the relative loss. In large slow-speed engines the loss may be insignificant if the clearance volumes are small, while in small high-speed engines it may be con- siderable. The ratio of expansion is decreased by clearance; for example, an engine cutting off at one-fifth, neglecting clearance has an apparent ratio of expansion of 5, but if the clearance volume is 10 per cent the actual ratio is only 3.66. One of the few recorded tests relative to the influence of clearance on the economy of a high-speed engine was con- ducted on a 14x15 Allfree engine. (Power, May, 1901.) With a clearance volume of 2.2 per cent, initial pressure 105 pounds gauge, and 172 r.p.m., the best performance was 23.7 pounds of dry steam per I.H.P. hour. With the same steam pressure and speed, but with clearance volume increased to 6 per cent by the use of a shorter piston, the best performance was 28.3 pounds per I.H.P. hour. In both cases the compression was carried up to admission pressure. Loss by Clearance in Steam Engines ; Trans. A.S.M.E., XVIII, 176. Clearance in Compound Engines: ibid., I, 173. Clearance in M ulti-cy Under Engines : ibid., XI, 151. Effect of Clearance: Ewing, Steam Engine, p. 145; Hutton, Heat Engines, p. 334; 282 STEAM POWER PLANT ENGINEERING Peabody, Thermodynamics, p. 407; Popplewell, Heat Engines, p. 332; Reeve, Ther- modynamics, pp. 223, 256; Ripper, Steam Engine, pp. 63, 101; Spangler, Steam Engineering, p. 114; Wood, Thermodynamics, p. 197. 165. Loss Due to Incomplete Expansion and Compression. — Theo- retically the loss due to incomplete expansion is considerable. For example, the theoretical steam consumption of a perfect engine (Ran- kine cycle) expanding from 120 pounds absolute to a condenser pres- sure of 2 pounds absolute is 9.6 pounds per horse-power hour. If the expansion were carried to only 5 pounds absolute, the exhaust pressure remaining the same, the steam consumption would be increased to 11.8 pounds per horse-power hour, a difference of 22 per cent for an increase in terminal pressure of only 3 pounds per square inch. The theoretical water rates for various terminal pressures are given below. Terminal Pressure, Pounds per Square Inch Absolute. Steam Consumption of Perfect Engine. Terminal Pressure, Pounds per Square Inch Absolute. Steam Consumption of Perfect Engine. 1 1.5 2 2.5 8.5 9.1 9.6 10 3 4 5 6 10.4 11.1 11.8 12.3 In actual engines expansion is seldom complete, since it would necessitate increased bulk and weight of engine, and the work done by the steam in the last stages would not compensate for the increased cost. In single-cylinder engines maximum economy is effected when the terminal pressure is considerably above that of the exhaust, since the gain due to complete expansion is more than offset by the increased cylinder condensation. This is true to a certain extent in all engines irrespective of the number of cylinders. Tests by G.H. Barrus ("Engine Tests," 1900) to determine the terminal pressures effecting maximum economy for various types of engine gave results as follows : Simple slide-valve engines, non-condensing Simple slide-valve engines, condensing .... Simple Corliss engines, non-condensing. . . . Simple Corliss engines, condensing Compound engines, non-condensing Compound engines, condensing Terminal Pressure, Pounds Absolute. 30 to 40 25 to 30 20 to 25 15 to 18 18 to 22 3 to 5 STEAM ENGINES 283 In high-speed engines a certain amount of compression is desirable for its cushioning effect; outside of this mechanical feature compression may or may not be of benefit to the engine as will be explained later. Zuener in his treatise on theoretical thermodynamics proves deduc- tively that in an engine with a large clearance volume the loss due to clearance is completely eliminated if the compression is carried up to admission pressure, a conclusion which tests by Jacobus, Carpenter, and others fail to confirm. A series of tests by Professor Jacobus (Trans. A.S.M.E., 15-918) on a 10x11 high-speed automatic engine at Stevens Institute show decreasing economy with increase of com- pression, the initial pressure, cut-off, and release remaining constant. The results were as follows : Proportion of initial pressure up steam is compressed Steam, pounds per I.H.P. hour. . . . to which the Full 38 Tests by Carpenter (Trans. A.S.M.E., 16-957) on the high-pressure cylinders of the Corliss engine at Sibley College gave: Compression, per cent Brake horse power Steam, pounds perB.H.P. hour. 11.4 25 30 29 33 33.3 35.2 26 34 6.8 6.6 6.5 6.4 6.3 6.2 6.1 6.0 5.9 5.8' 5.7 5.6 5.5 . 48 47 46 15 100 Lb. Gauge ^^ \ ;^w ^^"t < H ^* ^vV 13 n %^ li ^ ^*Oj e ^ •10 \^ 39 38 37 ,36 35 s* Influence of Back Pressure on the Economy of an 8 x 10 Automatic High Speed Non Condensing Engine 1 1 1 *> 6 8 10 12 14 Back Pressure.Q). Per Sq.In.Gauge Fig. 145a. 16 18 Opposed to these figures are tests which show an improvement in economy when compression is increased. Fig. 145a shows the influence of increasing back pressure on the 284 STEAM POWER PLANT ENGINEERING economy of an 8 x 10 automatic high-speed engine at Armour Institute of Technology. Cut-off. — Best for Different Pressures; Trans. A.S.M.E., 4-89; In Compound Engines: ibid., p. 549; Most Economical Point of: ibid., 8-486; Hutton, Heat Engines, p. 232; Peabody, Thermodynamics, p. 210; Spangler, Steam Engineering, p. 109; Wood, Thermodynamics, pp. 200, 433; Klein, High-Speed Engines, p. 7; Ewing, Steam Engine, p. 84. Ratio of Expansion: Trans. A.S.M.E., 2-19, 128, 10-576, 11-166; Ewing, Steam Engine, pp. 47, 159; Wood, Thermodynamics, pp. 154, 172, 197, 295; Rankine, Steam Engine, pp. 378, 553; Reeve, Thermodynamics, p. 228; Thurston, Manual of the Steam Engine, 1-271, 725, 2-14. Incomplete Expansion: Boulvin, Entropy Diagram, p. 28; Cotterill, Steam Engine, p. 240; Perry, Steam Engine, p. 364; Popplewell, Heat Engines, p. 332; Peabody, Thermodynamics, p. 238; Reeve, Thermodynamics, p. 222; Heck, Steam Engine, p. 78. Compression. — Efficiency of Compression in Steam Engine : Engr., Lond., Nov. 3, 1905, p. 434; Compression as a Factor in Steam Engine Economy : Trans. A.S.M.E., 14-1067; Effect of Compression on Water Consumption: ibid., 15-815; Engine Com- pression: ibid., 7-708; In High-Speed Engines: ibid., 7-202; In Steam Cylinder: ibid., 2-341. Back Pressure. — Back Pressure as Modifying Economy : Trans. A.S.M.E., 18-283; On Valves : ibid., 3-150; General : Ewing, Steam Engine, pp. 84, 145; Klein, High- Speed Engines, p. 11; Perry, Steam Engine, p. 75; Reeve, Thermodynamics, p. 223; Ripper, Steam Engine, p. 53. 166. Loss due to Wire Drawing. — Wire drawing, or the drop in pressure due to the resistances of the ports and passages, has the effect of reducing the output of the engine to some extent, since the pressure within the cylinder is less than that at the throttle during admission and greater than discharge pressure at exhaust. The steam may be dried to a small extent during admission. In single- valve engines the effects of wire drawing are decidedly marked and the true points of cut-off and release are sometimes difficult to locate on the indicator card. In engines of the Corliss or gridiron-valve type the effects are hardly noticeable. Wire Drawing: Trans. A.S.M.E., 2-344, 1-174; Ewing, Steam Engine, pp. 95, 143, 207; Boulvin, Entropy Diagram, p. 56; Popplewell, Heat Engine, p. 320; Rankine, Steam Engine, p. 413; Reeve, Thermodynamics, pp. 105, 221; Ripper, Steam Engine, p. 73; Wood, Thermodynamics, p. 195; Heck, Steam Engine, pp. 183, 224, 230. 167. Loss due to Friction of the Mechanism. — The difference between the indicated horse power and that actually developed is the power consumed in overcoming friction, and varies from 4 to 20 per cent of the indicated power, depending upon the type and condition of the engine. Engine friction may be divided into (1) initial or no-load STEAM ENGINES 285 friction and (2) load friction. The stuffing-box and piston-ring friction is practically independent of the load, while that of the guides, bearings, 25 50 100 125 150 175 200 225 250 2T5 Developed Horse-Power Fig. 146. Typical Curves of Steam Engine Friction. and the like increases with the load. In Fig. 146, curve A gives the relation between the frictions for a four-slide-valve horizontal cross compound engine, and B that for a simple non-condensing Corliss. TABLE 38. DISTRIBUTION OF FRICTION IN SOME DIRECT-ACTING STEAM ENGINES. (Thurston.)* Percentage of Total Engine Friction. Parts of Engines where Friction is Measured. " Straight Line " Balanced Valve. " Straight Line " Unbalanced Valve. Traction Engine Locomotive Valve Gear. Automatic Balanced Valve. Condensing Engine Balanced Valve. Main bearings 47.0 35.4 35.0 41.6 46.0 Piston and piston rod 32.9 25.0 21.0 49.1 Crank pin 6.8 5.1 13.0 21.8 Crosshead and wrist pin. ...... 5.4 4.1 Valve and valve rod 2.5 26.4 22.0 9.3 21.0 Eccentric strap 5.4 4.0 Link and eccentric 9.0 Air pump 12.0 100.0 100.0 100.0 100.0 100.0 * " Friction and Lost Work in Machinery," p. 13. 286 STEAM POWER PLANT ENGINEERING (Peabody's " Thermodynamics," pp. 433 and 437.) Curve C is plotted from the tests of a Reeves vertical cross, compound condensing engine (Engineering Record, July 1, 1905, p. 24), and D from the test of an Ames simple high-speed non-condensing engine. (Engineering Record, Vol. 27, p. 225.) A large number of recorded tests show less friction at full load than at no load, but this is probably due to error or to variations in lubrication. With first-class lubrication it is usually sufficiently accurate to assume the friction to be constant and equal to the initial friction at zero load. The distribution of the frictional losses in a number of engines is given in Table 38. Friction in Engines : Trans. A.S.M.E., 8-86, 10-10, 8-108, 9-74, 82, 7-639, 641, 1-153; Ewing, Steam Engine, p. 186; Ripper, Steam Engine, p. 275; Peabody, Thermodynamics, p. 430; Perry, Steam Engine, p. 270; Ripper, Steam Engine, 1-540; Heck, Steam Engine, 316-318. The efficiency of the fluid in the steam engine cylinder may be in- creased by (1) raising the boiler pressure, (2) compounding, (3) use of reheater-receiver, (4) steam jacketing the cylinders, (5) increasing the rotative speed, (6) superheating, (7) diminishing the back pressure by securing a more perfect vacuum. 168. Effect of Increased Steam Pressure. — A consideration of the Rankine and Carnot cycles indicates that theoretically the greater the temperature range the greater will be the efficiency. (See Table 36.) In the actual engine the temperature range is most readily increased by raising the boiler pressure, since the limit of the back pressure is practically fixed by the cooling medium in the condenser. The theoreti- cal gain resulting from increased pressure range is, however, very con- siderably affected by the increased losses due to cylinder condensation. Fig. 147 shows the results of tests made at the Armour Institute of Technology on an 8 x 10 automatic high-speed piston-valve engine, showing marked gain with increase of initial pressure up to a certain point when the condensation losses became sufficiently great to neu- tralize the advantage which would otherwise be gained. The following figures were obtained in tests of a small Willans engine, non-condensing, under different steam pressures : Initial Pressure, Gauge. Pounds Steam per I.H.P. Hour. B.T.U. per I.H.P. per Minute. 36.3 42.8 700 51.0 36.0 595 74.0 32.6 544 85.0 29.7 495 97.0 26.9 450 110.0 27.8 465 122.0 26.0 436 STEAM ENGINES 287 Referring to Table 36, it may be noted that both the theoretical and the actual efficiencies increase very slowly for pressures above 150 pounds. Practically, gain in efficiency due to increasing the pressure 6.2 6.1 S 6.0 f | 6.0 B 3 5.8 a 5.7 5.6 5.5 - V x •^ >,* ^ N^K y/ 4^ \ -* ps % V - _ • * "^ 48 47 W w 44 43 0, 75 Fig. 147. 85 90 95 100 105 Initial Gauge Pressure.Lb.per Sq.In. 110 115 120 Influence of Initial Pressure on the Economy of a Small, High-Speed, Non-Condensing Engine. above about 200 pounds is at the expense of increased first cost and maintenance and is only resorted to when small weight and space are the most important considerations. The range of pressures sanctioned by modern practice for different types of engines is as follows: Type of Engine. Simple slow speed Simple high speed Compound high speed, non-condensing Compound high speed, condensing Compound slow speed, condensing Triple expansion, condensing Quadruple expansion, condensing Range in Pres- sure (Gauge). 60-120 70-125 100-170 100-160 125-200 140-210 125-225 Average. 90 100 130 125 150 175 200 Steam Pressure : Trans. A.S.M.E., 4-88, 5-269, 6-572; Peabody, Thermo- dynamics, p. 248; Ripper, Steam Engine, p. 306; Engine Tests, Barrus, p. 258. 169. Receiver-Reheaters. — The receivers between the cylinders of multi-expansion engines are frequently equipped with heating coils as illustrated in Fig. 332, the function of which is to superheat the exhaust steam before delivering it to the cylinder immediately follow- 288 STEAM POWER PLANT ENGINEERING ing, with a view of reducing the losses occasioned by cylinder conden- sation. The coils are supplied with live steam under boiler pressure and may serve to evaporate a portion of the moisture or to actually superheat the steam supplied to the following cylinder. The question of the propriety of using reheaters is an open one, since reliable data relative to their use are meager and discordant. The conditions under which the few recorded tests were made are too diverse to warrant definite conclusions. Some show an appreciable gain in economy, others a decided loss. A reheater is of little value in improving the thermodynamic action of the engine, and is probably a loss unless it produces a superheat of at least 30 degrees F., and to be fully effective should superheat above 100 degrees F, (L. S. Marks, Trans. A.S.M.E., 25-500.) The effectiveness of the reheater will evidently be increased by the removal of the greater portion of the moisture from the exhaust steam before it enters the receiver. In the 5000-horse-power engine at the Waterside Station in New York it was shown that both jackets and reheaters, either together or alone, were practically valueless throughout the working range of load. {Power, July, 1904, p. 424.) Many similar cases may be cited which show no gain in economy with the use of the reheaters. On the other hand, with properly propor- tioned reheaters, the gain may be considerable and particularly with superheated steam. Practically all European engines operating with highly superheated steam are equipped with receiver-reheaters. In all cases the reheater effects a great reduction in the condensation in the low-pressure cylinders, but the resulting gain, considering the conden- sation in the reheater coils, may be little if any. In triple expansion pumping engines receiver-reheaters are found to effect an appreciable gain in economy, and practically all such engines are equipped with them. In electric traction work or where the load is a widely fluctuating one the reheater has been virtually abandoned. Apart from the consideration of fuel economy, all tests show a marked increase in the indicated power of the low-pressure cylinder (5 to 15 per cent), and to that extent it increases the capacity of the entire engine. (G. H. Barrus, Power, September, 1903, p. 516.) Receivers: Trans. A.S.M.E., 1-174, 178, 9-549; Ewing, Steam Engine, p. 222; Holmes, Steam Engine, p. 456; Spangler, Steam Engineering, p. 249; Whitham, Steam Engine, p. 395; Power, June, 1896, p. 20; ibid., Nov., 1905, p. 684. Receiver-Reheaters: Trans. A.S.M.E., 1-178, 17-509, 25-443; Power, Sept., 1903, p. 516; Am. Elecn., Oct., 1902, p. 480; ibid., July, 1904, p. 328; Eng. Rec., May 9, 1903, p. 496; Engineering, Aug. 8, 1902, p. 197; ibid., Aug. 21, 1902, p. 125; Eng. Rec, May 28, 1904, p. 690. 170. Jackets. — If the walls of the cylinder are made double and the space between is filled with live steam under boiler pressure, the STEAM ENGINES 289 cylinder is said to be steam jacketed. The function of the jacket is to reduce initial condensation by maintaining the temperature of the internal walls as nearly as possible equal to that of the entering steam. The heat given up by the jacket steam, and the resulting condensa- tion, is usually a smaller loss than would otherwise result from cylinder condensation. However, tests of numerous engines with and without steam jackets do not agree as to the conditions under which their use is profitable, the apparent gain ranging from zero to 30 per cent. According to Peabody, a saving of 5 to 10 per cent may be made by jacketing simple and compound condensing engines, and a saving of 10 to 15 per cent by jacketing triple expansion engines of 300 horse power and under. On large engines of 1000 horse power or more the gain, if any, is very small. (Peabody, "Thermodynamics," p. 400.) Other things being equal, the smaller the cylinder and the lower the piston speed the greater is the value of the jacket. Experiments show no advantage in increasing the jacket pressure more than a few pounds above that of the initial steam in the cylinder, and it is usual to reduce the pressure in the jackets of the second and succeeding cylinders of multi-expansion engines. (Ripper, " Steam Engine," p. 170.) To be effective, jackets should be well drained, kept full of live steam, and the water of condensation returned directly to the boiler. Pumping engines and other slow-speed engines running at practi- cally constant load are generally jacketed, but in street-railway work and in the majority of manufacturing plants carrying fluctuating load, jackets are not considered advantageous. Whatever may be the actual economy due to jacketing, there is no question but that the jacket greatly influences the action of the steam in the cylinders, and whether beneficially or not depends upon the design and construction of the engine. Unless otherwise specified, manufacturers usually build their engines without jackets. Steam Jackets: Trans. A.S.M.E., 1-175, 190, 2-198, 9-554, 11-141, 149, 328, 1038, 12-873, 462, 13-176, 14-1356, 15-779, 137; Golding, 6 Diagram, p. 39; Hutton, Heat Engines; Perry, Steam Engine, p. 369; Rankine, Steam Engine, p. 395; Reeve, Thermodynamics, p. 29; Ripper, Steam Engine, p. 166; Thurston, Manual of the Steam Engine, 1-598, 622; Peabody, Thermodynamics, p. 322; Heck, Steam Engine, p. 123; Am. Mach., Jan. 30, 1896, p. 126; Engr., Lond., April 21, 1905, p. 401 ; Engineering, Jan. 30, 1905, p. 829; Eng. Rec, April 16, 1898, p. 423; Power, Feb., 1898, p. 17; ibid., Feb., 1899, p. 9; Eng. Mag., June, 1898, p. 479, June, 1899, p. 496, Aug., 1905, p. 755. Increasing Rotative Speed: See High-Speed vs. Low-Speed Engines, paragraph 172. Compounding : See Compound Engines, paragraph 177. Reducing Back Pressure : See Influence of Condensing, paragraph 179. Superheating : See paragraph 181. 290 STEAM POWER PLANT ENGINEERING 171. Single- and Double-Acting Engines. — When steam pressure is exerted on only one end of the piston the engine is said to be single acting, and when exerted alternately on one side and the other is said to be double acting. For high speed, minimum wear and tear, and comparatively cheap construction the single-acting engine offers some advantages. The Westinghouse Standard and the Willans Central- Valve engines are typical of this class. Silent running at high speed is possible because the pressure on the crank pin is not reversed. The output is only half that of a double-acting engine of the same size and speed, but a much higher rotative speed is permissible, which some- what offsets this disadvantage. Single-acting engines have been operated successfully with speed as high as 1000 r. p.m., while double-acting engines seldom exceed 350 r.p.m. and that only for strokes less than 12 inches. Single-Acting Engines. — Comparison between Different Types of Engines ; Trans. A.S.M.E., 2-294. Economy of Single-Acting Expansion Engines: ibid., 3-252. Single- Acting Compound ; ibid., 12-275. Steam Distribution in Compound : ibid., 13-557; Ewing, Steam Engine, p. 371 ; Rankine, Steam Engine, p. 478 ; Hutton, Power Plants, p. 77. Double-Acting Engines: Hutton, Power Plants, p. 73; Rankine, Steam Engine, p. 50; Ewing, Steam Engine, p. 20; Le Van, Steam Engine, p. 240. 173. High- and Low-Speed Engines. — High rotative speed does not necessarily mean high piston speed. An 8 x 10 engine running at 300 r.p.m. has a piston speed of only 500 feet per minute, whereas a 36 x 72 Corliss running at 60 r.p.m. has a piston speed of 720 feet per minute. The classification " high speed " and " low speed " refers to rotative speed only, the former above and the latter below say 150 r.p.m. On account of the reduction of thermodynamic wastes, a high-speed engine should give theoretically a higher efficiency than the same engine at a lower speed, all other conditions being the same. The effect of speed upon economy is decidedly marked in engines and pumps taking steam full stroke. For example, tests of a 12 x 7J x 12 simplex direct-acting steam pump at Armour Institute of Technology showed a steam consumption of 300 pounds per I.H.P. hour at 10 strokes per minute, and only 99 pounds at 100 strokes per minute. (See Figs. 274 and 275.) Tests of engines using steam expansively, however, do not furnish conclusive evidence on this point, some showing a decided gain (Pea- body, " Thermodynamics," p. 425), others little or no gain (Barrus, " Engine Tests," p. 260). For example, a small Willans engine showed an increase in economy of 20 per cent in increasing the rotative speed STEAM ENGINES 291 from 111 to 408r.p.m. (Peabody, "Thermodynamics," p. 402), whereas the compound locomotives at the Louisiana Purchase Exposition showed a loss in economy for the higher speeds (Publication by the Penn- sylvania Railroad Company). On the other hand, a comparison of the performances of high- and low-speed Corliss engines shows little differ- ence in economy, and a general comparison between high- and low-speed engines furnishes little information, since nearly all high-speed engines are of a different class from the low-speed ones. High-speed engines are comparatively small in size, require larger clearance volume, and are usually fitted with a single valve. Rotative speed is limited by design, material, workmanship, and cost of subsequent maintenance. Speeds of 400 r.p.m. and more are not unusual with single-acting engines, whereas 300 r.p.m. is about the limit for double-acting machines with strokes over 12 inches in length. A comparison of tests of high-speed and low-speed engines in this country, irrespective of design and con- struction, shows the former to be less economical than the latter in most cases. In Europe high-speed engines are developed to a high degree of efficiency, and their performances are comparable with the best grade of low-speed engines. High-speed engines as a class have the advantage of being more compact for a given power, are simple in construction and relatively low in first cost; on the other hand, they are subject to comparatively rapid depreciation, excessive vibration, and are less economical in steam consumption. High- and Low-Speed Engines. — Effect of Speed on Condensation : Peabody, Thermodynamics, p. 424. Effect of Speed on Economy: Barrus, Engine Tests, p. 257; Trans. A.S.M.E., 7-397, 2-198. Limitation of Speed : Trans. A.S.M.E., 14-806. Effect of Speed on Economy: Ripper, Steam Engine, p. 317. 173-4. High-Speed Single-Valve Simple Engines. — This style of engine is made in sizes varying from 10 to 500 horse power. The cylinder dimensions vary from 4x5 to 24 x 24 and the rotative speed from 300 to 175 r.p.m. When ground is limited or costly and exhaust steam is necessary for heating or manufacturing purposes, the high-speed non-condensing engine is most suitable for horse powers of 200 or less, being compact, simple in construction and operation, and low in first cost. For sizes larger than this the compound engine would probably prove a better investment, except in cases where fuel is very cheap or large quantities of exhaust steam are to be used for manufacturing purposes. Small high-speed engines are seldom operated condensing, since the 292 STEAM POWER PLANT ENGINEERING gain due to reduction of back pressure is more than offset by the extra cost of the condenser and appurtenances. Engines are ordinarily rated at about 75 per cent of their maximum output. For example, a 12 x 12 non-condensing engine ^ running at 300 r.p.m., with initial steam pressure of 80 pounds gauge, is nor- mally rated at 70 horse power, though it is capable of developing 90 horse power at the same speed. The steam consumption of high-speed single-valve non-condensing engines at full load ranges from 27 to 50 pounds per indicated horse- power hour, depending upon the size of the unit and the conditions of operation. An average for good practice is not far from 30 pounds. With superheated steam a steam consumption as low as 18 pounds per horse-power hour has been recorded. Table 39 gives the steam consumption of a number of single-valve high-speed engines running condensing and non-condensing, and 49 47 ($) 43 §41 W '.39 Pu W37 M ■3 32 (?) ca > ^0 *> QL .«* lhA 50 R .p.im :., 1 S.P- 1.00 S§ k§H £ (a In, : 4Ll &0lip.]fc c, 1 s.v 80 ^31 Ki4r *>- ^"~: •§29 s Nt^ 250 R.P M-, :.s.p . 95 © ^v J.B1 I. i>_ -Si-i&O^B El .,1.5 3.P. 1 00 25 f^f MJg^u or.i P.M. , IS .p.u 6 23 21 23 50 75 100 125 Per Cent of Rated Load Fig. 148. Typical Economy Curves of High-Speed, Single-Valve, Non-Condensing Engines. Saturated Steam. Fig. 148 shows some of the results for different loads. The steam consumption is fairly constant from 50 per cent of the rated load to 25 per cent overload, but for earlier loads the economy drops off rapidly. The desirability of operating the engine near its rated load is at once apparent. The curves show a marked economy in favor of the larger cylinders, but the engines are not of the same make, and the conditions of operation are somewhat different. STEAM ENGINES 293 Fig. 148a. Assembly of Valve Gear; Typical Corliss Engine. Steam pipe | ">f Steam Flange .Corliss Steam Valve I I Throttle Valve Vanished Sheet Steel ^^_V* •j-a-H p, o •sq V O o '•ui -bg J9d 'sqi § 'amssajti ^pug w •aSrvBO 'sqi 02 'amssaij rei^nr- o Q W <1 t>-00eOt^Os00C0 CO iO CO CO CO CO CO CO CO "* O O CO IO Ui •* t*- OS Ui CO -tf © OS CO O CO IO 00 O CO O CO CO CO b- OS CO OS H T^l f>. -H^ T^l T^ ^ "^ ^ CO Ui OMffliflNiOOOO "* OCSC0I>©I>»Oi>C0 O ■* CO rH (M CO (N IN CO i-l CO 'O0"0000<0 OCOOOOOOOO CO CO»00©00000 COCOOSt^rHOLOGCO COrHOSiOCOCOO^t^ CO CO rH rH ^H CO OS CO CO Tff CO CO ■* CO CO 00 CO XXXXXXXXX X ■* ■* l> O CO Ui CO I> CO CO < rH CO rH rH rH rH CO rH sa- cs c3 • S Co •o-o 1 ^ O O - O O w c3 c3 c3 £££ CD oo ei • « ft r^l rW T -T«o o> f-SJcf|i rH-""£ P- Is ■"El .5 ° ° .to 93 03 >>££ • • I § ° - r- ococo COCO CO ^ "rj< "^ iO-* co acca rHCO co eoico COCO CO cococo t>C35 O OOOOS COlO O CO"*-* O rH t^COOO CO CO IO rH l>0^ So o H o r> o te; r- r> b CO00 OCOCO C0OTt< rHIOCO o o 00 OS CO •HKlOTt* XXX OOHncO rHOOrH CO W^ 03 03 ~ m 3 « X O 1) >>rj 03 ft-M>^> ■^^ CQQw'S w* co O «H (H O 3 3 c3 oO o > eg J§ § Ph£ s pmnpq rHW5-*CO»OrH cOt^cOOOst^ rtHHMHH COiOCOCOt^CO O t^ Ttl Tj< OS IO CO CO CO CO CO CO lOCOtOTf OOOOOC aS S «_03 ^03 (h (h o o OCOOrHOO O CO CO CO t^ OS COCOOOSCOIO rH rHCO CO rH OOOCOiOOO Ui CO CO ■<# CO Ui IO CO rH IO rH ■* CO CO >0 CO rH COOOOOO CO>OOOOX ■* TtH CO CO CO ■* xxxxxx CO «o - o 'HftS-C'O'C O O h 0= O OPhOOOQO STEAM ENGINES 297 The performances given in Table 39 are exceptional. It is not ad- visable to count on a better steam consumption for this type of engine than 30 to 35 pounds of steam per I.H.P. hour. tt 3 O \ • 60 n. \ Q. I \ TEST OF REEVES SIMPLE ENGINE ■ 50 oc a HADE BY PROF. R. C CARPENTER AND PROF. H. DIEDEHICH8 < 40 cr O ' \ \ o CD a X o 3 2 NON- CONDENSING 30 AVERAGE 1/8. P. 114 LB o CONDENSINC 20 Developed Horse-Power, < > 2 4 e i JO 1C )0 u >Q I' iO 160 ' 180 Fig. 149. Fig. 149 shows the effects of condensing on a typical single-valve high-speed engine. The gain in fuel economy may be only an apparent one, since the steam consumption of the condensing apparatus should be rightfully charged to the engine. When used in connection with heating plants or manufacturing plants requiring large quantities of exhaust steam the thermal efficiency is very high and may reach 85 per cent as against 22 per cent for the best compound condensing engine. In general when the requirements for exhaust steam are in excess of the steam consumption of a simple non- condensing engine a high-grade economical engine is without purpose. 175. High-Speed Multi- Valve Engines. — The steam distribution in a single-valve engine may give good economy for a very small range in load but be far from satisfactory for a wide range. This must neces- sarily be so since admission, cut-off, release, and compression are all functions of one valve, and any change in one results in a change of the others. To obviate the limitations of the single valve, many builders design engines with two or more valves. With a two -valve engine cut- off is independent of the other events, and with four valves all events are independently adjustable. In addition to the flexibility of the valve gear, the chief feature of the four-valve engines lies in the reduction of clearance volume which is made possible by placing the valves directly 298 STEAM POWER PLANT ENGINEERING over the ports. The valves may be of the common slide-valve or rotary type. As a class, four-valve engines are more economical than those having a less number of valves. The advantages and disadvantages of the four-valve over the single-valve engines may be tabulated as below. Advantages. Disadvantages. 1. Better steam distribution. 1. Increased number of parts. 2. Better regulation. 2. Increased first cost. 3. Reduced clearance volume. 3. Requires greater attention. 4. Less valve leakage. 5. Better economy. The steam consumption of a high-speed four-valve non-condensing engine varies from 22 to 35 pounds of saturated steam per horse-power hour, with an average not far from 27 pounds. With superheated steam the steam consumption may run as low as 18 pounds per horse-power hour. An exceptional performance for a simple unjacketed high-speed four- valve engine is that of Engine No. 17, Table 39. With initial gauge pressure of 125 pounds the steam consumption is 22.24 pounds per I.H.P. hour, corresponding to a heat consumption of 374 B.T.U. per I.H.P. per minute. 100 90 o< 60 50 40 o Comparative Economy of a and a ( B) Pour Valve High Speed Non Condensing Engine 15 x 14 Reeves Simple (A) 16 x 16 Memming Simple ( B ) \ \ \ \\ ^A^ 3 S> \B ' — c p —a— 10 20 30 40 50 60 70 80 90 Per Cent of Rated Load Fig. 150. 100 110 130 130 140 Fig. 150 gives a comparison between a single- valve and a four- valve high-speed engine, and though the engines differ slightly in size, the STEAM ENGINES 299 conditions of operation were comparable and the marked gain in economy of the latter over the former is apparent. Both perform- ances are exceptional, and a 10 to 15 per cent greater steam consump- tion may be expected in average good practice. As a general rule single- valve simple engines do not exceed 500 horse power in size for stationary work, whereas 1000 horse power is not an uncommon size for the multi-valve type. High-Speed Engines, General Description: Trans. A.S.M.E., 2-75; ibid., 17-117; Engr. U.S., Jan. 15, 1903; Engr., Lond., April 15, 1904, p. 379, April 29, p. 433, May 13, p. 478, May 20, p. 529. Proportions of High-Speed Engines: Trans. A.S.M.E., 8-191; Klein, High-Speed Engines; Hutton, Power Plants, p. 70, Thurs- ton, Stationary Steam Engines. Tests of Simple High-Speed Engines : Am. Elecn., April, 1901, p. 197, Dec, 1903. p. 581; Trans. A.S.M.E., 11-723, 18-795; Elec. World, May 20, 1903, p. '897, Sept. 10, 1904, p. 404, Oct. 1, 1904, p. 587, Feb. 17, 1906, p. 369; Engr. U.S., June 1, 1903, p. 416, Nov. 1, 1904, p. 758; Engineering, July 22, 1898, p. 116; Eng. News, Dec. 3, 1903, p. 493; Eng. Rec, July 6, 1901, p. 225; Machinery, May, 1903, p. 481; Power, Jan., 1904, p. 44, Nov., 1904, p. 651, Jan., 1905, p. 56; Stevens Indicator, Jan., 1900, p. 9; St. Ry. Jour., Oct., 1904, p. 673; Technology Quarterly, Sept., 1899, p. 255. 176. Medium and Low-Speed Multi- Valve Engines. — A comparison of tests of high and low-speed single-valve engines irrespective of design and construction shows the former as a class to be less economical than the latter. With four-valve engines there is no such disparity, and the high-speed type has shown just as good economy as the slow-speed class. For example, Engine No. 17, Table 39, with Corliss valves and a speed of 210 r.p.m., gives practically the same economy as Corliss engine No. 15 operating at 62 r.p.m. By far the greater number of simple multi- valve slow-speed simple engines are of the Corliss type. They range in size from 50 to 3000 horse power, with cylinders varying from 12 x 30 to 48 x 72. The smaller sizes with trip-valve gear run at 90 to 100 r.p.m., and the larger at 50 to 75 r.p.m. Without the trip gear, speeds of 150 r.p.m. are not uncommon, but at this speed they are usually classified as high-speed engines. Table 39 gives the steam consumption, condensing and non-con- densing, of a number of four- valve slow-speed simple engines. Engine No. 15 shows an unusual performance for a simple Corliss engine operating both condensing and non-condensing. With initial gauge pressure of 103.5 pounds, the minimum steam consumption is 21.5 pounds per I.H.P. hour for the non-condensing run and 16.5 pounds for the condensing run. This corresponds to 358 B.T.U. per I.H.P. per minute, non-condensing, and 302 B.T.U. per I.H.P. per minute, 300 STEAM POWER PLANT ENGINEERING condensing. The efficiency ratios are 78.0 per cent and 53.2 per cent respectively. The cylinder was jacketed. Attention is also called to the record of engine No. 22, Table 39, which is of the Sulzer type with four balanced poppet valves, heads and cylinder barrel jacketed. With 79 pounds initial pressure and a vacuum of 1.36 pounds absolute, the steam consumption is 15 pounds per I.H.P. hour, corresponding to a heat consumption of 275 B.T.U. per I.H.P. per minute. 177. Compound Engines. — Compound engines may be divided into three classes, tandem, cross compound, and duplex. In the tandem the two cylinders are end to end, in the cross compound side by side, and in the duplex one above the other. The tandem and duplex compounds have the advantage of (1) compactness for a given power, (2) less complication and fewer parts, and (3) low first cost. The crank effort is more variable than in the cross com- pound. In very large engines the low-pressure stage is generally divided between two cylinders of equivalent size to avoid an excess- ively large single cylinder and to distribute the crank effort. High- speed non-condensing compounds are ordinarily of the tandem type and are finding much favor in isolated station work, as in the power plants of tall office buildings where ground space is limited, though the duplex compound is sometimes used. The vertical or horizontal cross compound is generally installed in street-railway plants. Cylinder ratios for high-speed single-valve compound engines vary from about 1 to 2\ with 100 pounds pressure to about 1 to 3 with a pressure of 150 pounds, and for slow-speed condensing engines from 1 to 3 with 125 pounds pressure to about 1 to 4 with a pressure of 175 pounds. G. I. Rockwood recommends a ratio as high as 7 to 1, and a number of engines designed along this line have shown exceptional economy. A cross compound Corliss engine at the Atlantic Mills, Providence, R.I., with cylinders 16 and 40x48 (ratio 6.128 to 1) gave the low steam consumption of 11.2 pounds of steam per I.H.P. hour, corresponding to a heat consumption of 222 B.T.U. per I.H.P. per minute. The 5500-horse-power engines of the New York Edison Company have a cylinder ratio of 6 to 1. The great majority of corn- pound engines, however, have cylinder ratios of 4 to 1 or less. The 8000-horse-power engines of the Interborough Rapid Transit system have a ratio of 4 to 1, and the 4000-horse-power units of the Metro- politan Elevated Company, New York, a ratio of 3.5 to 1. STEAM ENGINES 301 Fig. 150a. 3500 K.W. Vertical Cross-Compound Corliss Engine as Installed at the Power House of the Twin City Rapid Transit Co., Minneapolis, Minn. 302 STEAM POWER PLANT ENGINEERING Fig. 150b. 7500 K.W. Vertical-Horizontal Double Compound Engine as Installed at the 59th Street Station of the Interborough. (Manhattan Type.) STEAM ENGINES 303 The respective advantages and disadvantages of compounding may be tabulated as follows: Advantages. 1. Permits high range of expansion. 2. Decreased cylinder condensation. 3. Decreased clearance and leakage 4. Equalized crank effort. 5. Increased economy in steam consumption. Disadvantages. 1. Increased first cost due to multi- plication of parts. 2. Increased bulk. 3. Increased complexity. 4. Increased wear and tear. 5. Increased radiation loss. The ratio of expansion for a multi-expansion engine is usually taken to be the product of the ratio of the volume of large to small cylinder divided by the fraction of the stroke at cut-off in the high- pressure cylinder. For example, a compound engine with cylinders 24, 48 x 48 cutting off at ^ in the high-pressure cylinder has a nominal ratio of expansion of 4 -j- | = 12. The number of expansions at rated load in compound condensing engines varies widely, ranging from 10 to 33, with an average not far from 16. The steam consumption shown by tests of a number of compound engines using saturated steam, condensing and non-condensing, is given in Table 40. For tests with superheated steam see Table 43. <* V \ \ \ Belative Economy of a Simple and Compound Non-Condensing High Speed Engine \ \ \ \\ X ^ v -. 2 ft ^^ ie^ 5 ._— i »■ — ^ Ct mpound 1 30 30 40 50 60 70 80 90 100 120 140 Developed Horse Power (B. H. P.) Fig. 151. 160 180 Fig. 151 shows the relative economy under comparable conditions of a high-speed simple and a high-speed compound engine, both run- ning non-condensing and using saturated steam. The advantage of the compound at full load and overload is very marked, though its 304 STEAM POWER PLANT ENGINEERING economy drops off rapidly at light loads and may be less than that of the simple engine. Fig. 152 shows the relative economy of two compound Corliss engines running condensing and non-condensing, both using saturated steam. 1 A 21,41 x 30 Compound B 20,40 x 42 Compound 20 Ac< 18 sji £c Ofld easing 16 14 Be N L 12 — ^Sgdensing 300 400 500 600 700 800 900 1000 1100 1200 1300 1400 Indicated Horse Power Fig. 152. It should be borne in mind that the object of compounding is to permit the advantageous use of high pressures and large ratios of expansion. Under proper conditions compounding may increase the economy at rated load about 20 per cent for non-condensing engines and 30 per cent for condensing engines. An exceptional performance of a single-valve high-speed non-con- densing compound engine is that of engine No. 20, Table 40. With initial gauge pressure of 128 pounds the steam consumption is 22.3 pounds per I.H.P. hour, corresponding to a heat consumption of 376 B.T.U. per I.H.P. per minute. One of the best performances of a multi- valve high-speed compound non-condensing engine is that of engine No. 14, Table 40. With initial pressure of 175 pounds gauge the steam consumption at full load is 17.17 pounds per I.H.P. hour, corresponding to a heat con- sumption of 291 B.T.U. per I.H.P. per minute. The 8000-horse-power vertical cross compound Corliss engines of the Interborough Rapid Transit system (No. 6, Table 40), probably hold the record for economy for compound engines without jackets and reheaters, using saturated steam. With initial pressure of 175 pounds gauge and absolute back pressure of 2.2 pounds, the steam consumption is 11.96 pounds per I.H.P. hour, corresponding to a heat STEAM ENGINES 305 consumption of 220 B.T.U. per I.H.P. per minute. In estimating aver- age practice it would be safe to add 10 per cent or 20 per cent to the steam consumptions given in Table 40. Lb 27 .2 22 Jj.19 "§ 18 Oh* 17 > * 15 12 1 \ \ \ fici >nc} ^-or- Net-E J 1,-B Dividec \ "by ta P. €s& it \C^ 1 ~x \s v C ° 94 92 90 SS S6 84 82 SO ■4?° ggs .W; iter / S Wa ter ; igji 1-HburisTet Ougp *rf* / -Q-. \1 . H our Gt° ssC \ / / \ 4 oun ds"n Vater p IT I.JT.i our - i '■ 500 1000 1500 2000 2500 3000 Gross K.W. Output 3500 4000 4500 5000 Fig. 153. Economy Test of the 5500-Horse-Power Three-Cylinder Compound Engine and Generator at the Waterside Station of the New York Edison Co. Fig. 153 illustrates the performance of the 5500-horse-power three- cylinder compound engine at the Waterside Station of the New York Edison Company. The best economy is 11.93 pounds of steam per I.H.P. hour, corresponding to a heat consumption of 221 B.T.U. per I.H.P. per minute. Compound Engines. — Best Load for Compound Engine : Trans. A.S.M.E., 18-674. Cylinder Proportions for Compound and Triple Expansion Engines: Trans. A.S.M.E., 21-1002, 16-762; Engr. U.S., Sept. 1, 1906, p. 586; Eng. News, March 2, 1899, p. 137; Eng. Rec, Jan. 7, 1899, p. 122; Power, June, 1904, p. 47. Laws of the Average Simple vs. Compound Engines under Variable Load: Am. Mach., Sept. 27, 1900, p. 927; Non-condensing Compound Engine for Office Buildings: Eng. Rec, June 18, 1898, p. 45. Economical Use of Steam in Non-condensing Engines: Eng. Mag., May, 1898, p. 213, July, 1898, p. 603. Compound Engine Tests: Trans. A.S.M.E., 24-1274, 25-264; Engr. Lond., 99-546; Eng. News, Jan. 11, 1906, p. 44; Eng. Rec, April 16, 1898, p. 431, June 4, 1898, p. 1; Nov. 18, 1899, p. 579; Sibley Jour., May, 1901, p. 346; St. Ry. Jour., 27-41. 178, Triple and Quadruple Engines. — Triple and quadruple expan- sion engines are in general use where the load is practically constant, as in marine and pumping-station practice, but have been abandoned in street-railway work and in plants where the load fluctuates widely, •iH 'd'H'a J»d lt?oo jo -sqi •JU80 J8d 'opBH Aouaioing- •Sua ^paj-^d: '"«ii\[ •^uaQ aad '.fouaping; x'BuiJsqx •nij\[ jad •J'Hl'j^ 'II' 1/8. 'W ' J'H'I J9d ureais jo -sq^ ' .19512 Al P^M -duiax Ol p9JJ9J9H "* lO I 5 * CO CS * IQ o © o © LO g CO X CO s 1 00 •3 eo o g § Ol 01 § 8 -* Tf CD OS 1 1 oo CO CO § © to CO © © 3 3 OS s § CN cn >* CO LO »o CO Ol © Ol o CN OS Ol LO 00 CN t- CD a CO io •xopaj bjo a Sd Sfl, Ph - 60§ O^ § s S 3 rt CN rH ® co 00 5 2" bo = 3 S -' oX) • W *?W 3 Oil • 53.2030 S as Pi oP-r U3 rHCN 3S OS OS 0000 00 OS t- 00 M 01 01 3 wos II 51 § 8 8S oeo t-os H CM 01 CNCN 01 00 OS s 00 wiro 01 © CNcq IH^H * CD X X M X X « X '3 X oT CO a § s s< 01 g s 01 I S CN c > C <• ccft | 5 ha w w H r ! H«s r s s g " bX>, a ce (« cot A? h go <1 -< t3 c ^ ^ > w 02 CD 2 a 5& II ^ c Sh "$ H H HP- PQPh > ... 05 ^ S •£ >' aT g (P < . ^^ 4) "So 3 -/. Si O >r1 ~ -d 02 s-Chalmers E ork Subway. ss-Compound ntic Mills, Pro vitt Pumping I lie, Ky. 3 & Sargent C agar Retinery, ning Four-Val Home VorHna.l ' . ft Si. 5 OS^ M. S^2 od ct >.B ■525 e8 ^ < H^ X : HH co t^ 00 CO lO CO ^< CN CO OS 2 000000 ©OS 00 CO loeo'co x©x©id© cSiico TfcocNioo'i >o °. .'°. tj< t-0000 00 00 t-.l3t-; CNCNCN CiTJ^^'^'H 2 SI 00 © ■* «o ia ^-SCN CNcMi-ieNCNcO ~ ©'o'^" t-'cD'cNCNHS^ H P^PhS £Sft si sf O O sjo ssf CD 32 ^ O o« ce cS bfl ft^-i =« , OS i-i • *S £? 52 >» - « ""• n o 75 aj "11-= bij ^-"E ^02^ ® ® .; - C °S a o c--t c cS o> ^ PhPhWWWH 02*^ °.|H O . OK ^Z 55""^^ cj M P5 ^^pa cOce«« «Sft§©^ . kfl v rb> 5 ®»2y S ® ° «U fi > •£ a « ? 02 -^ m * a sa 5 2 >».L ®^ 3 ea SO* 3 ^coMPSQrr S S-SS: SSSSol^" STEAM ENGINES 307 in favor of the two or three-cylinder compound. The best economy on a heat-unit basis ever recorded for an engine using saturated steam was that of the Nordburg quadruple pumping engine at Wildwood, Pa., which gave a consumption of 12.26 pounds per I.H.P. hour and a heat consumption of 186 B.T.U. per I.H.P. per minute reckoned above the feed-water temperature.* The Allis triple expansion pumping engine at Chestnut Hill holds the record for saturated steam consumption, 10 pounds per I.H.P. hour, and its exceptional performance of one developed horse power per 1.09 pounds of coal has, perhaps, never been excelled. An inverted vertical marine cross compound engine, 21 and 36 x 36, built by Cole, Marchent & Morley, Bradford, England, holds the record for superheated steam consumption, 8.58 pounds per I.H.P. hour. (Table 47.) On the heat basis (192 B.T.U. per I.H.P. per minute), however, it does not equal the performance of the Nordburg engine. The above efficiencies have been exceeded by the binary vapor engine at Berlin; but this belongs in a class by itself and should hardly be compared with the ordinary form of steam engine. (See paragraph 182.) Triple Expansion Engines. — Cylinder Proportions for Triple Expansion Engines : Trans. A.S.M.E., 21-1002, 10-576. Economy of Triple Expansion Engines : Trans. A.S.M.E., 8-496. 179. Effects of Condensing. — The effect of the condenser upon the power and economy of engines is indicated in Table 41. The curves in Figs. 154 and 155 were plotted from tests made by Professor R. L. Weighton on a 7, 10^, 15J x 18 triple expansion engine at Durham College of Science, Newcastle-on-Tyne. The straight line shows how the mean effective pressure would vary with the degree of vacuum if the power increased directly with the reduction in back pressure. The curved line shows the actual M.E.P., which increases almost along the theoretical line up to a 10-inch vacuum, from which point on the increase is less marked. At 26 inches the actual M.E.P. reaches an apparent maximum. These figures are not applicable to all engines but give a good idea of the limitation of the vacuum with the reciprocating engine. The gain in steam consumption due to the condenser does not indicate a corresponding gain in heat consumption. For example, Engine No. 2, Table 41, shows an apparent gain in steam consumption, due to condensing, of 12.5 per cent, the temperature of the feed water returned to the boiler being 120 degrees F. With a suitable heater the exhaust of the non-condensing engine would be capable of heating * Replaced in 1905 by a Riedler pumping engine on account of high maintenance cost. 308 STEAM POWER PLANT ENGINEERING 42 41 o ~ 40 sr S 39 P. S ® £ 36 !■ F-l |. fc 31 £ 30 ( 29 88 # X* 1 pS* (> O /> o - Inc rease Trip inPc leEx wer ] pansi Duet on Ei ) Vac lgine uum 10 12 14 16 18 Vacuum in Inches of Mercury 20 22 24 26 28 Fig. 154. 19 < U t O <■ W 1R S. a 380^ a> 370 £ HI 360 £ o 350^ 340 a 330 S 320 S B 310 Sj 300g 05 290 \ I acres tse ir Tripl Eco e Ex lomj >ansi due on E toV igixa icuu n Oh a 17 se&_ a> N ^ > i GO o .216 >w < V rf ^ 1 S& ^B/r U.? P^ j&$ V^ 15 ■^^^ ^ ^ ^ J-. *.p? -Er.- ~5 ' — "i D C 2 i 4 t i 1 IS r acuu J 1 min] F i 1 [nche IG. 1 3 1 3 of 2d 55. 3 2< ercur ) 2 y 2 & I. 2< i 2J j 3C STEAM ENGINES 309 the feed water to 210 degrees F. The non-condensing engine should therefore be credited with 210 - 120 or 90 heat units per pound of steam used, or, in round numbers, 9 per cent. The difference between 12.5 per cent and 9 per cent, or 3.5 per cent, represents the net gain in favor of condensing provided the power necessary to create the vacuum is ignored. Actually the steam consumption of the condenser pumps might be equal to or greater than 3.5 per cent of the steam generated and the net gain becomes zero or even negative. Referring to Fig. 155, plotted from tests of the 7, 10i, 15J x 18 triple expansion engine mentioned above, the solid lines show the feed-water consumption per I.H.P. hour and the broken line the heat units consumed per brake t a & 210 I 3 235 1100 12.50 225 12.00 Recta Lb .-10 Lb. . Vacialn .-2i TT \ \ V \ L V \ > \ i V b S| \ s s V ,\ s V o N ^s V \ *5 - — — -. .. _ ._ .. ti .. V te r F er U.P. Ur rv u.vin g a rs ° p. FT 1 r, E s s; -o ^ • t er H .Varying '-. L ?}« at ;r pe H .PHr.Vai T i V 8. ic.l'n J& B. ['. J' p ;r Hi. b\J So ur \ ar ric B Re il re ST 1 ! 1 s 15 Lb 24.5 20 Lb. <*5 25 Lb. 26.5 80 Lb 26 35 Lb 40 Lb . 20.5 27 Fig. 156. Performance of 5500 H. P. Engine at Waterside Station of New York Edison Company. horse power per minute measured above the hot-well temperature. The engine efficiency, based upon the water consumption, increases as the vacuum increases, reaching a maximum between 26 and 28 inches, whereas the heat-unit curve gives the maximum between 20 and 21 inches. Between 22 and 28 inches the heat-unit curve shows a rapid falling off in economy. Tests of the 5500-horse-power engine at the New York Edison Company's Waterside Station showed that increasing the vacuum from 25.3 to 27.3 inches decreased the water rate only 0.06 pound per I.H.P. {Power, July, 1904, p. 424.) The results are illustrated in Fig. 156. In most cases, and particularly with large 310 STEAM POWER PLANT ENGINEERING compound engines, the net gain due to condensing is considerable, but the feed-water temperatures and power consumed by the auxiliaries should be taken into account.* Fig. 149 shows the effect of vacuum on the steam consumption of a small high-speed simple engine, and Fig. 152 of a cross compound Corliss. (See also paragraph 210.) TABLE 41. EXAMPLES OF THE EFFECT OF CONDENSING ON THE ECONOMY OF RECIPROCATING ENGINES. Non-Condensing. Condensing. Increase Due to Condensing. O) CD © & u r 8 . o 1 :1s el if Pi v CD CD o 4 M K Steam Con- sumption, Pounds per H.-P. Hour. CD be 3 . o a S £ Back Pressure, Pounds per Square Inch Abs. 5r! T3 cd a) £ a Ph 13 <» -S: O M Steam Con- sumption, Pounds per H.-P. Hour. §3 a i— i 1 2 3 4 5 6 7 8 9 10 11 147 148 126 67.6 103.8 114 96 118 75.9 62.5 186.7 54.7 540 83 209 177.5 160 120 267 310 451 40.4 19.2 19.3 23.8 28.9 22.1 31 23.9 23.24 25.6 30.1 18.7 149 147 130 67 103.8 114 96 119 79 63.6 184.6 1.6 4 7.4 4.5 1.2 "4' 4.2 6.4 7.8 1.6 83.4 116 213 155 168 145 276.9 336 444 29.8 14.8 16.9 19.1 22 16.5 27 19.4 16 20.5 23 12.7 52.5 * 39.8 1.9 * 2 20.8 3.7 8.7 * * 25 12.5 19.7 23.5 25.1 12.9 18.8 31 19.9 23.6 32 Cut-off changed for best economy. 7, 10J, 15i x 18 triple ; Eng. News, Aug. 21, 1902, p. 127. 17, 27 x 24 Westinghouse marine, non-condensing ; Power, August, 1903. 1, 18 x 10 Buffalo tandem compound ; Elec. World, Sept. 10, 1904, p. 404. 18 x 30 four-valve (slide) ; Engine Tests, Barrus, p. 88. 21, 65 x 43.31 Corliss ; Peabody's Thermodynamics, p. 382. 12 x 12 Reeves simple ; Elec. World, Oct. 1, 1904, p. 587. 18 x 48 simple Corliss ; Peabody's Thermodynamics, p. 354. 14, 28 x 24 two- valve (slide) ; Engine Tests, Barrus, p. 175. 17 x 24 two-valve; Engine Tests, Barrus, p. 70. 28 x 36 Corliss ; Engine Tests, Barrus, p. 97. Willans triple expansion central valve engine ; Peabody, Thermodynamics, p. 406. 180. Throttling vs. Automatic Cut-Off. — The action of the gov- ernor in the throttling engine is shown by the superposed indicator cards (Fig. 157) taken between zero or friction load and maximum load. The effect of throttling is to reduce the pressure during admis- sion, but does not change the point of cut-off or other events of the stroke. The steam may be partially dried or even superheated by throttling, thus tending to reduce cylinder condensation. Initially dry saturated steam at a pressure of 125 pounds gauge would be super- * See Power, Feb. 23, 1909, p. 381. STEAM ENGINES 311 heated about 12 degrees in expanding through a throttle to 90 pounds, or if it contained initially 2 per cent moisture would be perfectly dried in expanding to 40 pounds. (See Table 42.) Friction through the valve also tends to dry the steam. Thus with very light loads the superheat may be decidedly appreciable. The possible gain due Fig. 157. Typical Indicator Cards. High-Speed Throttling Engine. to decreased cylinder condensation is to some extent offset by incom plete expansion. The best efficiency for a given load is realized b} a proper compromise between cut-off and initial pressure. Experi- ments made by Professor Denton (Trans. A.S.M.E., 2-150) on a 17 x 30 non-condensing double-valve engine showed the most economical results with J cut-off for 90 pounds pressure, J cut-off for 60 pounds, and tVq for 30 pounds. The average throttling engine does not give close regulation, the governor usually lacking sensitiveness. Tests show the economy to be better than that of the automatic engine on light loads, and the crank effort more uniform. Fig. 158. Typical Indicator Cards. High-Speed Automatic Engine. The indicator cards shown in Fig. 158 were taken from a single- valve high-speed automatic engine operating between friction load and maximum load. The mean effective pressure is adjusted to suit the load by the automatic variation in the cut-off, the initial pressure remaining the same. Since the cut-off is controlled by the action of the governor on the single valve, all other events of the stroke are 312 STEAM POWER PLANT ENGINEERING likewise cnanged. With a four-valve engine the variation in cut-off does not affect the other events. The chief advantage of the automatic over the throttling engine lies in its sensitive regulation, and while, in general, it gives a lower steam consumption than the throttling engine, this is probably in most cases due to superior construction and not to the method of governing. TABLE 42. SHOWING THE INITIAL PER CENT OF MOISTURE THAT WILL BE EVAPORATED IN THROTTLING FROM A HIGHER TO A LOWER PRESSURE. Based on Marks' and Davis' Steam Tables. Final Pressures. I nitial Pressure, Absolute 80 85 90 95 100 0.45 0.59 0.74 0.88 1.06 1.23 1.43 1.66 1.91 2.20 2.53 2.92 3.40 4.01 105 110 115 120 80 0.13 0.26 0.40 0.55 0.71 0.89 1.09 1.32 1.56 1.85 2.18 2.56 3.04 3.65 0.24 0.37 0.52 0.66 0.83 1.01 1.21 1.44 1.68 1.97 2.30 2.69 3.16 3.78 0.36 0.49 0.64 0.78 0.95 1.13 1.33 1.56 1.80 2.10 2.42 2.82 3.29 3.90 0.55 0.70 0.84 0.99 1.16 1.34 1.54 1.76 2.02 2.31 2.64 3.03 3.51 4.13 0.65 0.79 0.93 1.08 1.25 1.44 1.64 1.86 2.12 2.41 2.74 3.13 3.61 4.23 0.74 0.88 1.03 1.18 1.34 1.53 1.74 1.96 2.M 2.51 2.84 3.23 3.71 4.33 83 75 0.14 0.28 0.43 0.59 0.77 0.97 1.19 1.44 1.72 2.05 2.44 2.90 3.51 0.97 70 . . 1 12 65 60 55 50 45 40 35 30 25 20 1.26 1.44 1.62 1.82 2.05 2.30 2.60 2.93 3.32 3.80 15 4.43 Final Pressures. Initial Pressure, Absolute 125 130 135 140 145 150 155 160 165 80 0.91 1.05 1.19 1.34 1.52 1.70 1.90 2.13 2.39 2.68 3.01 3.41 3.88 4.51 0.99 1.13 1.27 1.43 1.60 1.78 1.99 2.21 2.47 2.77 3.10 3.49 3.97 4.60 1.08 . 1.21 1.36 1.51 1.68 1.86 2.08 2.30 2.55 2.85 3.18 3.58 4.06 4.70 1.15 1.28 1.43 1.59 1.76 1.94 2.15 2.38 2.63 2.93 3.26 3.66 4.15 4.78 1.22 1.36 1.50 1.66 1.83 2.02 2.22 2.45 2.71 3.01 3.34 3.74 4.22 4.86 1.29 1.43 1.58 1.73 1.90 2.09 2.30 2.52 2.78 3.08 3.41 3.81 4.30 4.94 1.35 1.49 1.64 1.79 1.96 2.15 2.36 2.59 2.84 3.14 3.48 3.88 4.37 5.01 1.41 1.55 1.70 1.85 2.03 2.21 2.42 2.65 2.91 3.21 3.55 3.96 4.45 5.09 1.48 75 1.62 70 1.77 65 1.93 60 2.10 55 50 45 2.29 2.50 2.73 40 2.99 35 3.29 30 25 20 15 3.63 4.04 4.53 5.17 STEAM ENGINES 313 The following performances of a Belliss 250-horse-power high-speed condensing engine fitted with both automatic and throttling govern- ing devices give results decidedly in favor of the throttling engine. (Pro. Inst, of Mech. Engrs., 1897, p. 331.) Percentage of load .... Electrical horse power. Steam per I.H.P. hour. Automatic Cut-Off. Throttling. 100 62.5 33 25 100 62.5 33 213 132 77.8 53 213 132 77.8 22.5 22.9 28.5 34.3 21 21.7 25.6 25 53 28.4 Some of the comparative advantages and disadvantages of the automatic and throttling engines are as follows: Automatic. 1. Sensitiveness of regulation. 2. Increased ratio of expansion. 3. Low terminal pressures. Throttling. Advantages. 1. Low first cost. 2. Crank effort more uniform. 3. Reduced cylinder condensation. 4. Simplicity of regulating device. 1. Increased cylinder condensation. 2. Greater variation in crank effort. 3. Complicated valve gear. 4. Low economy at very early loads. 1. Low ratio of expansion. 2. High terminal pressure. 3. Low initial pressure at early loads. 181. Influence of Superheat. — (See also paragraph 103.) Table 43 gives test results for several different types of engines employing super- heated steam. These figures may be compared with the perform- ances of engines using saturated steam as given in Tables 39 and 40. A decided gain in economy is shown in favor of superheat for single- cylinder engines. With compound engines the advantage is not so apparent, while triple expansion engines show the least gain. Tables 44 to 46 show the effect of superheating on simple, compound, and triple expansion engines. (Proc. A.S.M.E., September, 1907.) As far as steam consumption is concerned, most engines show greater economy with superheated than with saturated steam, but the gain in thermal efficiency is not so marked, and when the economy is measured in dollars and cents per developed horse power, taking all things into consideration the gain is still further reduced and in many cases com- pletely neutralized. 314 STEAM POWER PLANT ENGINEERING Fig. 159 gives the results of a series of tests made on a number of Belliss & Morcom engines using superheated steam. (Pro. Inst, of Mech. Engrs., March, 1905, p. 302.) The engines were from 200 to 1500 kilowatts capacity and were tested at full load. It is noticeable that the curves all converge to a single point and will meet at about 400 100 200 Superheat, Deg. F. Fig. 159. Effect of Superheat on Steam Consumption. degrees F. The results show that if sufficient superheat is put into the steam all engines of whatever size are equally economical. Fig. 160 shows the relationship between degree of superheat and the heat consumption at various loads for a 300-horse-power Belliss & Morcom high-speed triple expansion engine. (Pro. Inst. Mech. Engrs., March, 1905, p. 303.) It will be noted that the variation in heat consump- tion at different percentages of load becomes less marked as the degree of superheat increases. With superheat of 350 degrees F. the heat consumption from \ load to full load is practically constant. . ureo?g po^oqjodng + oq; jo ;reoH - oodg •j[ *Soq; 'uoissnupv %v ureoiig 'duiox 'JL '%3Q. 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CD 43 v. b "° .. oJ3 O -0 d a. fl '^ ^ a si 2 -0 b£ it fa « c ^T3 S 5 cd O 2£ Q C *> § ?5W Q U STEAM ENGINES 319 Fig. 159a 3000 H.P. Sulzer Engine Designed for Highly Superheated Steam. 320 STEAM POWER PLANT ENGINEERING 3,. : to J a J : KJ d ft -^ J -d STEAM ENGINES 321 Table 47 gives the results of a test made on a 21, 36 x 36 inverted vertical marine engine. (Engr., Lond., June 2, 1905, p. 546.) TABLE 47. PERFORMANCE OF 21, 36 X 36 INVERTED VERTICAL MARINE CROSS COMPOUND ENGINE. {Engr., Lond., June 2, 1905, p. 546.) Pressure on boiler side of throttle valve, gauge Temperature of steam on boiler side of throttle, degrees F Degrees superheat of steam on boiler side of throttle, F Temperature of steam at admission, F. . . Degrees superheat of steam at admis- sion, F Vacuum, inches of mercury, absolute. . . I.H.P M.E.P. referred to L.P. cylinder Revolutions per minute Pounds of steam per I.H.P. hour B.T.U. per I.H.P. per minute See footnote * B.T.U. per I.H.P. per minute, perfect engine Thermodynamic efficiency Efficiency ratio Equivalent evaporation of saturated steam reckoned from hot well Temperature of hot well 117.5 117.5 117.5 117 114.5 743 738 749 751 732 395 390 401 403 384 601 590 569 580 558 253 242 221 232 210 3.4 3 2.22 1.82 1.82 481.3 461.1 347.5 333.5 258 26 29.9 18.8 18 14.4 100.6 100.7 100.6 100.7 100.7 9.09 9.26 8.88 8.68 8.74 192.2 201.7 197.6 194 194 187 189 181 179 179 142.4 142.5 130.2 126 128.5 21.4 21 21.4 21.8 21.8 72 72 66 65 66 10.63 10.81 10.38 10.07 10.12 102 101 78 71 64 114.5 726 378 550 202 1.7 145.52 7.87 100.7 8.58 192.1 175 128 22 67 10.03 70 * B.T.U. per I.H.P. per minute based on latest (May, 1908) values for specific heat of super- heated steam. The relationship between the weight of steam consumed per I.H.P. hour and the equivalent heat consumption of a 250-horse-power tan- dem compound Van Den Kerchove engine is illustrated in Fig. 161. The performances of engines using superheated steam should be expressed in B.T.U. per I.H.P. per minute or the equivalent, as the steam consumption alone gives no idea of the true heat consumption. 182. Binary- Vapor Engines. — A consideration of the Carnot or Rankine cycles shows that theoretically the efficiency of the steam engine may be increased by raising the temperature of the steam supplied or by lowering the temperature of the exhaust, that is to say, by increasing the range. Superheated steam development has prac- tically determined the upper limit, and economical practice indicates a vacuum of about 26 inches, corresponding to 126 degrees F., as the average lower limit for most efficient results from a commercial stand- point. 322 STEAM POWER PLANT ENGINEERING • 20 \\ IS I I — Satnr 16 I pteu StP«T» \ - S"Pe rheat.-50 F. 14 \ a •« -loc °F. \ r, " -150 °F. \ v v ,, » -200 'F. 12 (( " -250 3 F. «. " -300 D F. " -350 |f. 10 ___^ 40 60 Per Cent of Bated Load. loo Fig. 160. Effect of Superheat on Steam Consumption. ( 350 245 13- £> • 940 '§^ 0/> ? 235 m ii. < 5 ^ 3*i ^ 5.CL ^ ^ W 10 o^ > •P f 9,90 ^N K 9,15 y <^ fa. >D 910 '>05 8 Influence of Superheat on Economy (250 H.P. Tandem Comp'd Van Den) Kerchove Engine 900 195 i 190 K )0 2( K) a 30 Degrees of Superheat, Fahrenheit Fig. 161. STEAM ENGINES 323 In the binary-vapor engine the working range has been considerably- increased by substituting a highly volatile liquid, as sulphur dioxide, for the water which is ordinarily used as the cooling medium in the surface condenser. The S0 2 in condensing the exhaust steam is itself vaporized and the vapor, under a pressure of about 175 pounds per square inch, used expansively in a secondary reciprocating engine. The exhausted S0 2 is discharged into a surface condenser in which it is liquefied by cooling water much the same as in refrigerating practice and used over and r ! -To AicPump F SO Vaporizer and Steam Condenser l87#A1jsolutel48 F SOgVapor 1 1 " - H .7.6° Liquid S0 2 Tank 4,2rfABsJA 155 ° p ..!l L_ 56.8I.H.P. SO2 Cylinder 42.4I.H.P. r-' - c L.P. Cylinder 72°F J A i I I B l.P. Cylinder 43.2 1.H.P. *—j ll _J 49.8 F Circulating Water Inlet 62.7 °_ Circulating Water Outlet A [ G 49.5* kbsolute so„ Condenser \f H-P. 68.6I.H.P. P— • M Cylinder _590°F | 173 # A1 Steam Inlet jsolute 143.5 R.P.M. Fig. 162. Diagram of Binary- Vapor Engine. over again. Referring to Fig. 162, which illustrates diagrammatically a binary-vapor engine at the Royal Technical High School, Berlin: A, B, and C are the three steam cylinders of an ordinary triple expan- sion engine and D the S0 2 cylinder. All four cylinders drive a common crank shaft E. F is a high-pressure surface condenser which acts as a vaporizer for the S0 2 and a condenser for the steam. G is a surface condenser which serves to condense the S0 2 vapor. H is a liquid S0 2 tank. The operation is as follows: Highly superheated steam enters the high-pressure steam cylinder at I and leaves the low-pressure cylinder at /, just as in any steam engine. The exhaust steam enters chamber F and is condensed by the liquid S0 2 passing through the coils. The condensed steam and entrained air are removed from the chamber by a suitable air pump. The steam in condensing gives up its latent heat to the liquid S0 2 and causes it to vaporize. The S0 2 vapor passes from the coils in chamber F to the S0 2 engine D and performs work. The exhausted S0 2 vapor flows from cylinder D to chamber G, and is condensed by cooling water flowing through a series 324 STEAM POWER PLANT ENGINEERING of tubes. The liquid S0 2 is collected in liquid tank H and thence is pumped into the coils in vaporizer F. The approximate temperatures and pressures at different points of the cycle are indicated on the diagram. A number of experiments made by Professor E. Josse in the labora- tory of the Royal Technical High School of Berlin on an experimental plant of about 200 horse power gave some remarkable results. A few of the tests made with highly superheated steam gave the following average figures : I.H.P. (steam end) 146.4 Steam consumption per I.H.P. hour 12.8 I.H.P. (S0 2 end) , 52.7 Percentage of power of S0 2 engine 35.9 Steam consumption per I.H.P. hour of combined engine 9.43 When operating under the most satisfactory conditions a perform- ance of 8.36 pounds of steam per I.H.P. hour was recorded, correspond- ing to a heat consumption of 158.3 B.T.U. per minute, which is the best recorded performance to date (1907) in the history of steam- engine economy. 4000 3500 U" ^ 3000 & & 1* ^ 2500 ^ .--- ,*• *-* 2000 .^ ^^ 1500 •^ M £? 1000 tf^ ? V s -- — ( v-r- ■=1 J= =1 1= tJ L = = 50 +VbJ i.*. 50 75 100 150 200 250 300 350 400 Horse-power Fig. 163. Cost of Simple High-Speed Engines. S0 2 does not attack the metal surface of the engine unless combined with water, in which case sulphurous acid is formed. There is, how- ever, no danger from this cause, since the S0 2 being under greater pressure effectually prevents leakage of water into the S0 2 system. STEAM ENGINES 325 The S0 2 cylinder requires no other lubrication than the S0 2 itself, which is of a greasy nature. / s y 40000 y y 35000 .,"■■ 30000 /< ^ £ 25000 i t K » s '\ k*< 20000 s' J 15000 S » 10000 s / 5000 8500 J* y * 200 400 600 800 1000 1200 1400 1600 1800 2000 2200 2400 2600 ■SDrse-powner Fig. 164. Cost of High-Speed Compound Engines. 100 200 300 400 500 600 700 800 900 1000 1100 1200 Hoige-power Fig. 165. Cost of Low-Speed Engines, Simple and Compound. Properties of S0 2 : Trans. A.S.M.E., 25-181 . Binary-Vapor Engines : Jour. Frank. Inst., June, 1903; Elec. World and Engr., Aug. 10, 1901; U.S. Cons. Reports, No. 1139, Sept. 14, 1901; Engr. U.S., Aug. 1, 1903; Sib. Jour, of Eng., March, 1902. 326 STEAM POWER PLANT ENGINEERING 183. Cost of Engines. — In general the cost of engines per horse power diminishes as the size increases, but is of course governed by the style and workmanship. Average figures may be expressed as follows (Engr. U.S., Nov. 15, 1902, p. 750): Simple high-speed engines Cost in dollars = 300 + 8 X horse power Setting, high-speed engines Cost in dollars = 60 4- 0.75 X horse power Compound high-speed engines Cost in dollars = 1000 + 15 X horse power Simple low-speed engines Cost in dollars = 1000 + 10 X horse power Compound low-speed engines Cost in dollars = 2000 + 13 X horse power Setting, low-speed engines. Cost in dollars = 500 + 1.3 X horse power These equations were deduced from the curves in Figs. 163 to 165, which were plotted from the actual costs of a large number of engines. Rules for testing steam engines. — See Appendix C. CHAPTER X. STEAM TURBINES. 184. Classification. — The following outline gives a classification of a few well-known steam turbines: Steam Turbines. Impulse Type. Reaction Type. Single J Stage. 1 Multi- stage. Multi- Stage. De Laval. Sturtevant. Nozzle • Expansion. Curtis. Kerr. Terry. Rateau. Hamilton-Holzworth. Blade Parsons. • Expansion. Schulz. Allis-Chalmers. In the impulse type the steam is expanded by suitable means before doing useful work; that is, its potential energy is first converted into kinetic energy. In the reaction type the conversion is not complete, the expansion taking place partly before doing work upon the wheel and partly within the blades of the wheel itself. Thus the steam gives up a portion of its energy by direct impulse in impinging against the blades or buckets and the balance by reaction in leaving them. The impulse type may be either single- or multi-stage, depending upon the number of divisions in which expansion takes place, but the reaction type is always multi-stage. The single-stage impulse machine has one row of buckets or vanes mounted on the periphery of a revolving disk and one set of stationary nozzles. The peripheral velocity is very high, ranging in practice from 700 to 1400 feet per second. In the multi-stage impulse machines the expansion is divided between a number of stages, each one exhausting through suitably porportioned nozzles into the next succeeding stage. The steam velocity is thereby very much reduced, and the peripheral velocity may be considerably lower for good efficiency, ranging in practice from 200 to 600 feet per second. In the reaction type the steam flows through a large number of rows of blades alternately fixed and revolving. 327 328 STEAM POWER PLANT ENGINEERING 184a. General Elementary Theory. — A given weight of steam at a given pressure and temperature occupies a certain known volume and contains a known amount of heat energy. If the steam is permitted to expand to a lower pressure without receiving additional heat or giving up heat to surrounding bodies it is capable of doing a certain amount of work which will be the same whether the expansion takes place in the cylinder of a reciprocating piston engine, a rotary piston engine or the nozzles and blades of a steam turbine. Let W = weight of steam, lbs. per sec. E = energy given up by 1 pound of steam, ft.-lbs. P x = initial pressure, lbs. per sq. in. abs. P n = final pressure, lbs. per sq. in. abs. H 1 = initial heat contents per lb., B.T.U. H n = final heat contents per lb., B.T.U. Then the heat available for doing useful work is If the steam expands against a resistance, as, for example, the piston of a reciprocating engine, the energy given up in forcing the piston for- ward may be expressed E x = 778 W (H t - H n ) ft.-lbs. (72) If the steam expands within a perfect nozzle the energy will be given up in imparting velocity to the steam itself, thus: E 2 =W^- ft.-lbs., (72a) 2 2g in which V t = velocity of the jet in feet per second. If the velocity of the jet is retarded to V n feet per second, as by placing a series of vanes in its path, then the energy given up to the vanes (neglecting all losses) is V 2 — V 2 E = W * n - (72b) 2g If the jet is brought to rest by the vanes (neglecting all losses), then V n = and the energy given up is E 3 = W^. (72c) But E X =E 3 . Hence, from which 778W(H l -H n )=W^-, & 9 V x = 223.9 VX - H^* (73) * For most purposes it is sufficiently accurate to make 223.9 = 224. STEAM TURBINES 329 The jet issuing from the nozzle is capable of exerting an impulse equal to F upon any object in its path, thus: WV F = i^- 1 lbs. (74) 9 If A = the area of cross section of the jet in sq. ft. and y = weight of steam, lbs. per cubic foot, then W = yAV 1} or F = *^Il lbs. (74a) 9 The reaction, R, of the jet against the nozzles is equal in value and opposite in direction to the impulse, or B = F = EX± = rlR. (74b) 9 9 The theoretical horse power developed by a jet of steam flowing at the rate of one pound per second may be expressed E v 2 —V 2 H * P - = 550 = 2^X550' (75) in which y ^ = initial velocity of the j et> ft per sec> V n = final velocity of the jet, ft. per sec. Steam consumption per horse power hour: W-jg. (75a) Heat consumption, B.T.U. per horse power, per min. 60 in which ^ = k eat Q £ fa e iiq U i(j at p ressure p n . (75b) Impulse efficiency of the jet = equation (72b) ~ equation (72c). Thermal efficiency V 2 — V 2 ft = v * (76) E t = ^ — ^2- See equation (68). (76a) «l — On Efficiency ratio or " kinetic " efficiency: 2545 F r = w \jj _ H , ■ See equation (71). (76b) Equations (72) to (76b) are general and are applicable to all turbines of whatever make. The more important types of turbines will be discussed separately and an application of above equations will be made in each specific case. 330 STEAM POWER PLANT ENGINEERING STEAM TURBINES 331 185. The De Laval Turbine. — Fig. 166 shows a section through a De Laval steam turbine and gear case and illustrates the principles of the single-stage " velocity " type. The turbine proper, to the right of the figure, consists of a high-carbon steel disk W fitted at the periphery with a single row of drop-forged steel blades and inclosed in a cast-steel casing. The disk is secured to a light flexible shaft and is of such a cross section that the radial and tangential stresses through- out its mass are of constant value. A flexible shaft is employed which allows the wheel to assume its proper center of rotation and thus to operate like a truly balanced rotating body.* The shaft is supported by three bearings, F, P, and 2V. N is self-aligning and carries the greater part of the weight of the disk. P is a flexible bearing, entirely free to oscillate with the shaft, and its only function is to seal the wheel casing against leakage. The power is transmitted through a steel helical pinion K mounted on the extension of the turbine shaft X, to two large gears E, E at a reduction in speed of about 10 to 1. The blades, Fig. 167, are made with a bulb shank and fitted in slots milled Fig. 167. De Laval Blades. in the rim of the wheel. The flanges, at the outer end of the blades, are brought in contact with each other and calked so as to form a continu- ous ring. The inlet and outlet angles of the blades are made alike and are 32 degrees for smaller sizes and 36 degrees for larger sizes. The operation is as follows: Steam enters the steam chest D, Figs. 166 and 168, through the governor (shown in detail in Fig. 169) and is distributed to the various adjustable nozzles, varying in number from 1 to 15 according to the size of turbine. In the earlier types the nozzles were uniformly distributed around the circumference, but in the later types are arranged in groups. As illustrated in Fig. 168, * The shaft diameter for a 100-H.P. turbine is but 1 inch and for a 300-H.P. approximately 1{% inches. 332 STEAM POWER PLANT ENGINEERING the nozzles are placed at an angle of 20 degrees with the plane of the disk. The steam is expanded adiabatically in the nozzles to the existing back pressure before it impinges at high velocity against the De Laval Nozzle. blades. After giving up its energy the steam passes into chamber G, Fig. 166, and out through the exhaust opening. Fig. 169 gives the details of the governor and vacuum valves. Two weights B are Fig. 169. De Laval Governor. pivoted on knife edges A with hardened pins C bearing on the spring D. E is the governor body, fitted in the end of the gear wheel shaft K, and has seats milled for the knife edges A. The spring seat D is STEAM TURBINES 333 held against pins A by spiral concentric springs, the tension on which is adjusted by a milled nut /. When the speed exceeds the normal, centrifugal force causes the weights to fly outward and overcome the resistance of the springs. This pushes pin G against bell crank L, which in turn closes the double-seated valve, thus throttling the supply of steam. To prevent racing in case the load is suddenly removed the vacuum valve T is added to the governor mechanism. Its operation is as follows: The governor pin G actuates the plunger H under normal conditions without moving the plunger relative to the bell crank. In case the load is suddenly removed, centrifugal force pushes pin G against bell crank L until it reaches its extreme position and the valve is nearly closed and little steam enters the turbine. If this does not check the speed, plunger G overcomes the resistance of spring M, and H moves relative to L, and its adjustable projection presses against valve stem T and allows air to rush into the turbine chamber through passage P. The power of the turbine depends upon the number of nozzles in action, and these can be opened or closed by a hand wheel on each. Each nozzle performs its function as perfectly when operating alone as when operating in conjunction with others. De Laval turbines are made in sizes ranging from 1J to 300 horse power, condensing and non-condensing, and are designed to regulate within an extreme variation of 2 per cent from no load to full load. The speeds vary from 10,600 r.p.m. for the largest size to 30,000 r.p.m. for the smallest, the gearing reducing these to 900 and 3000 r.p.m., respectively, at the shaft. The diameter of the wheel varies from 4 inches in the smallest tur- bine to 30 inches in the largest, thus giving peripheral velocities of from 520 to 1310 feet per second. De Laval Turbine : Prac. Engr., Jan. 1, 1910; Trans. A.S.M.E., 25-1056; Elec. World, July 29, 1905, p. 194, Oct. 26, 1901, p. 693; Eng. Rec, Oct. 19, 1901, p. 371; West. Elecn., June 4, 1904, p. 463; Machinery, Oct., 1904, p. 6, Nov., 1904, p. 123; Electrician, March 4, 1904; Power, Oct., 1905, p. 593. See Table 48 for results of tests of turbines of this type. 186. Elementary Theory. — De Laval Turbine. — The maximum theo- retical power developed by a jet of steam flowing through a nozzle is dependent only upon the weight of steam flowing per unit of time and the initial velocity. Therefore the higher the initial velocity for a given rate of flow the greater will be the power developed and the higher the efficiency. * In Europe De Laval turbines are made as large as 3750 H.P. (See Power and Engr., May 10, 1910, p. 708.) 334 STEAM POWER PLANT ENGINEERING The maximum weight of steam discharged through a nozzle of any shape and for a given initial pressure is determined by the area of the narrowest cross section or throat. To obtain the maximum velocity at the exit or mouth, for a given rate of flow, the nozzle should be proportioned so that expansion to the external pressure into which the nozzle delivers shall take place within the nozzle itself. If expansion in the nozzle is incomplete, sound waves will be produced and there will be irregular action and loss of energy. On the other hand, if expansion in the nozzle is carried below that of the external pressure at the mouth, sound waves will be produced with subsequent loss of energy even greater than in the former case. Experimental and mathematical investigations indicate that the pressure at the narrowest section of an orifice or the throat of a nozzle through which steam is flowing falls to approximately 0.58 of the initial absolute pressure (with resultant velocity of about 1400 to 1500 feet per second) and any farther fall in pressure must take place beyond the narrowest section. Thus for back pressures greater than 0.58 of the initial (conveniently takes as f ), maximum exit velocity may be ob- ^00?0MMmmm, Fig. 170. Theoretically Proportioned Expanding Nozzle. tained from orifices or nozzles of uniform cross section or with sides convergent. For back pressure less than 0.58 of the initial the nozzle must first converge from inlet to throat and then diverge from throat to mouth in order to obtain maximum velocity. Without the divergent portion of the nozzle the jet will begin to spread after passing the throat, and its energy will be given up in directions other than that of the original jet. Fig. 170 shows a section through a theoretically proportioned expand- ing nozzle. The cross section of the tube at any point n may be cal- culated by means of equation WS n A n V n (76c) STEAM TURBINES 335 in which A n = area in square feet. W = maximum weight of steam discharged, lbs. per sec. S n = specific volume of the steam at pressure P n . For saturated steam S n = x n u n> in which x n = quality of steam at pressure P n after adiabatic expansion from pressure P v u n = specific volume of saturated steam at pressure P n . For superheated steam, see equation at bottom of page 131. V n = velocity of the jet, feet per second. V n may be determined from equation (73): V n = 223.9 VH l - H n . By substituting H n = heat contents corresponding to pressure P n = 0.58 P t in equation (73) and (76c) the area at the throat may be readily determined. The cross-sectional area for other points in the tube may be determined in a similar manner by assigning values of H n corresponding to the various pressures. In case of a perfect nozzle H x — H n represents the heat given up toward producing velocity by adiabatic expansion from pressure P x to P n . In the actual nozzle the frictional resistance of the tube serves to increase its dryness fraction, but in doing so it decreases the amount of energy the steam is capable of giving up towards increasing its own velocity. If y one-hundredths of the heat H t — H n is utilized in over- coming frictional resistance, then the resulting velocity will be V = 223.9 V(l -i^iH.-Hn). (76d) The quality of the steam after expanding to P n against the resistance will be higher by an amount t ,-, y (H ± — H n ) f l n = increase in quality = > (7oe) in which r n = heat of vaporization at pressure P n . The curves in Fig. 171, calculated by means of equations (76b) and (73), show the relationship between velocity, quality, pressure and kinetic energy for all points in a theoretically perfect nozzle expanding one pound of dry steam per second from an initial absolute pressure of 190 pounds to a condenser pressure of one pound. The curves in Fig. 172 are based upon the experiments of Gutermuth 336 STEAM POWER PLANT ENGINEERING {Zeit. d. Ner. Ingr., Jan. 16, 1904) and show the effect of a few shapes of nozzles and orifices on the actual weight of steam discharged for various rates of initial and final pressures, the smallest section of the tube remaining constant. The nozzles of most commercial types of steam turbines are made with straight sides as in Fig. 168, so that only the area at the mouth need 5000 4500 4000 3500 2000 1500 1000 500 2500 - THEORETICAL DESIGN OF A DIVERGENT NOZZLE 190 180 \ 170 160 \ \ S * \ 140 \ 130 \ 120 \ 4 K^S 110 \ 4* \ 90 Q J *L, > 70 60 50 40 30 J * A ^> I s 10 5S ^5£, 40,000 80,000 120,000 160,000 200,000 240,000 280,000 kinetic energy of the jet, in foot pounds Fig. 171. 100 90 80 70 be determined in addition to that at the throat in order to lay out the shape of the tube. Equations (73) and (76b) are general and are applicable to steam of any quality, wet, dry, or superheated. For steam initially dry and saturated Napier's rule offers a simple means of determining the area at the throat, thus: W = A . 70 for P n = or < - P x (76e) in which STEAM TURBINES 337 W = 0.029 A VP n (P l -P n ) for P n > f P lt W = maximum weight of steam discharged, lbs. per sec. A = area at the throat, sq. in. P l = absolute initial pressure, lbs. per sq. in. P n = absolute back pressure, lbs. per sq. in. .06 05 .04 03 02 .01 4 .00 V \ \ VJ >3 \ "v .05 \ \ \ .04 ._,.. r i -> P 2 \ ■ " .03 , i H 3 \ P i- ■ " p r p- 1 ^ P \ * , .02 zMtW, 1 2 p— i 1 P r~ '■■. .01 P = 132 Lb. Per Sq. In. Absolute l 4 A rea )fOi ifioc 0.0 555 S q.Ir . .1 .2 .3 A .5 .7 .8 .9 1.0 .1 .2 .3 A .5 .6 .7 .8 .9 1.0 Fig. 172. Flow of Steam through Nozzles. Moyer ("The Steam Turbine," 1st Edition, p. 40) states that the ratio of the area of a correctly proportioned nozzle at the throat A to the area at any point A n is very nearly proportional to the ratio of the pres- sure at point A n to the initial pressure, or A P, Pn (76f) The entrance to the tube is rounded by any convenient curve. The length of the tube may be roughly approximated by the following formula: L = Vl5 A , (76g) in which L = length between the throat and mouth, in inches. A = area at the throat, sq. in. Practice shows that the cross section of a nozzle, whether circular, elliptical, square or rectangular (the latter with rounded comers), has very little influence on the efficiency provided the inner surfaces are smooth and the ratio of the area at the throat to that of the mouth is 338 STEAM POWER PLANT ENGINEERING cDrrectly proportioned. The velocity efficiency of a properly propor- tioned nozzle with straight sides is about 95 to 97 per cent, corresponding to an energy efficiency of 92 to 94 per cent, so that it is not considered worth while to attempt to follow the more difficult exact curves. Example : — Find the smallest cross section of a frictionless conically divergent nozzle for expanding one pound of steam per second from an absolute initial pressure of 190 pounds to an absolute back pressure of 2 pounds and find six intermediate cross sections where the pressures will be 70, 30, 14.7, 8, 4 and 2 lbs. respectively. Compare the velocity and energy of the jet issuing from this nozzle with those of an actual nozzle in which 10 per cent of the heat energy is lost in friction. From steam and entropy tables we find the values of H, x, u, for absolute pressures corresponding to 190, 0.58 X 190 = 110, 70, 30, etc., lbs. per square inch as follows (theoretical nozzle): H. X. u. 5 = xu. P x = 190 1197.3 1.00 2.405 2.394 P 2 = 110* 1152.6 0.960 4.047 3.878 P,= 70 1117.9 0.932 6.199 5.775 P 4 = 30 1057.2 0.887 13.75 12.27 P 5 = 14.7 1011.3 0.857 26.78 22.95 P." 8 947.8 0.834 47.26 39.29 P 7 = 4 935.6 0.810 90.4 73.2 P«= 2 899.3 0.788 173.1 137.0 * P 2 = 0.58 Pi ( = pressure at throat). If entropy tables or charts are not available, values H l to H 8 and Xj to x s may be determined as outlined in equations (66b) to (67g). The different quantities for the theoretical nozzle will be calculated for the exit pressure P n = P s = 2 lbs. per sq. in absolute. V 8 = 223.9 VH, - H 8 = 223.9 V1197.3 - 899.3 = 3865 feet per second. E 8 = 778 (H t - H 8 ) = 778 (1197.3 - 899.3) = 232,000 foot-pounds. WS V 1 X 137 A* = 3865 = .0353 square foot. STEAM TURBINES 339 dg = y/pt^A = 13.58 VI = 13.56 V3353 = 2.54 inches. WV a F» = 9 3865 32.2 = 120 pounds. THEORETICAL NOZZLE. Quantity < V Ft. per Sec. E Ft.-Lbs. A Sq. Ft. d Inches. F Pounds. (73) (72) (76c) (74) Pressures - 110 70 30 14.7 8 4 2 1,496 1,995 2,650 3,053 3,339 3,624 3,865 34,767 61,853 107,485 144,742 173,207 203,968 232,000 .00259 .00269 .00461 .00745 .0119 .0202 .0353 0.693 0.702 0.919 1.1 1.46 1.92 2.54 46.4 61.98 82.3 94.8 103.7 112.5 120.0 In the actual nozzle these values will be modified because of the frictional losses. Thus for P n = 2 lbs., V s = 223.9 \/ (l - y) (H 1 - H 8 ) = 223.9 V(l - 0.1) (1197.3 - 899.3) = 3667 ft. per sec. E 8 = 778 (1 - 0.1) (1197.3 - 899.3) = 208,800 ft.-lbs. y (# i - h 8 ) ^8 3*H ~l~lfi #8 1 Qs = 0.788 + 0.1 (1197.3 - 899.3) 1021 = 0.788 + 0.029 = 0.817. A» = WxJuc _ 0.817 X 173.1 3667 = 0.0386 sq. ft., 340 from which STEAM POWER PLANT ENGINEERING d 8 = 2.66 inches. WV 8 3668 8 a 32.2 114 pounds. These various factors for all given pressures have been calculated in a similar manner and are as follows: ACTUAL NOZZLE. Quantities < V Ft. per Sec. E Ft. -Lbs. x'. A Sq. Ft. d Inches. F Ft. -Lbs. 110 1,420 31,317 .9658 .00275 0.711 44.1 70 1,893 55,632 .9414 .00286 0.723 58.8 30 2,515 98,257 .9026 .00493 0.951 78.12 Pressures ■ 14.7 2,894 130,050 .876 .0080 1.2 98.8 8 3,168 155,858 .856 .0127 1.53 98.4 4 3,438 183,581 .836 .0220 2.01 106.8 2 3,667 208,800 .817 .0386 2.66 114.0 Many of these values may be determined directly from the Mollier or total heat-entropy diagram as described in Appendix H; in fact, the Mollier diagram has to all intents and purposes supplanted the steam tables in this connection. For superheated steam the diagram is extremely useful in avoiding laborious calculations. Fig. 172a. Velocity Diagram. Ideal Single-stage Impulse Turbine. Fig. 172a gives a diagrammatic arrangement of the blades in a De Laval turbine. The nozzle directs the steam against the blades with absolute velocity V x and at an angle a with the plane of the wheel XX. Since the wheel is moving at a velocity of u feet per second, the velocity STEAM TURBINES 341 v x of the steam relative to the wheel is the resultant of V t and u. The angle ( /? 1 between v x and XX will be the proper blade angle at entrance. If the blade curve makes this angle with the direction of motion of the wheel no shock will be experienced when the steam enters the blades. For convenience in construction the exit angle /? 2 is made the same as the entrance angle j3 v Neglecting frictional losses in the blade channels the relative exit velocity will be v 2 = v l} and the absolute velocity V 2 is the resultant of v 2 and u. The impulse exerted by the jet in striking W the vanes is — v v and its component in the direction of motion is W W — v t cos /?! = — (V l cos a — u). As the jet leaves the vanes the im- W W pulse is v 2 cos /? 2 .= (V 2 cos y + u). The total pressure acting on the vanes, or the actual driving impulse, is W 9 V, cos a (— V 2 cos y + u) W (V l cos a + V 2 cos y). (77) Equation (77) may also be expressed W p = _ . 2 (7, cos a -u). 9 The resultant axial force or end thrust is W F = — {V 1 sin a — V 2 sin y). (77a) (77b) Evidently if a = y and V\ = V 2 there will be no end thrust, since V l sin a — V 2 sin y will be zero. The work done is W Pu = — u (V\ cos a + F 2 cos y), (77c) or, using equation (77a) in place of (77), W Pu = — ' 2 u (V t cos a — u) W = — • 2 (uV, cos a— u 2 ). g I J By making the first derivative equal to zero d ( W ) j-) — 2 {uV 1 cos a — u 2 ) \ = V x cos a — 2 u = 0, or ^ y t cos a. (77d) 342 STEAM POWER PLANT ENGINEERING That is, for any nozzle angle a the work done, Pu, has its greatest value when u = \ V x cos a or y = 90°, whence Pu= W ^-cos 2 a. (77e) The work for any initial velocity V t becomes a maximum when a = and tt = J V v This condition can only occur for a complete reversal of jet and zero final velocity. Substitute a = and u = \ V x in equation (77d). TFT 2 Pu = — — — » which is necessarily the same as equation (72c). In the actual turbine the various velocities will be less than those as obtained on account of the frictional resistance in the blades, and the velocity diagram should be modified accordingly. Example. Lay out the blades (theoretical and actual) for the nozzle in the preceding example, assuming that the jet impinges against the wheel at an angle of 20 degrees and that the peripheral velocity is 1250 feet per second. Theoretical Case. Lay off V x = 3865 feet per second in direction and amount as shown in Fig. 172a and combine it with u = 1250 feet per second; this gives v v the relative entrance velocity as 2725 feet per second and /?, the entrance angle as 29 degrees. Lay off v 2 = v t at an angle /? 2 = & and combine with u; this gives V 2 , the absolute exit velocity, as 1740 feet per second. The theoretical energy available for doing work is W E = f g (V 1 >-V 2 >) = -7TT-A (3865 2 - 1740 2 ) = 185,000 ft.-lbs. 64.4 The difference between 232,000 and 185,000 = 47,000 ft.-lbs. is evi- dently the kinetic energy lost in the exhaust due to the exit velocity. The pressure exerted by the steam on the buckets is W P = — (V t cos a + V 2 cos 7) = ^ (3865 X 0.9397 + 1740 X 0.65166) = 148 pounds. The theoretical impulse efficiency is V t 2 - V 2 2 _ 3865 2 - 1740 2 _ V7 - ~ 3865 2 " U '' y/ STEAM TURBINES 343 The theoretical horse power per pound of steam flowing per second is 185,000 H.P. 550 = 336. Theoretical steam consumption per H.P.-hr. is 3600 336 10.7 pounds. Actual Case. Proceed as in the theoretical case, using the actual absolute velocity V x = 3865 Vl - y = 3865 Vl -0.10 = 3667 feet per second in place of the theoretical value V t = 3870. Lay off V t = 3667 at an angle of 20° as before and combine with u = 1250, Fig. 172b. U = 1250 U = 1260 Fig. 172b. Velocity Diagram as Modified by Friction Losses. The resultant v l = 2530 is the velocity of the jet relative to the wheel, and the entrance angle /? is found to be 29.7 degrees. The relative exit velocity v 2 will be less than v x because of the blade friction. Assume the loss of energy from this cause to be 14 per cent; then, since the velocity varies as the square root of the energy, u, = v. Vl — d> = 2530 VI -0.14 = 2346 ft. per second. The resulting absolute velocity V 2 is found from the diagram to be V 2 = 1405 ft. per second. Since the loss of energy in the nozzle is V 2_ {X _ y)V 2 and that in the blade 2g (1-0) TV 2.9 344 STEAM POWER PLANT ENGINEERING the remaining energy, deducting both losses in the nozzle and the blades, is = j^j (3865 2 - 0.1 X 3865 2 - 0.14 X 2530 2 - 1405 2 ) = 164,200. The losses due to windage, leakage past the buckets and mechanical friction must be deducted from these figures to give the actual energy available for doing useful work. Assuming a loss of 15% due to this cause, the work delivered is 0.85 X 164,200 '= 139,570 ft.-lbs. The efficiency in the ideal case was found to be 0.797 and the available energy 185,000 ft.-lbs. The efficiency, deducting the loss due to friction, etc., is 139,570 185,000 The horse power delivered is 139,570 0.797 = 0.60. = 254. 550 Steam consumption per horse-power hour is 3600 tin A -T-— - = 14.2 pounds. 254 r The heat consumption, B.T.U. per H.P., per minute is 16.4 (1197.3 - 94) 60 = 316 ' Assuming the revolutions per minute to be 10,000, the mean diameter of the wheel to give a peripheral velocity of 1250 ft. per second is 1250 X 60 _ QO , + oc „. , 1000 X 3.14 = 2 - 39 ft " ° r 28 ' 6 mcheS ' The determination of the height and width of vanes, clearance between nozzles and blades, etc., are beyond the scope of this work and the reader is referred to the accompanying bibliography. Blade Design for De Laval Turbines: Moyer, Steam Turbine, Chap. IV; Power, Mar. 17, 1908, p. 391. Flow of Steam through Nozzles: Jour. A.S.M.E. Mid. Nov., 1909, April, 1910, p. 537; Engineering, Feb. 2, 1906; Engr., Lond., Dec. 22, 1905; Eng. Rec, Oct. 26, 1901; Power, May, 1905; Eng. News, Sept. 19, 1905, p. 204; French, Steam Tur- bines, Chap. XI; Pro. Inst. Civ. Engrs., Feb. 2, 1906. Design of Turbine Disks: Engr., Lond., Jan. 8, 1904, p. 34, May 13, 1904, p. 481; Jude, Theory of the Steam Turbine, Chap. XIII; Thomas, Steam Turbines, Chap. VI. STEAM TURBINES 345 Steam Turbine Efficiency: Power, Feb., 1906, p. 83; Jude, Theory of the Steam Turbine, Chap. VIII. Critical Velocity of Shafting: Jour. A.S.M.E., June, 1910, p. 1060; Power, Sept., 1903, p. 484; Stodola, Steam Turbines, p. 177; Jude, Theory of the Steam Turbine, Chap. XVI; French, Steam Turbines, Chap. XV. Tests of De Laval Turbines: Eng. Rec., Aug. 2, 1902, p. 100; Am. Elecn., Aug., 1905, p. 445; Engr. U. S., Aug. 1, 1905, p. 526; Eng. and Min. Jour., Nov. 3, 1904, p. 706; Machinery, Aug., 1904, p. 560; Eng. News, June 19, 1905, p. 62. 187. Terry Turbine. — Fig. 173 shows a section through a Terry turbine, illustrating an application of the impulse type with two or more velocity stages. The rotor, a single wheel consisting of two steel Fig. 173. Section through Terry Steam Turbine. disks held together by bolts over a steel center, is fitted at its periphery with pressed steel buckets of semicircular cross section. The inner surface of the casing is fitted with a series of gun-metal reversing bucket arranged in groups, each group being supplied with a separate nozzle. The steam issuing from nozzle N, Fig. 174, strikes one of the buckets, B, on the wheel and, since the velocity of the buckets is comparatively low, is reversed in direction and directed into the first one of the revers- ing chambers. The chamber redirects the jet against the wheel, from which it is again deflected; this is repeated four or more times until the available energy has been absorbed by the rotor. Terry turbines are made in a number of sizes varying from 2 to 800 horse power, and operate 346 STEAM POWER PLANT ENGINEERING at speeds varying from 210 feet per second in the smaller machine to 260 feet per second in the larger. These low speed limits compared with the speed of single-stage De Laval turbines are made possible by the application of the velocity stage principle in the use of the reversing buckets. The rotor of the smaller machine is 12 inches in diameter and runs at 4000 r.p.m., and that of the larger, 48 inches, running at 1250 Fig. 174. Arrangement of Buckets and Reversing Chambers in a Terry Steam Turbine. r.p.m. Since the flow of steam into and from the buckets is in the plane of the wheel there is no end thrust. For a description of the Bliss, Dake, Sturtevant and Wilkinson steam turbines with results of tests see " Small Steam Turbines," by G. A. Orrok, Jour. A.S.M.E., May, 1909, and contributed discussion, Sept., 1909. See also, " The Development of the Small Steam Turbine," Eng. Mag., Dec, 1908, and Jan., 1909. 188. Kerr Turbine. — Fig. 175 shows a longitudinal section through, and Fig. 176 a sectional elevation of, a Kerr steam turbine. This turbine is of the impulse type and built on the principle of the Pelton water wheel, which it resembles in many respects. The rotor consists of a series of steel disks R, R, Fig. 175, mounted on a steel shaft. A series of drop-forged mild steel buckets of the double-cup type are secured to the periphery and riveted in dovetail slots. The stator is made up of a number of cast-iron diaphragms S, S, with cir- cular rims, which are tongued and grooved and when drawn together form a continuous cylinder. Square cold-rolled steel nozzle bodies N, N are expanded and beaded in the diaphragm near the rim and the nozzles screwed into them. The bearings B, B are of the oil-ring type; no thrust blocks are necessary, as each element is practically balanced. The operation is as follows: Steam enters the turbine at inlet A and passes through balanced throttling valve V (controlled by governor G) to the circular cored space H, H extending around the STEAM TURBINES 347 348 STEAM POWER PLANT ENGINEERING entire casing. Space H, H acts as an equalizer and insures uniform admission to the first row of nozzles, where steam is partially expanded and the kinetic energy, imparted to the rotor through the medium of the buckets. The steam leaves the buckets at practically zero velocity Fig. 176. Kerr Steam Turbine: Sectional End Elevation. and is again expanded through the second set of nozzles. This process is repeated in each stage and the exhaust steam leaves the turbine at 0. Fig. 177 shows a diagrammatic arrangement of the governor. The governor weight is turned from solid steel and split into two piece, 7 Ctr. Line of Shaft -? Section- A-A Fig. 177. Kerr Steam Turbine Governor. of semi-cylindrical form with the center of gravity near the center of the shaft. The weights are supported at three points. The hardened steel knife edge at B is of sufficient length for the stresses involved. The curve of rolling contact C is such that the bearing between the weigh STEAM TURBINES 349 TURBO-M-TERNATQR Fig. 178. Four-stage Vertical Curtis Turbo-Generator. Base Condenser T ype. 350 STEAM POWER PLANT ENGINEERING and the cam collar is always on the line of centers. The outward move- ment of the weights compresses the spring and operates, through lever connections, the balanced piston valve controlling the flow of steam. The movement of the center of gravity is indicated. The Kerr turbine is very simple in design, compact, noiseless, and low in cost of repairs. Its performance compares favorably with all other types of turbines of similar size and capacity. An 18-inch Kerr turbine direct connected to a multi-stage Worthing- ton centrifugal pump at the Armour Institute of Technology gives a steam consumption when running non-condensing comparable with that of high-grade non-condensing engines. Rateau Turbine : Trans. A.S.M.E., 25-782; Eng. Mag. T Oct., 1903, p. 49; St. Ry. Jour., April 18, 1903. Zolly Steam Turbine : Elec. Rev., Sept. 2, 1904. 189. The Curtis Steam Turbine. — Figs. 178 to 183 show the general arrangement and a few details of the Curtis steam turbine, which is of the compound or multi-stage velocity type. The total expansion is carried out in one or more compartments or stages, each stage compris- ing a set of expanding nozzles and a wheel carrying two or more rows of buckets. A high initial velocity is given to the jet in each stage by expansion in the nozzles as in the De Laval, and the energy absorbed by successive action upon the series of moving and stationary vanes arranged somewhat as in the Parsons turbines, paragraph 192. In the latter, however, the difference in pressure between the two sides of each vane induces flow by continuous expansion, while in the former the moving vanes- in any one stage simply absorb the kinetic energy already created by expansion in the nozzle. The action is as follows : Steam enters stage (1), Fig. 180, through the first set of nozzles, and is partially expanded. With the resulting initial velocity it impinges against the first row of moving blades and gives up part of its energy, and is then deflected through the adjoining stationary blades to the next set of moving vanes, where its velocity is still further reduced, and so on until it has been brought practically to rest. From this stage the steam flows at reduced pressure through nozzles of stage (2), which are suffi- cient in number and in size to afford the greater area required by the increased volume. In expanding in these nozzles it acquires new velocity and gives up energy to the moving blades as before. This process is repeated through two to five stages, depending upon the size of turbine. Fig. 178 shows a partial section of a four-stage 5000-kilo- watt machine. R, R are sections through the revolving wheels, which in this particular turbine are nine feet in diameter and keyed to the STEAM TURBINES 351 352 STEAM POWER PLANT ENGINEERING vertical shaft S. On the periphery of each wheel are bolted two rows of blades or vanes, with a stationary or intermediate row attached to the casing between them. The buckets are made of rolled nickel bronze, hammered to shape and finish. The roots are dovetailed into the holders and the tips are tenoned and riveted into a shroud ring, thus insuring positive spacing and a rigid construction. Between each pair of wheels is a stationary steam-tight diaphragm P, which contains the nozzles through which the steam is expanded from the preceding stage. It will be noticed that the buckets and nozzles increase rapidly in size in suc- ceeding stages as the pressure falls and the volume of steam increases. MOVING BLADE9 STATIONARY BLADES MOVING BLADES STAGE No. 2 S NOZZLE DIAPHRAGM ffiHIffl !<<<<<<<<< ) \i MOVING BLADES ^^m STATIONARY BLADES n MOVING BLADES EXHAUST Fig. 180. The parts are so proportioned that the steam gives up approximately -of its energy in each stage, n representing the number of stages. n The number of stages and the number of vanes in a stage are governed by the degree of expansion, the peripheral velocity which is desirable or practicable, and by various conditions of mechanical expediency. The number of admission valves vary in number and in location with the size of turbine. The automatic stage valve G connects the first stage directly to a set of auxiliary second-stage nozzles. Thus the overload capacity is increased by widening the steam belt and not by admitting high-pressure steam into an intermediate stage as was for- merly the practice with Curtis turbines. This method of overload con- trol results in higher efficiency than with the older system. STEAM TURBINES 353 Curtis turbines appear to have a wider range of economical application than any other type, commercial sizes ranging from a small horizontal unit of 7 kilowatts rated output to vertical units of 20,000 kilowatts capacity on the continuous 24-hour basis. The smaller machines, 1000 kilowatts and under, are usually of the horizontal type, and the larger units, 3500 kilowatts and larger, are of the vertical type. Be- tween 500 and 3500 kilowatts they are made both vertical and horizontal. All Curtis turbines are governed by " cutting-out nozzles"; that is, full initial pressure is maintained in all the nozzles that are open and Fig. 181. Section through Curtis Governor. the capacity of the machine is controlled by varying the number in operation. Units under 1500 kilowatts are ordinarily controlled by a mechanical valve gear and the larger units by an indirect or relay system. In the older types this relay system was electrically operated; in the modern machines the valves are hydraulically controlled. Fig. 181 shows a section through a typical Curtis governor. Speed regulation is accomplished by the balance maintained between the centrifugal force of moving weights A A and the static force exerted by spring D. The governor is provided with an auxiliary spring F, for varying its speed when synchronizing, the tension in which is varied by 354 STEAM POWER PLANT ENGINEERING a small pilot motor controlled from the switchboard. The movement of the governor weights is transmitted through rod C to arm H and by means of the latter to the controlling mechanism of the valve gear. Fig. 182 gives an assembly view of the mechanical valve gear as Fig. 182. Assembly of Mechanical Valve Gears for 300-Kw. Curtis Steam Turbine. applied to a 300-kilowatt unit. The valve stems extend upward through ordinary stuffing boxes and are attached to notched crossheads 8, 8. Each crosshead is actuated by a pair of reciprocating pawls or dogs, 6, 6, the lower one of which closes the valve and the upper one opens it. The several pairs of pawls are hung on a common shaft which receives a rocking motion from a crank driven by the turbine shaft. The cross- STEAM TURBINES 355 heads have notches milled in the side in which the pawls engage to open or close the valve, the engagement being determined by shield plates 2, the positions of which are controlled by the governor through the medium of suitable levers. Shield plates 6 are set one a little ahead of the other to obtain successive opening or closing of the v^es. The pawls are held in position when not in contact with the shield plates by springs W. Fig. 183 gives a diagrammatic arrangement of the hydraulically controlled valve gear mechanism. The motion of governor g is trans- mitted through lever i to lever a of the pilot valve /. Pilot valve j controls the supply of oil (under pressure) in cylinder k the piston of which actuates rods I, I. The movement of rod I is transmitted through •Q p Turbine Fig. 183. Diagrammatic Arrangement of Hydraulically Operated Valve Gear, Curtis Turbine. rack m to a small pinion. This pinion is mounted on the end of a shaft fitted with a number of cams, one a little ahead of the other, each cam controlling the opening and closing of a steam valve through the medium of rocker arm/. As the load on the turbine increases the governor slows down and causes the cam shaft to rotate in a reverse direction indicated by the arrow points in Fig. 183. This causes a proportionate number of valves to be lifted and held open, the number increasing as the load increases, until all are open. Should the load continue to increase, as in the case of overload, the secondary valve opens as previously described, connecting the first stage with a set of auxiliary second stage nozzles. Only the nozzles in the first stage are controlled by the governor. Should the turbine run above normal speed the emergency stop valve automatically closes the admission of steam to the nozzles. This device 356 STEAM POWER PLANT ENGINEERING consists of a steel ring placed around the shaft between the turbine and the generator. This ring is eccentrically mounted and the unbalanced centrifugal force is balanced by a helical spring. When the predeter- mined speed is reached the centrifugal force overcomes the spring ten- sion and the ring moves in a still more eccentric position. In this posi- tion the ring strikes a bell crank lever which trips the throttle valve 1 ■ 5? 2 M 3 2* *5 12345G78 Steam Belt Area Fig. 183a. Steam Belt Area in Five-Stage Curtis Turbine. and permits it to close by its own weight and the unbalanced pressure on the valve stem. In the Curtis turbine the area of the steam admission is limited to a small portion of the circumference in the first stages and does not OIL DRAIN OIL SUPPLY Fig. 183b. Step Bearing for Curtis Turbine. extend around the entire circumference until the last stage is reached. See Fig. 183a. The step bearing of a vertical machine is illustrated in Fig. 183b. The weight of the rotor is supported by oil under pressure forced between the bearing blocks M and P, thus permitting the shaft S to revolve on a film of oil. The smaller disk M is attached by dowels F to the main shaft. Carbon packing rings 0, are used above the bearing to prevent leakage, and adjustment is provided in the lower bearing block STEAM TURBINES 357 RELIEF VALVE, by means of set screws. The oil pressure varies from 150 to 750 pounds per square inch according to the size of machine, the higher pressures being used in the larger machines. Fig. 183c gives a diagrammatic outline of the oiling system. A tank, of sufficient capacity to contain all the oil and fitted with suitable straining devices and a cooling coil, is located at a level low enough to receive oil by gravity from all points lubricated. A pump draws oil from this tank and delivers it at a pressure about 25 per cent higher than that required to sustain the weight of the turbine in the step bearing. A spiral duct baffle connects the source of pressure to the step bearing and serves to regulate the oil supply to the lower end of the shaft. This source of pressure is also connected through a reducing valve to the upper oiling sys- tem of the machine, in which a pressure of about 60 lbs. to the square inch is maintained. This system, which includes a storage tank partly filled with com- pressed air, operates the hydraulic gov- ernor mechanism and supplies oil to the upper bearings. Delivery of oil to these bearings is regulated by adjustable baffles designed to offer resistance to the oil flow without forcing the oil to pass through any very small opening which might easily become clogged. A relief valve is provided to prevent the pressure in the upper part of the oiling system from rising above a desirable limit. Drain pipes from the upper bear- ings and from the hydraulic cylinder and relief valve all discharge into a common chamber, in which the streams are visible, so that the oil distribution can always be easily observed. At some point in the high- pressure system adjacent to the pump it is desirable to install a device to equalize the delivery of oil from the pump, as is done by the air cham- ber commonly used with pumps designed for low pressure. A small spring accumulator is furnished for this purpose, except in cases where weighted storage accumulators are used. In large stations where several machines are installed, a storage accumulator is desirable and can advantageously be so arranged that it will normally remain full, but will discharge if pressure fails, and in doing so will start auxiliary pumping apparatus. TO ACCUMULATOR Fig. 183c. Arrangement of Oiling System, for Curtis Turbine. 358 STEAM POWER PLANT ENGINEERING DIRECT CURRENT. Kw. R.p.m. Kw. R.p.m. 15 25 75 4,000 3,600 2,400 150 300 500 2,000 1,800 1,500 ALTERNATING CURRENT. 300 500 1,000 1,500 1,800 1,800 1,200 900 2,000 3,000 to 20,000 900 I 600-750 For the description of a typical steam turbine station equipped with Curtis turbines see Chapter XX. General Description of Curtis Turbines: Power, March, 1909; Engr. U. S., Jan. 1, 1908, p. 115; Power & Engr., Feb. 25, 1908, p. 284, Feb. 25, 1908, March 3, 1908; Elec. Wld., June 17, 1905, p. 1136. Guide Bearings, Oil Distribution and Carbon Packing: Power & Engr., April 14, 1908. Mechanical Valve Gear: Power & Engr., March 10, 1908, p. 356. Hydraulic Valve Gear: Power, March, 1909, p. 189. 190. Elementary Theory, Curtis Turbine. — Fig. 184 gives a dia- grammatic arrangement of the blades and nozzles in the first stage of a two-stage Curtis turbine, each stage consisting of one set of nozzles and two moving and one stationary sets of blades. The object of employing a number of stages is to permit of a low peripheral velocity without reducing the efficiency. Since the velocity of steam varies as the square root of the kinetic energy, the theoretical stage velocity may be determined by dividing the maximum initial velocity by the square root of the number of stages, assuming that the entire velocity of the jet is abstracted in each stage. Thus if the turbine were constructed with four stages the theoretical stage velocity would be reduced from say 3600 feet per second to 1800 feet per second. In general, in order to reduce the stage velocity to V\ feet per second, the number of stages n may be determined from the equation V 2 n = ~> (77) in which V = maximum initial velocity. Referring to Fig. 184: the steam is expanded in the first stage from pressure P l to P 2 and issues from the first set of nozzles with absolute velocity V lf striking the first set of moving blades at an angle a with the line of motion of the wheel. The resultant v x of V x and the STEAM TURBINES 359 peripheral velocity u, is the velocity of the steam relative to the vanes ; and the angle /? which the line v l makes with the line of motion of the wheel is the proper entrance angle of the blades for the first set. Neglecting friction the exit angle y will be the same as the entrance angle /?. The resultant of v 2 the exit velocity relative to the blade, and u, the peripheral velocity, is V 2 , the absolute exit velocity. Nozzles Fig. 184. Velocity Diagram, Curtis Turbine. Since the second set of blades is fixed and serves as a means of changing the direction of flow, the absolute velocity entering them is V 2 . The angle d formed by V 2 and the center line of the stationary blades is the proper entrance angle. Neglecting friction the absolute exit velocity will be 7 3 = V 2 , and the exit angle will be e = d. The steam flowing from the stationary blades strikes the second set of moving blades at an angle £ = d with absolute velocity V 3 . Combining V 3 with the 360 STEAM POWER PLANT ENGINEERING peripheral velocity u we get v 3 , the velocity of the steam relative to the second set of moving blades. The angle 6, formed by v 3 and the line of motion of the wheel, is the proper entrance angle for the second set of moving blades. The resultant of v 4 (■= v 3 ) and u is V 4 , the absolute exit velocity for the first stage.* In the second stage the steam is expanded from pressure P 2 to that in the condenser and acquires initial velocity V a , leaving the last bucket with residual velocity V n - The theoretical velocities and blade angles for this stage may be found as above. Example: A two-stage Curtis turbine develops 800 horse power on a steam consumption of 12.5 pounds per horse-power hour, steam dry and saturated. Initial gauge pressure 135 pounds per square inch and back pressure 2 pounds per square inch absolute. Peripheral velocity 600 feet per second. The steam expands in the first stage from 150 pounds to 20 pounds absolute, and in the second stage from 20 to 2 pounds absolute. Angle of the nozzles with the plane of rotation, 20 degrees. Compare the performance of the actual turbine with its theoretical possibilities. Actual turbine: Steam consumed per hour = 800 X 12.5 = 10,000 pounds. Steam consumed per second = 10,000 -5- 3,600 = 2.78 pounds. Horse power developed per pound of steam flowing per second, 800 ■*- 2.78 = 288. Kinetic energy = 288 X 550 = 158,400 foot-pounds per second. Thermal efficiency, equation (68), 2545 E > " 12.5 (1191.1 - 94.4) = 187 Per Cent - Heat consumption, B.T.U. per horse power per minute, 12.5 (1191.1 - 94.4) 60 T? 1 8 7 Efficiency ratio = -=r = ^^ = 0.739. Ideal turbine: Stage velocities. The theoretical velocity in the first stage in expanding from a pres- sure of 150 pounds to 20 pounds absolute will be V l = 224\ / H l - H 2 = 224 V1191.1 - 1042.9 = 2727 feet per second. * In the actual turbine the velocities will be less than the theoretical on account of f rictional resistances in the nozzles and blades, and the velocity diagram must be modified as indicated in Fig. 172b and described on p. STEAM TURBINES 361 In the second stage the steam expands from 20 pounds to 2 pounds absolute. 7 = 224V#3-# 4 * - 224 V1042.9 - 914.8 = 2530 feet per second. The kinetic energy per pound of steam in each division of the first stage will be: In the first set of moving blades, E = (V 2 — V 2 ) = ^ (2727 2 - 1620 2 ) = 74,722 foot-pounds per second. The value of V 2 is conveniently obtained from the velocity diagram. In the second set of moving blades, E=t^-7(V 2 2 -V 3 >) 1 64.4 1 64.4 (1620 2 - 925 2 ) = 27,448. Total energy in first stage = 102,170. In a similar manner the total energy in the second stage will be found io be 93,700. Total for entire turbine = 195,870 foot-pounds per second. Theoretical horse power per pound of steam: Theoretical steam consumption per horse-power hour, 3600 ini , — - = 10.1 pounds. Sob Heat consumption, B.T.U. per horse power per minute, 10.1 (1191.1 - 94.4) 60 Thermal efficiency ratio, = 184. _ 1191.1 - 914.8 _ Er ~ 1191.1 - 94.4 " ° 253 - * In the actual turbine the heat contents H 2 , H 3 and H 4 will be greater than that of the ideal mechanism on account of frictional losses. See equation (76d). 362 STEAM POWER PLANT ENGINEERING SUMMARY. Horse power developed per pound of steam Steam consumption, pounds per H.P. hour B.T.U. consumed per H.P. per minute Thermal efficiency, per cent Efficiency ratio, per cent Actual Turbine. Perfect Turbine. 288 356 12.5 10.1 288 184 18.7 25.3 73.9 191. The Hamilton-Holzworth Steam Turbine. — Figs. 185 to 188 give a general view and some of the details of the Hamilton-Holzworth turbine, which belongs to the compound multi-stage " velocity " type. The steam flows through the annular space between rotor and stator CONDENSER OR EXHAUST PRESSURE Fig. 185. Principles of Hamilton-Holzworth Steam Turbine. as in the Parsons, but differs from the latter in that expansion takes place only in the stationary vanes. The rotor consists of a number of steel disks of varying diameters riveted to both sides of steel hubs and fitted at the periphery with drop-forged vanes as shown in Fig. 186. A tough steel ring is shrunk on the outside periphery of the vanes as indicated. The number of wheels and vanes is considerably less than in the Parsons type. The stationary vanes are fitted in steel disks as shown in Fig. 185, and the latter are located in grooves in the turbine casing. The vanes have a varying radial height increasing in the direction in which the steam flows. Sizes under 750 kilowatts STEAM TURBINES 363 have but one turbine casing, but larger sizes are divided into two, a high and a low-pressure turbine. The operation is as follows: Steam enters the high-pressure casing as indicated by arrows in Fig. 185, and passes through the first set of stationary vanes, extending around the whole periphery, which direct the steam at the proper angle against the wheel blades. In passing through the stationary vanes the steam is expanded down to the pressure in the first stage which is the same on both sides of the rotating disk. After giving up part of its energy, the steam expands again through the second set of stationary vanes to the pressure in the second stage, giving up energy to the second set of moving vanes. This pro- cess is repeated until the last stage is reached, from which the steam is discharged to the condenser in the simple turbine, or to the low-pressure steam chest in the compound turbine. In the low-pressure casing the steam is distributed in the same manner as in the high-pressure turbine. The diagram in the lower part of Fig. 185 shows the variation in steam pressure and velocity. The low-pressure front head is provided with an auxiliary nozzle which may be supplied with live steam in case of overload. The builders claim that since the pressures on both sides of the wheel are the same, no provision is necessary for axial balancing as in the Parsons standard turbine. Fig. 187 shows a sectional view of the bearing and stuffing box for the shaft at the point where it passes through the end of the turbine casing. The shaft is turned to a smaller diameter at its end and runs in a bushing G having a flange bearing against the inner side of the pillow block. At A is a cylindrical piece attached to and rotating with the shaft. This piece projects into an annular groove in the piece B, but it does not completely fill the groove and a circuitous passage is formed through which the steam must pass before reaching the stuffing box C. The object of the passage is to provide condensing DETAILS OP VANES. SECTION OF WHEEL. Fig. 186. Details of Vanes, Hamilton- Holz worth Steam Turbine. 364 STEAM POWER PLANT ENGINEERING surface so the steam itself will not reach the packing. The joint at the stuffing box is thus practically water-sealed. I Fig. 187. Details of Bearing, Hamilton-Holzworth Turbine. To prevent the oil from working into the turbine a bushing F is attached to -the shaft which throws off the oil into the space D by cen- trifugal force, where it drips down through a channel into a compart- ment in the pillow block. Any water escaping through the stuffing box is also collected in the same compart- ment. The bearing is oiled by a forced-oil system, the oil being sup- plied to the bottom of the bushing. Fig. 188 gives a diagrammatic view of the governor mechanism. A is the friction disk, L the roller, C the splined shaft with which the roller turns but upon which it is free to slide, bevel gears connecting shaft C with the throttle valve, and E a worm wheel driving disk A by means of a worm shaft F. At normal speed the gov- ernor sleeve is in mid position and roller L is at the center of the disk. If the turbine speeds up, however, the governor sleeve will rise, carrying with it the right-hand arm of lever T, which in turn will push the roller L a cor- responding distance downward. At the same time the cam H will be thrown to the right by contact with the roller R and by means of lever U will move the disk A and its shaft to the left, bringing it in Fig. 188. Governor Mechanism, Hamilton-Holzworth Turbine. STEAM TURBINES 365 contact with roller L, thus imparting a rotary motion to the shaft C and closing the throttle valve until the turbine assumes normal speed, when the several parts assume the first position. If the speed is reduced below normal the operation is just the same except that the various motions are reversed. At an increase in speed of 2 J per cent above normal steam is cut off entirely. HamiUon-Holzworth Turbine: Am. Elecn., Oct., 1904, p. 549; Eng. Rec, Oct. 1, 1904, p. 405; Power, Dec, 1907, p. 878, Nov., 1904, p. 659; Machinery, Nov., 1904, p. 134; Engr. U.S., Oct. 1, 1904, p. 690. 192. Westinghouse-Parsons Steam Turbine. — Fig. 189 shows a section through a Westinghouse-Parsons multi-stage reaction turbine. In this type no nozzles are employed and expansion of the steam is effected by a series of stationary and movable blades. The rotor is a steel barrel or drum divided into three sections of varying diameter, upon the periphery of which bronze blades are radially inserted in dove- tailed grooves. The adoption of three sections of varying diameter has no bearing on the design of this machine but is merely for mechanical convenience. The blades increase in length and cross section from the high-pressure to the low-pressure end of each section. The stator is of cast iron and its inner surface is studded with rows of blades projecting radially inward and conforming in size with the adjoin- ing blades of the rotor. The relative positions of the blades in the rotor and stator are shown in Fig. 190. The operation of the turbine is as follows: Steam enters at S, Fig. 189, through poppet valve V, which is actuated by the governor shown in detail in Fig. 191, and flows through the annular space between rotor and stator to the exhaust opening at B. The entire expansion is carried out within this annular compart- ment and resembles in effect a simple divergent nozzle with the excep- tion that the dynamic relationship of jet and vane is such as to secure a comparatively low velocity from inlet to exhaust. The velocity varies from 150 feet per second at the high-pressure end to about 600 feet per second as a maximum at the low-pressure end. The action of the steam on the blades is illustrated in Fig. 190. The steam strikes the first set of stationary blades as at P with initial velocity of about 150 feet per second and is deflected against the moving blades immedi- ately adjoining. In passing from P to P t the steam is partly expanded and gives up a portion of its energy to the moving blades. The steam is deflected from P x to P n and thus has a reactive effect on the moving blades in addition to the impulse imparted at P v The total torque produced at the shaft in element A is therefore due to impulse from 366 STEAM POWER PLAXT ENGINEERING STEAM TURBINES 367 1 and reaction from 2. This process is repeated in each element of the turbine, the steam expanding as it flows from element to element in its passage to the condenser. The angular velocity of the rotor UJ -J 5 CO Stationary ccuccuc HHHUD Blades "Moving Blades Vs Stationary ccccccccc )) )) )) B )) )) )) )) Blades Moving Blades Fig. 190. Flow of Steam in Parsons Turbine. varies from 3600 r.p.m. in a 400-kilowatt unit to 750 r.p.m. in the 7500- kilowatt size. Opposed to the three sets of blades the spindle also carries three rotating balance pistons P, P, Fig. 189, each of such diameter as Fig. 191. Governor Mechanism, Westingho use-Parsons Turbine. to exactly balance, through passage E, the axial thrust of the steam against its corresponding drum of blades. Steam enters the turbine intermittently as shown in Fig. 192, which represents indicator cards from a 1250-kilowatt turbine at various loads. 368 STEAM POWER PLANT ENGINEERING At light load the valve opens for a very short period and remains closed during the greater part of the interval. As the load increases, the period lengthens until finally, at about full load, the valve does not reach its seat at all, and continuous pressure is obtained in the high-pressure end of the turbine. The intermittent admission of steam is produced and controlled as follows: Lever T, Fig. 191, is given a reciprocating motion by an eccentric actuated by a worm and worm wheel on the main shaft. This motion is transmitted through lever H (with fixed fulcrum B) to lever A (with floating fulcrum D) and finally to pilot valve G. This reciprocating pilot valve admits puffs of steam from pipe to the under side of piston M, the rod R of which is attached to the admission valve V in Fig. 189. A spiral spring holds piston M in its lowest position until Fig. 192. Indicator Cards Showing Initial Pressure in a Westinghouse-Parsons * Steam Turbine. steam admitted by the pilot overcomes the spring tension and lifts the main valve from its seat, thereby permitting steam to enter the turbine. The fulcrum D of lever A is raised and lowered by the governor and therefore the pilot valve is controlled both by the motion of the eccen- tric and the motion of the governor. The eccentric keeps the pilot valve, and hence the main throttle, in constant oscillation, while the movement of the governor changes the limits of this motion. If an overload is sufficiently great to cause the governor balls to drop to their lowest position, the auxiliary or secondary valve V s , Fig. 189, begins to open and admits high-pressure steam to the later stage where the working steam areas are greater, thus increasing in proportion the total power of the turbine. The operation of this valve is the same as the main admission valve and is controlled by the governor. Fig. 193 shows the details of this mechanism. The speed varies about 2 per cent from no load to full load. STEAM TURBINES 369 In the smaller size machines running above 1200 r.p.m., flexible bear- ings are employed to absorb the vibration incident to the critical velo- city. They consist of a nest of loosely fitting concentric bronze sleeves with sufficient clearance between them to insure the formation of a film of oil. In the larger machines running below 1200 r.p.m. a split self-aligning bearing is used instead of the flexible bearing. The ends of the casing are fitted with water-sealed glands of special design to prevent the escape of steam or inflow of air at the point of entry of the shaft. The water used for sealing them is small in quantity and may be returned to the feed-water system. . Fig. 193. By-Pass Valve, Westinghouse-P arsons Turbine. Double-flow Type. — In reaction turbines of the single-flow type, as illustrated in Fig. 189, the high-pressure portion dealing with the high-pressure incoming steam is the least efficient. This is due to the fact that the blade lengths are approximately proportional to the specific volume of the steam, and consequently the initial expansion in the turbine requires blade passages of very small dimensions. This results in greater leakage past the tips of the blades than in the low- pressure elements where the blades are long. Again, in the single-flow type the high-pressure balance piston occupies fully one-half of the total balance piston length of the shaft, while the low-pressure piston is 2\ times the high-pressure diameter, so that balance pistons occupy a large portion of the total bulk of the machine. By making the high- pressure element of the impulse type and by arranging the low-pressure reaction elements on either side as illustrated in Fig. 193b the efficiency may be increased and the bulk of the turbine may be greatly decreased. 370 STEAM POWER PLANT ENGINEERING There are two rows of moving buckets upon the impulse wheel with an intermediate set of reversing blades, the operation being practically the same as in the first stage of a Curtis turbine. The drop in pressure in the nozzles is such that approximately 20 per cent of the total energy developed is absorbed by this impulse element. After leaving the im- pulse element the steam divides, one portion passing directly to the low-pressure blading at the left, while the rest passes through the hollow shell of the rotor to the similar pressure blades upon the right. As these sections are equal and symmetrical they counterbalance each other, and the balance or " dummy " pistons may be dispensed with. The advantages of the double-flow type over a single-flow unit of equal capacity are: (1) reduction of nearly 50 per cent in the shaft span willl Plan View Side View End View Fig. 193a. Method of Fastening Blades in Westinghouse-Parsons Turbines. between bearings; (2) the diameters of the casing and rotating part are more uniform, thus tending to greater rigidity; (3) a reduction of about 70 per cent in the bulk of the main parts of the machine, and (4) internal stresses due to high-pressure and high-temperature steam are avoided by isolating the incoming steam, without separate nozzle chambers. Westinghouse-Parsons turbines are made in a number of sizes, varying from 400 kw. to 15,000 kw. In Europe, however, Parsons turbines are made as small as 20 kw. and as large as 25,000 H.P. In sizes up to 3500 kw. the single-flow turbine has established itself as the most suitable prime mover, but for larger sizes the double- flow is given preference. The double-flow turbine is admirably adapted to low-pressure work. Fig. 194b shows a section through a Westing- house-Parsons double-flow low-pressure turbine. For results of tests of Parsons and W T estinghouse-Parsons turbines see Table 48. Double-flow Turbine: Power & Engr., March 16, 1909, Aug., 1908, p. 471 ; Eng. Rec., May 30, 1908, p. 693; Elec. Review, June 26, 1908, p. 1089. 10,000-kw. Westinghouse-Parsons Double-flow Turbo Generator for the Metropolitan Street Railway Company, Kansas City, Kansas: Power & Engr., May, 17, 1910, p. 890. STEAM TURBINES 371 $ 03 372 STEAM POWER PLANT ENGINEERING 192a. Allis-Chalmers Steam Turbine. — Fig. 193c shows a section through an Allis-Chalmers standard steam turbine, which is of the Parsons type but differs from the original Parsons machine and the Westing- house-Parsons construc- tion principally in manu- facturing details. In the older Parsons type, three balancepistons are placed at the high-pressure end. In the Allis-Chalmers de- sign, the larger piston is placed at the low-pressure end of the rotor, behind the last row of blades, the other two remaining at the high-pressure end. This construction per- mits of a smaller balance piston and allows a smaller working clearance in the high-pressure and inter- mediate cylinders. In the Allis-Chalmers turbine the roots of the blades are dovetailed and fitted into a foundation ring, and the tips are encased in a chan- nel-shaped shroud ring,, thereby insuring a rigid and positively spaced con- struction. The governor is of the Parsons type, ex- cept that the main valve and pilot valve are actu- ated by hydraulic instead of steam pressure. The bearings are of the self- adjusting ball and socket pattern and are kept " floating in oil " by a small pump geared to the turbine shaft. The oil is passed through a tubular cooler with water circulation after it leaves the bearings and is used over and over again. STEAM TURBINES 373 193. Elementary Theory, Parsons Turbine. — Fig. 194 gives a dia- grammatic arrangement of fixed and stationary blades in the first stages of a multi-stage ideal reaction turbine. The steam enters the stationary blades at practically zero velocity and is there partially expanded and impinges against the movable blades at velocity V v part of the energy of the steam being thus absorbed. In passing through the movable blades the steam is still further expanded and leaves at an absolute velocity V 2 , exerting an additional pressure on the blades from the reaction. The steam enters the second set of stationary blades with velocity V 2 and is still further expanded to velocity V 3 , and so on. "P=150 W=300 Fig. 194. Velocity Diagram. Westingho use-Parsons Turbine. The energy imparted to the steam in the first set of stationary blades is E,= W (78; V t = absolute velocity of the steam leaving the blades. The energy imparted to the steam in the first set of moving blades is W E n 2g (v 2 2 - v 2 ). (79) i\ = relative velocity of the steam entering the moving blades. v 2 = relative velocity of the steam leaving the moving blades. The total energy acquired by the steam in the first stage is E, + E 2 . The energy converted into work in this stage is WV 2 E = E t + E 2 - V^ 2 g V 2 = absolute velocity of the steam leaving the moving blades. Each stage may be analyzed in a similar manner. (80) (81) 374 STEAM POWER PLANT ENGINEERING Example: A Westinghouse-Parsons turbine develops 1000 horse power on a steam consumption of 12 pounds of steam per horse-power hour. Initial steam pressure 150 pounds per square inch absolute; back pressure 1 pound per square inch absolute; drop in pressure in each set of fixed and moving blades 15 pounds per square inch; peripheral velocity 300 feet per second; a x = a 2 = 30 degrees. Com- pare the performance of the actual and ideal turbine. Actual turbine: Steam consumed per hour, 1000 X 12 = 12,000 pounds. Steam consumed per second, 12,000 ^ 3,600 = 3.33 pounds. Horse power developed per pound of steam flowing per second, 1000 -^ 3.33 = 300. Kinetic energy per pound of steam, 300 X 550 = 165,000 foot-pounds per second. Thermal efficiency, g '- 12(llSS-70) - 18 - <>P " 0mt - Heat consumption, B.T.U. per horse-power hour per minute, 12(1191.2-70) _ 221 60 Efficiency ratio, Ml = 1^1 = 67.5 per cent. E r 28 F Ideal turbine : The velocity imparted to the steam in the first set of stationary blades due to the drop from 150 to 135 pounds per square inch is V t =224 VH 1 - H 2 =224 V1191.2- 1182.4 = 662 feet per second. Lay off the value of V r in direction and amount and combine with u, the peripheral velocity, Fig. 194. The resultant is v lf the velocity of the steam relative to the blades. The angle between v x and the line of motion of the wheel will be the angle with the blade at entrance. From the velocity diagram, v* = 429. STEAM TURBINES 375 E 2 , the energy given up by one pound of steam in expanding from 135 to 120 pounds, is E 2 =77S (H 2 -H 3 ) = 778 (1182.4-1172.8) = 7468 foot-pounds per second. Substitute v x = 429 and E 2 = 7468 in equation (79), 7468 = — (v 2 2 - 429 2 > 64.4 V 2 ; v 2 = 816 feet per second. The resultant of v 2 and u is V 2 , the absolute velocity of the steam leaving the moving blades of the first stage. From the diagram, V 2 = 573 feet per second. The energy converted into work in the first stage is determined by substituting the proper values in equation (81), thus: E = (662 2 + 816 2 - 429 2 -573*) -i- 64.4 = 9200 foot-pounds per second. The various stages may be analyzed in a similar manner. The theoretical output of the entire turbine per pound of steam will be that corresponding to adiabatic expansion from a pressure of 150 to 1 pound absolute. E= 77S(H t - H n ) = 778(1191.2-877) = 244,447 foot-pounds per second. Horse power per pound of steam, H.P. = ?M£17 _ 445 550 Steam consumption per horse-power hour, 3600 -, , =8.1 pounds. 445 Thermal efficiency, E = 1191.2-877 r 1191.2-70 = 28 per cent. 376 STEAM POWER PLANT ENGINEERING 194. Low-pressure and Mixed-pressure Turbines. — A promising field for the steam turbine is in its application as a secondary or low- pressure unit in connection with non-condensing or condensing engines, or, combined with a regenerator, in connection with engines using steam intermittently. Numerous examples may be cited showing great gains in both capacity and economy in existing power plants involving the Fig. 194a. Low-pressure Turbine Installation at the 59th Street Station of the Interborough Rapid Transit Company, New York. abandonment of but a negligible part of the equipment and accom- plishing this result with a minimum additional investment. The most notable installation (June, 1910) of low-pressure turbines to con- densing reciprocating engines is at the 59th Street Station of the Interborough Rapid Transit Co., New York. Three of the nine 7500-kw. Manhattan-type compound Corliss engines have been equipped with Curtis three-stage, low-pressure turbo-generators of equal capacity, and provision is made for the installation of six additional units. The low- pressure turbine is installed between the exhaust of the low-pressure cylinders and the condenser as shown in Fig. 194a. Running with the STEAM TURBINES 37T i^^^^^^^^^i y////////////rf& 378 STEAM POWER PLANT ENGINEERING engine the low-pressure turbine generator carries a variable load without governor regulation. The turbine generator takes care of the speed by automatically taking such a load as will keep the frequency in unison Engine Load K.W. 3000 | 4000 | 5000 Fig. 194c. Performance of 7500-Kw. Engine at 59th Street Station of Interborougli Rapid Transit Company, New York, with Varying Receiver Pressure. 19 18 K 17 u 8.15 13 a « 10 Act lal 1 'est) En§ ;ine _IU_ Jarai itee) Hif f h Pressu re r J ■"urbine i> _^A ■ fuit iue I*" Co nbic ed 1 :ngu ie~ai id-t jw-tr« 3 = A:- Constant Nozzle Pressure @ 16 #■ abs. B:- Variable Nozzle Pressure All Nozzles Open Cor Tur rectec sine £ I for team Moist ure li i 7000 8000 9000 10000 11000 12000 13000 Unit Load JC.W. 14000 15000 16000 Fig. 194d. Comparison of Economy Curves: 7500-Kw. High-pressure Turbine, 7500-Kw. Engine and Combined Engine and Low-pressure Turbine at the 59th Street Station of the Interborougli Rapid Transit Company, New York. with that of the engine-driven generator. The turbine is equipped with the usual emergency speed limit attachment for cutting off the steam supply should the speed exceed a predetermined limit. The STEAM TURBINES 379 performance of one set of engines, a high-pressure turbine of the equiva- lent total capacity, and that of the combined engine and low-pressure turbine, are illustrated in Fig. 194d. The conclusions drawn from an exhaustive series of tests at this station are that the addition of low- pressure turbines effected : a. An increase of 100 per cent in maximum capacity of plant. b. An increase of 146 per cent in economic capacity of plant. c. A saving of approximately 85 per cent of the condensed steam for return to the boiler. d. An average improvement in economy of 13 per cent over the best high-pressure turbine results. e. An average improvement in economy of 25 per cent (between the vfc:zGzzzz Fig. 195. Rateau Low-Pressure Steam Turbine Installation. limits of 7000 kw. and 15,000 kw.) over the results obtained by the engine units alone. /. An average unit thermal efficiency of 20.6 per cent between the limits of 6500 kw. and 15,500 kw. Low-pressure turbines are frequently installed in connection with regenerative accumulators, to rolling-mill engines, steam hammers, and other appliances using steam intermittently, and have proved to be paying investments. A typical installation of this character is to be found at the South Chicago Division of the International Harvester Company. The front elevation of the turbine and regenerator installa- tion is shown in Fig. 195 and the general arrangement of the regenerator is shown in Fig. 196. The regenerative accumulator is intended to regulate the intermittent flow of steam before it passes to the turbine. 380 STEAM POWER PLANT ENGINEERING The steam collects and is condensed as it enters the apparatus and is again vaporized during the time when the exhaust of the engines dimin- ishes or ceases. The regenerator consists of a cylindrical boiler-steel shell divided into two similar chambers by a central horizontal diaphragm. In each compartment are a number of elliptical tubes A, each of which is per- forated with a number of f-inch holes. The spaces surrounding the tubes and, under certain conditions, the tubes themselves are filled with water to a height of about four inches above the top of the upper tubes. Baffle plate B serves to separate the entrained moisture from the steam. The operation is as follows: Exhaust steam enters the apparatus at N, passes to the interior of the elliptical tubes, and escapes Fig. 196. Rateau Regenerator Accumulator. into the steam space through the perforations and thence to the turbine. When the supply of steam from the main engine ceases, the pressure in the regenerator decreases, the water liberates part of the heat it has absorbed and a uniform flow of low-pressure steam is given off. The continued demand of the turbine reduces the pressure in the accumula- tor and causes the steam still retained in the tubes to escape, thereby maintaining the circulation of the water (indicated by arrowheads) and facilitating the liberation of steam. Suitable valves regulate the limits of pressure in the accumulator and prevent the return of water to the main engine. Low-pressure turbines develop one electrical horse- power hour on a steam consumption of about 30 pounds with initial pressure of 15 pounds absolute and a back pressure of 1.5 pounds absolute. Fig. 197 gives the performance of the 500-kilowatt Rateau turbine at the International Harvester Works, South Chicago, I., , STEAM TURBINES 381 average initial pressure of 16 pounds absolute, condenser pressure 1.5 pounds absolute. Low-pressure turbines equipped with special expanding nozzles, or the equivalent, to receive steam at high pressure direct from the boilers are known as mixed pressure turbines. With this construction the full power of the turbine can be developed with (1) all low-pressure steam, (2) all high-pressure steam, (3) any proportion of high and low Reinforced Cinder Concrete Fig. 196a. Typical Double-deck Installation, Fort Wayne and Wabash Valley Traction Company, Spy Run Station. pressure steam. In the Curtis mixed-pressure turbine this transition from all low pressure to all high pressure, through all the conditions intermediate between these extremes, is provided for automatically by the turbine governor; a deficiency of low-pressure steam causes the high-pressure nozzles to open automatically. With this arrangement it is not necessary for purposes of economy to proportion exactly the low-pressure turbine to the amount of exhaust steam available, but within limits it may be made as large as the load demands. Low-pressure Turbines: Power & Eagr., July 6, 1909, p. 1, Nov 30, 1909, p. 905; Prac Engr. U. S., Mar. 1, 1909, p. 169; Eng. Mag., Apr. and May, 1907; Iron Age, Jan. 7, 19*09. 382 STEAM POWER PLANT ENGINEERING 195. Advantages of the Steam Turbine. — The principal advantages of the steam turbine are (1) simplicity; (2) economy of space and foundation; (3) absence of oil in condensed steam; (4) freedom from vibration; (5) uniform angular velocity; (6) large overload capacity; and (7) high efficiencies for large variations in load. The reciprocating engine is well adapted for pumping stations, direct-current generators, compressor plants, hoisting engines, and the like, requiring low angular velocity, but its place is being rapidly taken by the steam turbine for alternating-current dynamos, centrifugal pumps and blowers, requiring high angular velocity. — i rt mi i mi hi m 1 1 1 1 1 1 1 1 1 1 1 1 1 i — r i ii 1 1 ii 1 1 1 1 M ii i ii ii 1 1 ii i ii 1 1 1 1 li ' \ 1 ! 1 1 ^ 1 1 \ t 1 ^ di x _:::_ v v ±::::::_ : ::: ^ i \ ' - r; n _ it- \_ X_ _ _ J 50 - - _ s __ _ _ _ j---- ---- v t ::::: ::: \ _: :: \ it _:: _ 4^_ _ s _ _ _ __ \ : : _::: 5v * -- 'ST ~Z if) t "v ./ . % „? : z: : _ **« ,s" : ~: : *>* »88 OJE PC u- d-Co rli 3S- -I Si gsbr dj e J T-J P_e 0.2 i e ti -al T irpe Pylin rip rf ss NY Sdj — Pr 0.1 West a? he us e_- ?a sc ns -S ea w- Turrito aU I ■N< rmal Rate d Elec. H ara e Po ver 1000 2000 3000 400.0 Fig. 198. 5000 6000 8000 metric condenser is finding much favor, and in such instances the curves may be taken to indicate the relative floor space for the entire equip- ment of prime mover and auxiliaries. The weight of the steam turbine is very small compared with a recip- rocating engine of the same horse power. The New York Edison engines and generators weigh more than eight times as much as a turbine installation of equal capacity. The turbine for this reason, and also because of the total absence of vibration, requires a relatively light foundation. In many instances the foundation consists of steel beams with concrete arches sprung between them resting upon the floor, and the basement underneath may be used for the condenser instead of the massive foundation required for the reciprocating engine. 384 STEAM POWER PLANT ENGINEERING 198. Absence of Oil in Condensed Steam. — As the steam turbine requires no internal lubrication, oil does not come in contact with the steam, and the condensed steam from the surface condensers is available for boiler-feeding purposes without purification. In many cases the re-use of condensed steam effects a large saving in cost of feed water and in expense for maintenance and cleaning of boilers. The amount of entrained air is reduced to a minimum and consequently the work of the air pumps lessened. 199. Regulation. — The variable pressure at the crank pin of a reciprocating engine necessitates the use of a heavy fly wheel to keep the instantaneous angular fluctuation within practical limits. In the steam turbine the motion is purely rotary and a fly wheel is not neces- sary. In the former there are always instantaneous variations in velocity during each revolution, even with constant load, while in the latter the speed is practically constant. A number of published tests of Parsons and Curtis turbines show an average fluctuation of 2 per cent from no load to full load and 3 per cent from no load to 100 per cent overload. Although closer regulation than this is possible, it is not deemed necessary, particularly in alternating- current work where a comparatively wide range is desirable for parallel operation. 200. Overload Capacity. — A particular advantage of the turbine over the reciprocating engine lies in its greater overload. capacity and higher economy at overloads. The maximum economy of the average reciprocating engine lies between 0.75 and full load, whereas the turbine reaches its maximum at about 25 per cent overload. Thus a single turbine unit may economically take the place of two or more reciprocating units for a variable load. A turbine may be readily operated at 100 per cent overload, while the ordinary engine reaches its maximum capacity at about 50 per cent overload. In central lighting and power stations where there are one or more sharp peak loads of short duration, this extreme overload capacity is of marked importance. 201. Efficiency and Economy. — As far as steam consumption is concerned there is practically no difference between the performance of a high-grade piston engine and that of a first-class turbine for sizes under 2000 kw., the choice depending more upon rotative speed, over- load capacity and space requirements than upon the heat economy. For sizes over 2000 kw. the fuel consumption lies in favor of the turbine. A comparison of Fig. 148, showing typical economy curves of high-speed single- valve, non-condensing engines, and of Fig. 198c, show- ing similar curves for small non-condensing turbines, is somewhat STEAM TURBINES 385 in favor of the piston engine, though the difference is small; whereas a comparison of the turbine and engine curves in Fig. 194d, showing the performance of very large units, is decidedly in favor of the turbine. W 3000 Ph W S 3200 ? s | S800 ,v* 8 1> V v? V GO £T Steam Pressure 175* Back Pressure .68* Superheat 60° R.P.M. 2500 2400 ty* \ > k s 1 s 2000 \ \ 50 s \ 1000 \ \ 45 s 1200 V % f 40 | < Bf. 1 >/. -' K -t u Oj a 35 u £» *n ^ i 30 40 50 00 70 Brake Horse Power 80 90 Fig. 198a. Typical Performance of a 90 H.P. Terry Steam Turbine. Any number of individual tests may be cited showing superiority in fuel consumption of the piston engine over that of a turbine cf equiva- lent capacity and vice versa, but when the machines are designed for IP.14 12 5000 G000 rooo 8000 9000 Load in K.W. 10000 11000 12000 Fig. 198b. Typical Performance of 9000 Kw. Curtis Turbine; 200 Lbs. Gauge Pressure, 125° Superheat, 29 Inches Vacuum. the same operating conditions the results are practically the same for all sizes under 2000 kw. Tables 39 to 43 give the general condition of operation and the steam consumption of exceptionally good piston engines of various sizes and types, and Table 48 similar data of first- 386 STEAM POWER PLANT ENGINEERING class turbines. A study of these tables will show that the choice must be based on other factors than the steam consumption. In a general sense, the piston engine is superior to the turbine for high back pres- sures, slow rotative speeds and heavy starting torques, while the tur- bine has practically superseded the engine for large central station units and for auxiliaries requiring high rotative speed. Recent tests of the Melville reduction gear (Machinery, Feb., 1910) show exception- ally high efficiencies for sizes as large as 6000 kilowatts, and it is not 80 Steam Press. -150 Lb 70 w ft W 60 ft u A Atmospheric Exhaust J i^s. 40 ° R 3 -M. ■£°H^ B C C D A \ s ^ si fi 50 o ft ^ LOH.p II & 40 E fiiJOJi ej^ iSSj^ 1. B C * — Si 55s 2t£oo J^M * i E 1 h- 2°i •P. E 30 £URTls_ E' 20 1 2 4 Lo G. IS ad 8c. 4 A : h i '4 unlikely that the turbine equipped with this device will offset the low rotative speed factor of the piston engine. If the tests of steam turbines and piston engine could be made at some standard initial pressure, back pressure and quality or superheat, then a comparison could readily be made, but both types of prime movers are designed to give the best results for special operating conditions, and any marked departure from these conditions will result in loss of economy. It is frequently desired, however, to make a comparison between the economy of the different machines, and the following methods are in vogue: (1) Steam consumption under assumed conditions. (2) Heat consumption per horse-power per minute above the ideal feed-water temperature. (3) Efficiency ratio or ratio of ideal to actual. STEAM TURBINES 387 Standard Correction Curves. This method for comparing engines or turbines or both is best illus- trated by a specific example: Suppose it is required to compare the full-load performance of a 125-kw. direct-connected piston engine with that of a 125-kw. turbo-generator with operating conditions as follows: Steam Consump- tion, Lbs. per Kw.-Hour. Initial Pres- sure, Lbs. Absolute. Vacuum, Inches of Hg. Superheat, Deg. F. Engine Turbine 25.0 22.7 160 110 25.5 28.0 125 Manufacturers of steam turbines have provided correction curves as illustrated in Fig. 194d, showing the influence of varying vacuum, superheat and pressures on the steam consumption.* From curve B, we find that the steam consumption of the turbine should be 100 110 Steam Pressure, Lbs. per Sq. In. Absolute 120 130 140 150 160 170 180 190 200 30 t-1 I 24 3 20 A 0°F Superheat,165 *"Abs. B » » , 28 in Vacuum C 165* Absolute, » » A ^Vi c 0li t u-Co p^o^C W B Pr e& sure Corr ectic 23 1 C g"Perhei rve rrec ion ?urv e TYPICAL CORRECTION CURVES FOR 125 K.W. STEAM TURBINE FULL LOAD CONDITIONS. 20 40 CO 80 100 120 140 Superbeat, Deg. Fab. 160 20 21 22 23 24 25 26 Vacuum, Inches of Mercury M 180 200 29 Fig. 198d. decreased 2.5 pounds to give the equivalent at 160 pounds initial pressure; from curve A it should be increased 2.5 pounds to give the equivalent at 25.5 inches of vacuum, and from curve C it should be increased 2.5 pounds to give the equivalent at degree superheat. The full-load steam consumption for the turbine under the engine con- ditions is therefore 22.7 — 2.5 + 2.5 + 2.5 = 25.2 pounds per kw.-hour. * These curves are drawn to a much larger scale than the reproduction given here. 388 STEAM POWER PLANT ENGINEERING The ratio method is also used in this connection, thus: The full-load steam consumption at 160 pounds pressure, curve B, Fig. 194d, is 25 multiplied by the ratio 7 r=- - to give the equivalent consumption at 110 Zt .o pounds (25 is the steam consumption at 160 pounds and 27.5 the con- sumption at 110 pounds). Similarly the correction ratio to change 25 5 the consumption at 28 inches of vacuum to 25.5 is -^-, and to correct 25 125° F. superheat to 0° F. is — -• Summary. 25 Pressure correction 7 —— = 0.91 = 11 .5 - 9%. . 27.5 Vacuum correction — -^ = 1.10 = lb 10%. 25 Superheat correction 7 ^— = 1.11 = 11%. Net correction 12%. Corrected steam consumption = 22.7 + 0.12 X 22.7 = 25.4 pounds per kw.-hr. The ratio method is generally used if the difference between the corrected steam consumption and that of the correction curves for the same conditions is greater than 5 per cent ("The Steam Turbine," Moyer, p. 128). This ratio method for correcting steam consumption at full load may be used without appreciable error for half to one and one half load and is the only practical method for quarter load (Engrg. London, March 2, 1906). Heat Consumption. The heat consumption B.T.U. per unit output per minute above the ideal feed water temperature may be expressed — ^-~- — — • See equation (/5b). For the case cited above Engine, 25 (1194.1 - 60 98) = 455 B.T.U. Turbine, 22.7 (1264.2 -70) = 451 B.T.U. STEAM TURBINES 389 Efficiency Ratio. The efficiency ratio, or the extent to which the theoretical possibilities are realized, may be expressed 2545 Er W {H x - H 2 ) For the case cited above ™ • 2545 Engine, Turbine, 25(1194.1 - 915) 2545 22.7 (1264.2 - 915.3) See equation (76b). = 0.366. = 0.322. In the assumed case the turbine is the more economical in heat con- sumption, but the engine is the more perfect of the two as far as theo- retical possibilities are concerned. 30 29 \A u 27 §26 r \- — -Wewii (ghouse - Paisons ' "■ iV / 1250 Kv '. Tuubiu \f. Sllp€ tlieat 7b. :f *2l ■a B V\ \\ ■s " % 22 • 21 \ \ & \\ 5* v \ o N> V 3>-." c \a b>S^ s\ i ^^T f^fi 2 11 \ \* tr, n Avti K,00 J igeof'.ht osh fo Se ',*. Engir :e J lunour ^^^ r42i£u a- es o - st pir,e ii xo 4 .i .6 .7 .8 Load in terms of rated electrical load Fig. 199. 202. First Cost. — Steam turbines, generally speaking, are about 10 per cent lower in first cost than high-grade compound engines of equiv- alent power. The following table gives an idea how the price varies with the conditions of operation. The figures are approximate only and refer to the cost of the turbine and generator exclusive of auxil- iaries. 390 STEAM POWER PLANT ENGINEERING CM CM CO i— i i— i o ■* M DO CM CM CM i-l 00 t- »o CM » OS • O CM OO OS O »0 CM ^H tH »o »o *' *' * Pi *: fe *: P.' m m m w w a W IO »0 »0 CM CO . CM i>. CM i-H IO CO O CO CO CO OS m Ho ^ CP Ok CM o3 QQ -4-3 . I* it £ ° d o o T3 T3 QQ o o *. c bfi ^ a : Tt< oo CO ^ A ^ S S -3 r - > be o r - ~ a. o UJ . • « h <1 •^Kred^o pai^H £ £ £ ^ ^ ^ o o o o o o o o o o o >o OS IO CM £ £ £ ^ ^ r*) o o o o o o io »o — i £ £ £ ^ r* ^ io «o »o w co co i-i o o o -a -a -c o o o TJ T3 13 xapui t^ 00 OS o STEAM TURBINES 391 o ■<*» CO lo (M CM O ^H CM CM CO CM CO -^ LO LO CM CM CO CO o O CO t^ CO CO CO ^ OJ O) y-i CM CO i— I i— I LO . CO OO lo C5 OS o t~- o CO CO oo o o _ -* CO »* LO OS LO 05 r^ CO OS "tf t~ LO ■* >* ■* OS lo iO LO »o lO lO CM T* CM CM t^ o co OS CM CM CM CM CM OS OS co CM OO LO CO o CO o O O o o O OS o o o o o rH ,—l 1-1 rH 7-1 *"" ' 1-1 i—i o 1-1 rH 1-1 '-' 1-1 *" H ^ W Ph PU W W LO <* Tf LO CO LO Dh cm LO I>- ■8 ~ > S H o o #2 o o 13 13 H o o 13 13 13 oo O CM OQ bJO _ s £> feO 13 ci : o o o o 13 13 1! T3 o o I * t^ i-i 13 73 O o o J£ 13 13 ^ o o o co £ £ £ £ £ £ ^ ^ ^ ^ ^ ^ O O O O O T* O O O O LO CM O LO CM i-H o fc>C "^ 5 a O CD OQ o o o o 73 13 13 13 o o o 13 13 13 o o 13 13 O O 13 13 o o o o 13 13 13 13 392 STEAM POWER PLANT ENGINEERING APPROXIMATE COST OF STEAM TURBINES AND GENERATORS. In Dollars ner Kilowatt. Rated Capacity. Kilowatts. 25 55 60 50 47 51 75 42 46 100 38 43 200 32 36 300 500 1000 2000 4000 6000 Direct current: Non-condensing .... 32 36 35 35 Condensing Alternating current: 25 cycles 32 30 28 27 25 25 21 21 20 60 cycles 20 203. Cost of Operation. — Data pertaining to the cost of operating steam-turbine and reciprocating engine plants and combinations of both will be found in Chapter XVII. The following table, contributed by H. G. Stott, Superintendent Motive Power of the Interborough Rapid Transit Company, New York, gives an excellent comparison of the relative maintenance and operating costs (March, 1910) of the three types of steam power plants as applied to large central stations for electric street railways. RELATIVE COSTS PER KILOWATT-HOUR. DISTRIBUTION OF MAINTENANCE AND OPERATION. Reciprocating Steam Plant. Steam Turbine Plant. Reciprocating Engines and Low- pressure Steam Turbines. Maintenance. 1. Engine room, mechanical 2. Boiler or producer room 3. Coal and ash handling apparatus. . . . 4. Electrical apparatus Operation. 5. Coal 6. Water 2.59 4.65 0.58 1.13 61.70 7.20 6.75 7.20 2.28 1.07 2.54 1.78 0.30 0.17 0.51 4.33 0.54 1.13 55.53 0.65 1.36 6.74 2.13 0.95 2.54 0.35 0.30 0.17 1.55 3.55 0.44 1.13 46.48 0.61 7. Engine room labor 4.06 8. Boiler or producer room labor 9. Coal and ash handling labor 5.50 1.75 0.81 11 Electrical labor 2.54 12 Engine room lubrication 1.02 13. Engine room waste, etc 0.30 0.17 Relative operating cost, per cent. . . . Relative investment, per cent Probable average cost per kw 100.00 100.00 125.00 11% 77.23 75.00 93.75 11% 69.91 80.00 100.00 11% For steam turbine plants larger than 60,000 kw. the cost per kilowatt may be reduced to $75.00. STEAM TURBINES 393 204. Influence of Superheat. — The use of superheated steam in- creases the economy of the reciprocating engine about 1 per cent for every 10 to 20 degrees of superheat, depending upon the conditions of operation, the gain being due mainly to the reduction of cylinder con- densation. Cylinder condensation is reduced not only because of the excess heat available for the evaporation of moisture but also because superheated steam has a lower conductivity than wet steam, and less heat is given up to the cylinder walls for the same difference of tem- perature. In the steam turbine this difference of temperature is much smaller, since high- and low-pressure steam do not alternately come in « 1 1 i I i 10 RELAT K-W. Westinghouse-Parsons Turbine i0 I r„ £? vj 28 inches Vacuum Tested by Messrs. Dean & Main Eng'rs. it) -. te ai J -c OJ ^Ll >fl ? i~ Cc -*d t**- 3o % i 52 S 13 o T ^-2 ^ id 1 H -J « 12 u a—- 5-i a ~ - o n, — •^. --■ ^ > c ^ "> >< ^ n» &, V wgp \ 11 1 33* ^ gp5 v~ i & m 5- ? £ ^ & 10 ^^ K ft Sup srl 1« It Ueg. V. SJ0 40 80 100 Fig. 200. 120 140 160 180. contact with the same surface as is the case with the reciprocating engine, and the time of contact is considerably less, due to the com- paratively high velocities. With a well-lagged casing, therefore, the condensation due to this cause is insignificant compared with that of the reciprocating engine, and the beneficial effect of superheat is much more pronounced. Friction of the steam, which in the reciprocating engine is negligible, and which may be a source of considerable loss in the turbine, is greatly reduced by the use of superheated steam, as is also the " windage " loss due to the rapid revolution of the wheels. The problem of cylinder lubrication is sometimes a difficult one in steam engines using a high degree of superheat, and trouble is fre- quently experienced due to the unequal expansion of the metal. In 394 STEAM POWER PLANT ENGINEERING the steam turbine the latter difficulty is not so pronounced and no internal lubrication is necessary, hence higher degrees of superheat are permissible. For maximum economy the steam at the end of expansion should be free from moisture. Assuming purely adiabatic expansion, the steam in expanding from 165 pounds to 1 pound abso- lute would have to be superheated about 500 degrees F., giving the steam an actual temperature of 800 degrees F. A study of some 100 tests made in this country gives about 250 degrees superheat as a maxi- RELATI ON- VACUUM TO ECONOMY 3 K.W. Parsons Turbine 36 30 Hulton Colliery '.34 j u 3* o \ P]ul L W No Su ^heat 30 > sL <^" r ~> 28 h^ *«j pj 3p ^ 26 "^ N. 34 35 / 7 V ' 1 22 30 -6- 2 o > / N k \ 20 — 25 3- o -5- i / / 20 +3 o ^ > O* k > <* / a 9 f ^ 1 g jj5 £> P-. ■, v£> _^ & ' 10 09 L& % 3>, y 2 a p^ V*5 p* tt iVC ££ t? ^1 o &z jf Lg< re ase \ r ac jum [nohe P 8 1Q 12 14, 16 18 20 22 34 Fig. 201. 28__30 mum and 100 degrees to 150 degrees F. as an average. In Europe reciprocating engines are operating with superheat as high as 350 degrees F. and turbines 300 degrees F. The additional fixed and operating costs of superheating must be considered in determining the net gain, since the decrease in steam consumption does not represent the actual saving. With pressures of 175 pounds gauge or less, and not to exceed 200 degrees F. superheat, the net gain has in most cases proved a substantial one. With higher temperatures and pressures the cost of maintaining the superheat may increase more rapidly than the saving in steam consumption, until a limit is reached beyond which no STEAM TURBINES 395 advantage is gained. The relation between superheat and steam con- sumption for a 400-kilowatt Westinghouse-Parsons turbine is illustrated in Fig. 200. Fig. 202 gives a similar comparison for a 1500-kilowatt turbine. (J. R. Bibbins, Power, January, 1905.) 205. Influence of High Vacua. — The possible economy of the recip- rocating engine is greatly restricted by its limited range of expansion. Cylinders cannot be profitably designed to accommodate the rapid increase in the volume of steam when expanded to very low pressures. For example, the specific volume of 1 pound of steam under a vacuum of 29 inches (referred to a 30-inch barometer) is about 650 cubic feet, or nearly double its volume under a vacuum of 28 inches. Usually the exhaust is opened at a pressure of 6 or 8 pounds absolute and consequently a large proportion of the available energy is lost. The lower vacuum in the exhaust pipe, therefore, serves only to diminish the back pressure and does not affect the completeness of expansion. Even if it were practical to expand to 1 pound absolute, the increased condensation in the reciprocating engine would offset any gain due to expansion unless the steam were highly superheated. A study of a number of tests of reciprocating engines shows a slight improvement due to increasing the vacuum beyond 26 inches. Tests of steam tur- bines show a decrease of 3 to 4 per cent in steam consumption for each inch increase of vacuum between 25 and 29 inches, for with a well- lagged casing cylinder conden- _ sation is practically absent, since the high- and low-temper- ature steam do not alternately come in contact with the metal- lic surfaces as is the case with the reciprocating engine (Figs. 201 and 201a). Fig. 230 shows a relation between the power con- sumption of the auxiliaries and the total output of the station at different loads for a Parsons steam turbine installation and Fig. 231 shows a similar relation for the 2000-kilowatt Curtis tur- bine. The power consumption in the latter case is higher on account of the high temperature of cool- ing water. Table 55 gives the power required for the auxiliaries in a number of stations. A high vacuum may be limited by the initial tem- perature of the cooling water. The difference in temperature between I Mill II 1 1 1 1 1 1 | 1 1 1 1 1 II 1 1 1 1 1 1 1 1 1 ft k% W ID No. Superheat i %J.__ II IJ *te* N.B. Observations at 28 "takeu froi a subsequent test under same v conditions J .5 4- \A A fc*> s 9- a ^ H ^ n; .5 i- CO "* tri lftl s ■r -a>- > •*. <** £t° U -1 a ■o | ■ o. n 5jJ »io ©so* ^5-1 _ 1—1 «* K,V« a ~U .5 -{ t — £ ^get cl - 1 ^ v.^ ■?. i — k m It 1 1? ^ 1 r ,uu P i- 25 .2 A .6 I .2 .4 .6 .8 Fig. 201a. 27 .2 .4 .6 .8 28 396 STEAM POWER PLANT ENGINEERING 16 15.5 15 04.5 14 H'3.5 03 12.5 02 , ^ Efi. ect of Vacuum and Superheat a s. ^> 9 on Steam Consumption 1500 K.Vf . .Turbine ,Full Load a 0 Lbs. Stc 28 Inche im Pressui Vacuum j 3 l o. OLb Dry I. Steam Pressu Saturated Steam e 1 « ^ a fe* c/3 26 27 Vacuum Inches 20 40 60 80 100 120 140 Superheat Deg.E. Fig. 202. inlet and discharge should be greater than 10 degrees, since otherwise the amount of circulating water per pound of steam becomes excessive and increases the work of the pumps. For example, the temperature of steam corresponding to a vacuum of 28 inches or 1 pound absolute is 102 degrees, and with cooling water at 75 degrees F., and the discharge at 95 degrees F., the theoretical ratio of cooling water to steam necessary to maintain this vacuum will be about 50 and the actual nearer 70. From Table 50 it will be seen that a 28-inch vacuum referred to a 30-inch barometer is obtained with an average ratio of 50 pounds of cooling water at 70 degrees F. per pound of steam. The cost of high- vacuum apparatus is not proportional to the vacuum, but increases much more rapidly, as shown in Fig. 232. These estimates show averages and not specific costs. Fig. 202 shows the effect of superheat and vacuum on the economy of a 1500-kilowatt Westinghouse-Parsons turbine. Figs. 200 to 202 are taken from " Steam Power Plants," by J. R. Bibbins, as published in Power, January, 1905. Reciprocating Engine vs. Turbine: Power, April, 1904, p. 232, May, 1904, p. 298; Engr. U. S., Nov. 1, 1905, p. 711; Elec. World, April 2, 1904, p. 651; Eng. Mag., Sept., 1905, p. 935; Power, Feb., 1906, p. 83; Elec. Age, June, 1905, p. 478; Elec. Rev., Dec. 23, 1904. Steam Turbine Design: Eng. Rec, July 22, 1905, p. 101 ; St. Ry. Jour., Dec. 20, 1902, p. 988; Engr., Lond., Jan. 8, 1904, p. 34, May 13, 1904, p. 481, Dec. 27, 1907; Electrician, Lond., March 24, 1905; Power, Dec, 1905; Mech. Eng., Feb. 7, 1908; Engineering, Dec. 13, 1907. Theory and Design of Steam Turbines: Engr. U. S., Dec. 16, 1907, p. 1126 (serial), March 15, p. 201; Revue de Mecanique, Oct. 31, 1907; Engr., Lond., Oct. 4, 1907; Eng. Rev., May, 1904; Mech. Engr., Oct. 28, 1905; Eng. Rec, July 22, 1905, p. 101, May 7, 1904, p. 581. Modern Steam Turbine Plants: Power, Dec, 1906, p. 717, Dec, 1907; Engr. U.S., Nov. 15, 1906, p. 733, March 15, 1907, p. 304; St. Ry. Jour., Oct. 19, 1907; Elec. World, July 22, 1905, Feb. 15, 1908; Eng. Rec, March 4, 19Q5. Tests of Westinghouse-Parsons Turbines: Engr. U.S., Dec. 1, 1904, p. 802; St. Ry. Jour., Dec. 19, 1903, p. 1063; Power, March, 1904, p. 127, April, 1904, p. 239, Aug., 1905, p. 466; Elec. World, Sept. 6, 1902, p. 360; Eng. Rec, July 29, 1905, p. 134. Governing Steam Turbines: Harvard Engineering Jour., 1908. A Recent Comparison of Turbines and Engines: Eng. Rec, Feb. 19, 1910. Internal Losses of a Steam Turbine: Power & Engr., Aug. 24, 1909. The Principles of Steam-turbine Buckets: Power, Mar. 17, 1908. CHAPTER XI. CONDENSERS. 206. General. — A pound of dry steam at atmospheric pressure (30 inches mercury) occupies a volume of 26.8 cubic feet. Suppose these 26.8 cubic feet of steam were contained in a closed vessel, and that the steam was subsequently condensed and its temperature lowered by suitable means to say 110 degrees F. The condensed steam would occupy only about T 7V0 °f its original volume, and the pressure would fall to 2.6 inches of mercury, the latter pressure being due to the ten- sion of the aqueous vapor at the given temperature. That is to say, the best vacuum theoretically attainable under the given conditions would be 30 — 2.6 = 27.4 inches. The lower the temperature to which the condensed steam is reduced the more nearly perfect will be the vacuum attained. If air is mixed with the steam the vacuum will be still more imperfect. Thus, suppose the vessel to contain one pound of steam and one-tenth of a pound of air under atmospheric pressure. The volume of the closed vessel in this case must be 26.8 + 1.69 = 28.49 cubic feet. After the steam has been condensed and its temperature reduced to 110 degrees F. the remaining pressure will be due to the aqueous vapor tension plus the pressure due to the air, or 2.6 + 1.51 = 4.11 inches mercury, and the maximum vacuum attainable will be 25.89 inches. In practice air is always present in exhaust steam. A condenser is a device in which the process of condensation and subsequent removal of the air and condensed steam is continuous, the degree of vacuum obtained depending upon the tightness of valves and joints, the quan- tity of entrained air, and the temperature to which the condensed steam is reduced.* The degree of vacuum may be expressed in different ways. (1) Excess of the atmospheric pressure over the observed vacuum. For example, a 26-inch vacuum implies that the pressure of the atmosphere is 26 inches of mercury above the pressure in the condenser. (2) Per cent of vacuum, by which is meant the ratio of the observed vacuum to the atmospheric pressure. Thus with the barometer standing at 30 inches a vacuum of 26 inches may be expressed as 100 X ff = 86.6 per cent vacuum. This method of expression gives an idea of the * See " The Influence of Air on Vacuum in Surface Condensers," Engng., April 17, 1908; Power, Feb. 2, 1909, p. 235. 397 398 STEAM POWER PLANT ENGINEERING efficiency of the condensing system. For example, the degree of vacuum indicated by 26 inches would be 93 per cent with a barometric pressure of 28 inches but only 84 per cent when the barometer reads 31 inches. (3) Absolute pressure. Thus a 26-inch vacuum referred to a 30-inch barometer would be indicated as a pressure of 30—26 = 4 inches absolute, or 1.9 pounds per square inch. 207. Function of the Condenser. — The function of a condenser in connection with a steam engine or turbine is primarily the reduction of back pressure, though in some instances, notably in marine work, the recovery of the condensed steam may be of equal importance. The advantages to be gained by decreasing back pressure may be most readily illustrated by the following example: A non-condensing engine taking steam at a pressure of 100 pounds absolute and cutting off at one-quarter stroke will have, theoretically, a mean effective pressure on the piston of 44.6 pounds per square inch, the back pressure being 14.7 pounds per square inch absolute. If the engine exhausts into a condenser against a 26-inch vacuum (1.7 pounds absolute) the mean effective pressure will be increased to 44.6 + (14.7 — 1.7) = 57.6 pounds per square inch, resulting in a gain in power which may be expressed H.P. = IjzA/L , 33,000 in which (82) H.P. = horse power gained. P r = reduction in back pressure, pounds per square inch. A = area of the piston in square inches. S = piston speed in feet per minute. If P = mean effective pressure on the piston when running non-con- densing, the percentage of increase of power may be expressed Per cent = 100 Pr (83) In the above example the percentage of power gained would be 13 100 44.6 29.2 per cent. The actual gain due to the use of the condenser would be much less than this, depending upon the type of engine and conditions of operation, as shown in the results of engine performances outlined in Chapter X. CONDENSERS 399 TABLE 49. PRESSURE OF AQUEOUS VAPOR IN INCHES OF MERCURY FOR EACH DEGREE F. (Marks and Davis.) 0° 1° 2° 3° 4° 5° 6° -0 8° 9° 30° .180 .268 .390 .560 .790 1.10 1.51 2.04 2.74 3.63 4.76 6.18 .188 .278 .405 .580 .817 1.13 1.55 2.11 2.82 3.74 4.89 6.34 .195 .289 .420 .601 .845 1.17 1.60 2.17 2.90 3.84 5.02 6.51 .203 .300 .436 .622 .873 1.21 1.65 2.24 2.99 3.95 5.16 6.67 .212 .312 .452 .644 .903 1.25 1.71 2.30 3.07 4.06 5.29 6.84 .220 .324 .468 .667 .964 1.30 1.76 2.37 3.16 4.17 5.43 7.02 .229 .336 .486 .690 .996 1.33 1.81 2.44 3.25 4.28 5.58 7.20 .238 40° .248 .362 .522 .739 1.03 1.42 1.93 2.60 3.44 4.52 5.88 .257 .376 .541 .764 1.06 1.46 1.98 2.66 3.53 4.64 6.03 .349 50° 60° 70° .503 .714 1.03 80° 1.37 90° 100° 110° 1.87 2.51 3.34 120° 130° 4.40 5.73 140° 7.38 With steam turbines the advantage gained by reduction of back pressure is more marked than with the reciprocating engine, though theoretically the same for the same range of expansion. Initial con- densation, leakage past valves, and other sources of loss prevent a reciprocating engine from benefiting from a good vacuum to the same extent as a turbine. (See paragraph 205.) Referring again to the example given above, if the steam is cut off at about one-sixth stroke, the work done when running condensing will be the same as when running non-condensing and cutting off at one-quarter. Theoretically the steam consumption will be decreased nearly in pro- portion to the reduction in cut-off. Generally speaking, a condensing engine will require from 20 to 30 per cent less steam for a given power than a non-condensing engine. (See results of engine tests, paragraph 179.) This decrease in steam consumption is only an apparent one. If steam is used by the auxiliaries in creating the vacuum, the amount must be added to that consumed by the engine, unless the steam exhausted by the former is utilized to warm the feed water, in which case only the difference between the heat entering the auxiliaries and that returned to the heater should be charged against the engine. The power neces- sary to operate the condenser auxiliaries varies from one to six per cent of the main engine power, depending upon the type and conditions of operation. (See paragraph 228.) In power plants where the exhaust steam is not used for heating or manufacturing purposes, the engines are almost invariably operated condensing, provided there is an abundant supply of cooling water. Even if the water supply is limited, it is often found to be economical to 400 STEAM POWER PLANT ENGINEERING use some artificial cooling device, notwithstanding the high first cost and cost of operation of the latter. Some of the considerations affecting the propriety of running con- densing and the choice of condensing systems are taken up in para- graphs 230 and 231. The Law of Condensation of Steam : Pro. Inst, of Civil Engrs., Nov. 30, 1897; Engr., Lond., Dec. 17, 1897, p. 609. Relation of Pressure and Temperature in Con- densers : Power, April, 1902, p. 28, June, 1902, p. 30. The Measurement of Vacuum : Engr., Lond., April 21, 1905. Experiments on Condensation of Steam : Engr., Oct. 15, 1897, p. 481. Condensation Fallacies and Facts: Machinery, Sept., 1904, p. 38; Elec. Review, June 17, 1904. Condenser Pressure, a Neglected Point in Steam Condensers : Mech. Engr., Sept. 30, p. 484. Advantages of Condensing : Amer. Elecn., Sept., 1904, p. 469. Condensers for Steam Engines and Turbines (F. Foster) : Mech. Engr., Oct. 28, 1905, pp. 637, 655. Condensers, to What Extent Should They be Used : Power, Nov., 1899. The Value of the Condenser : Power, Oct., 1902. The Effects of Vacuum on Steam Engine Economy : Eng. Mag., June, 1905. Importance of Condensers : Power, May, 1897; Evaporating, Condensing, and Cooling Apparatus (E. Hausbrand), published by Scott, Greenwood & Co., London. The Use of Condensers : Electrician, Lond., Oct. 12, 1900. 208. Classification of Condensers. — The following is a classification of a few well-known steam engine condensers : 1. Jet condensers. 2. Surface condensers. Parallel current (a). Counter current (6). Worthington. Blake. Deane. Baragwanath. Bulkley. Ejector (3) Schutte. Weiss. Barometric. Ordinary (1) Siphon (2).., Alberger. Tomlinson. Single-flow Baragwanath. Double-flow Wheeler. Multi-flow Wainwright. Forced draft Fouche. i Natural draft Pennell. Evaporative (c) Ledward. Water cooled (a) Air cooled (6) . . . . Condensers may be divided into two general groups: 1. Jet condensers, in which the steam and cooling water mingle and the steam is condensed by direct contact, Figs. 203 to 211. 2. Surface condensers, in which the steam and cooling medium are in separate chambers and the heat is abstracted from the steam by con- duction, Figs. 212 to 215. CONDENSERS 401 Jet condensers may be further grouped into two classes, according to the direction of flow of the air and cooling water: (a) Parallel-current condensers, in which the condensed steam, cool- ing water, and air flow in the same direction, collect at the bottom of the condenser chamber, and are exhausted by a suitable pump, Fig. 203. (b) Counter-current condensers, in which the cooling water and con- densed steam flow from the bottom of the chamber, usually by gravity, while the air is drawn off at the top, Fig. 229b. Parallel-current condensers may be subdivided into three classes: (1) Standard condensers, in which the cooling water, condensed steam, and air are exhausted by a vacuum pump, Figs. 203 and 209. (2) Siphon condensers, in which the cooling water, condensed steam, and air are exhausted by a barometric column, Fig. 205. (3) Ejector condensers, in which the condensed steam and air are exhausted by the cooling water, on the ejector principle, Fig. 206. Surface condensers may be classified according to the nature of the cooling medium as (a) Water-cooled condensers, Fig. 212. (b) Air-cooled condensers, Fig. 217. (c) Evaporative condensers, in which the condensation of the steam is brought about by the evaporation of a fine stream of water trickling on the outside of the tubes. 309. Ordinary Jet Condensers. — Fig. 203 shows a section through a Worthington jet condenser, illustrating the parallel-current principle. When the pump is started a partial vacuum is created in the suction chamber above the valves H, H in the cone F. As soon as sufficient air has been exhausted, cooling water enters at B with a velocity depend- ing upon the degree of vacuum in chamber F and the suction head, and is divided into a fine spray by the adjustable serrated cone D. The spray mingles with the exhaust steam entering at A and both move downwards with diverse velocities. The steam gives up its heat to the water and condenses. The velocity of the steam diminishes in its downward path to zero, while the velocity of the water increases accord- ing to the laws of falling bodies. The condensed steam, cooling water, and air collect at the lower part of the condenser and are exhausted by 402 STEAM POWER PLANT ENGINEERING the wet air pump G, from which they are forced through opening J to the hot well. The vacuum in chamber F will depend upon the vapor tension of the warm water in the bottom of the well, the amount of air carried along by the cooling water and steam, and the tightness of Fig. 203. Worthington Independent Jet Condenser. valves and joints. In case the water accumulates in the condenser cone F, either by reason of an increased supply or by a sluggishness or even stoppage of the pump, the condensing surface is reduced to a mini- mum, as soon as the level of the water reaches the spray pipe and the spray becomes submerged, and only a small annular surface of water is exposed to the exhaust steam. The vacuum is immediately broken, and the exhaust steam escapes by blowing through the injection pipe CONDENSERS 403 and through the valves of the pump and out the discharge pipe at J, forcing the water ahead of it; consequently flooding of the steam cylinder cannot occur. In starting up the condenser a partial vacuum for inducing a flow of injection water into the condenser chamber may be created by the pump if the suction lift is not too great. Many engineers, however, prefer to install a small forced injection or priming pipe the function of which is to condense sufficient steam to produce the necessary partial vacuum. Fig. 222 shows such an installation. AIR PUA Fig. 204. Section through a Blake Jet Condenser. Fig. 204 shows a section through the condensing chamber and air pump of a Blake vertical jet condenser with an automatic vacuum- breaking device. The injection water enters at opening marked " injection " and flows through the adjustable " spray " nozzle in a fine spray, at an angle of about 45 degrees, and impinges on the conical sides of the upper condenser chamber. The spray falls from the sides to the projecting ledges shown in the illustration. The ledges prevent the 404 STEAM POWER PLANT ENGINEERING spray from falling directly to the bottom of the chamber and insure an efficient mingling of steam and cooling water. A perforated copper plate is substituted for the shelves when the force of the injection water is not sufficient to produce spray. The circulating water and con- densed steam together with the non-condensable gases are drawn off at the bottom of the chamber. The vacuum-breaking device is shown at the right of the figure. When the rising water reaches the level of the float chamber, as in the case of an accidental stoppage of the air pumps, the float is raised and forces a check valve from its seat and allows an inrush of air to break the vacuum, thus preventing further suction of water into the condenser and consequent flooding of the engine. A is the forced injection or " priming " inlet used in start- ing up when the suction lift is considerable. Condenser Types and Applications: Power, June, 1906, p. 44; Engr. U.S., Jan., 1906, pp. 55-66. Jet Condensers (McBride): Trans. A.S.M.E., Vol. 12, p. 187; American Machinist, March 7, 1895, p. 185; Engr. U.S., Jan., 1906, p. 61; Whitham's Steam Engine Design, p. 294; Seaton's Manual of the Marine Engine, Chapter XI. 210. Condensing Water, Jet Condensers. — In a jet condenser the cooling water and exhaust steam mingle, and the degree of vacuum is a function of the final or discharge temperature; thus the quantity of cooling water required depends upon its initial temperature, the tem- perature of the discharge water, and the total heat in the steam entering the condenser. If the steam in the low-pressure cylinder at release is dry and saturated, the heat entering the condenser will correspond to the total heat in steam at release pressure, but it usually contains considerable moisture, part of which is reevaporated when the exhaust valve opens to the condenser; however, it is sufficiently accurate for all practical purposes to assume the exhaust steam entering the con- denser to be dry and saturated and its heat to correspond to the pres- sure in the condenser. Let A = total heat of steam at condenser pressure above 32 degrees. T 2 = temperature of the discharge water. T = initial temperature of the cooling water. W = weight of cooling water in pounds necessary to condense and cool one pound of steam to the required discharge temperature. Then W = X ~ r »t 32 ' ( 84 ) T 2 - T Example: How many pounds of cooling water are necessary to con- dense one pound of steam under the following conditions: Barometer 29.92; vacuum 26 inches; temperature of injection water 60 degrees F. CONDENSERS 405 The temperature of aqueous vapor corresponding to an absolute pressure of 29.92 — 26 = 3.92 inches of mercury is 125 degrees F. (See Table 49.) The discharge temperature, however, must be less than this, as the pressure in the condenser is due not only to the aqueous vapor but to that of the air carried over with the circulating water and the con- densed steam. In a condenser of this type the discharge temperature will be from 10 degrees to 15 degrees lower than that corresponding to the vacuum as recorded by the gauge. In this case assume it to be 15 degrees lower, i.e., T 2 = 125 — 15 = 110 degrees. The total heat corresponding to a pressure of 3.92 inches of mercury is 1120 B.T.U. above 32 degrees (see steam tables); T = 60 degrees; T 2 = 110 degrees. w= 1120-110 + 32 110-60 Evidently the higher the temperature of the discharge water the less will be the quantity of cooling water required, and consequently the smaller the weight of air introduced into the condenser; but the warmer the discharge water the greater will be the vapor tension and the lower the degree of vacuum. For reciprocating engines a hot-well tempera- ture between 110 degrees and 130 degrees F. is average practice; with turbines the temperature ranges between 80 degrees and 100 degrees F. On account of the inefficient heat absorption in practical installations, from 5 per cent to 15 per cent is added to the theoretical weight of cool- ing water as determined from equation (84). Table 50 has been calcu- lated from equation (84). Cooling Water for Condensers: Am. Mach., May 18, 1905, p. 656; Evaporation and Condensing Apparatus, Hausbrand, pp. 227, 240, 301, 318; Steam Power Plants, Meyer, p. 106. Wet-Air Pump, Jet Condensers. (See paragraph 285.) Circulating Pumps. (See paragraph 297.) 211. Effect of Aqueous Vapor upon the Degree of Vacuum. — The futility of attempting to better the vacuum by exhausting the vapor is best illustrated by a specific problem. Required the volume of aqueous vapor to be withdrawn per hour from a condenser operating under the following conditions, in order that the vacuum may be increased one pound per square inch : Tem- perature of discharge water 125 degrees; corresponding vapor tension 4 inches of mercury; barometer 30 inches; relative vacuum 26 inches; horse power, 100; steam consumption 20 pounds per horse-power hour; cooling water 25 pounds per pound of steam condensed. 100 X 20 X 25 = 50,000 pounds of cooling water per hour. = 833 pounds of cooling water per minute. TABLE 50. RATIO, BY WEIGHT, OF COOLING WATER TO STEAM CONDENSED (THEORETICAL). (Barometer 29.92.) Vacuum 24". Temperature of Steam 141°. Temp. of In- jection. Vacuum 25". Temperature of Steam 134°. Temp, of In- jection. Temperature of Hot Well. Temperature of Hot Well. 110 115 120 125 130 105 110 115 120 125 40 50 60 70 80 90 15.0 17.5 21.0 26.2 35.0 52.4 13.9 16.0 18.9 23.2 29.8 49.7 12.9 14.8 17.3 20.7 25.9 34.6 12.1 13.7 15.8 18.7 23.0 29.5 11.4 12.8 14.6 17.1 20.5 25.6 40 50 60 70 80 90 16.1 19.0 23.2 30.0 42.0 70.0 14.9 17.4 20.9 26.1 34.8 52.1 13.8 16.0 18.9 23.0 29.6 41.5 12.9 14.8 17.2 20.7 25.9 34.5 12.1 13.7 15.8 18.7 22 8 29.4 Vacuum 26 ,; '. Temperature of Steam 125°. Temp. of In- jection. Vacuum 27". Temperature of Steam 114°. Temp, of In- jection. Temperature of Hot Well. Temperature of Hot Well. 100 105 110 115 90 95 100 105 40 50 60 70 80 17.5 21.0 26.3 35.0 57.6 16.1 19.0 23.2 30.0 42.0 14.8 17.4 20.9 26.0 34.7 13.8 16.0 18.8 23.0 29.6 40 50 60 70 80 21.2 26.5 35.3 52.9 19.1 23.4 30.1 42.1 17.4 20.9 26.2 34.9 52.3 16.0 19.0 23.2 29.8 41.5 Temp. of In- jection. Vacuum 27.5". Temperature of Steam 108°. Temp, of In- jection. Vacuum 28". Temperature of Steam 100°. Temperature of Hot Well. Temperature of Hot Well. 80 85 90 95 75 80 85 90 40 50 60 70 26.6 35.6 52.3 23.6 30.3 42.5 70.8 21.1 26.4 35.2 52.8 19.1 23.4 30.0 42.0 40 50 60 70 30.5 42.7 71.2 26.6 35.5 53.2 23.5 30.2 42.3 70.6 21.1 26.3 35.1 52.7 Temp, of In- jection. Vacuum 28.5". Temperature of Steam 90 o Temp, of In- jection. Vacuum 29". Temperature of Steam 77 3 Temperature of Hot Well. Temperature of Hot Well. 60 65 70 75 55 60 65 67 35 40 45 50 42.2 54.0 72.0 35.8 43.0 53.5 72.0 30.6 35.6 42.8 53.5 29.2 33.4 38.8 46.6 35 40 45 50 52.0 69.3 43.0 54.0 71.5 35.8 43.0 54.0 72.0 33.4 38.4 47.0 61.0 CONDENSERS 407 Now to increase the vacuum one pound per square inch, approxi- mately 2 inches of mercury, the temperature of the water must be lowered to 102 degrees F., that is, 833 (125-102)= 19,159 B.T.U. 19 159 must be abstracted from the water in one minute, or — = 18.6 pounds of water to be evaporated per minute. (1030 = average heat of vaporization of water under 26 to 28 inches of vacuum.) Now, one pound of vapor at 102 to 125 degrees F. has an average volume of 270 cubic feet. Therefore 18.6 X 270 = 5022 cubic feet of vapor must be exhausted per minute to increase the vacuum from 26 to 28 inches, which is man- ifestly impracticable. 212. Injection Orifice. — The velocity of water entering a jet con- denser, neglecting friction, may be determined from the formula V = y/Tgh, (85) where V = velocity of the water in feet per second. g = acceleration of gravity = 32.2. h = total head in feet. If p = pressure below the atmosphere in pounds per square inch, h x = distance in feet between the source of supply and the injection orifice, then h = 2.3 p ± h v (86) and equation (85) may be written V= 8.025 V2.Sp± h v (87) If the supply is under pressure, h x is positive; if under suction, it is negative. Example: What is the theoretical velocity of water entering a con- denser with 26-inch vacuum (referred to 30-inch barometer); suction head 8 feet Here p = pressure in pounds per square inch, corresponding to 26 inches of mercury = 12.8 pounds per square inch. h t = 8. V = 8.025 V2.3 X 12.8-8 = 37.1 feet per second = 2226 feet per minute. In proportioning the injection orifice in practice the maximum velocity of flow is assumed to be between 1500 and 1800 feet per minute, 408 STEAM POWER PLANT ENGINEERING or, approximately, area of injection orifice in square inches = weight of injection water in pounds -r- 650 to 780. (" Manual of Marine Engineer- ing," Seaton, p. 204.) A rough rule gives area of orifice = area of low pressure piston in square inches -^ 250. (Seaton, p. 204.) 213. Volume of the Condenser Chamber. — According to Thurs- ton the volume of a jet condenser should be from one-fourth to one- half that of the low-pressure engine cylinder. ( " Steam Engine Manual," Thurston, II, 127.) According to Hutton the volume should not be less than that of the aif pump and should approximate three- fourths of that of the engine cylinder in communication with it. 214. Injection and Discharge Pipes. — In practice the diameter of the injec- tion pipe is based on a velocity of 400 to 600 feet per minute and that of the discharge pipe of 200 to 400 feet per minute; the lower figures for pipes under 8 inches in diameter, the upper range for larger diameters. (Atmospheric relief valves. — See paragraph 351.) 215. Siphon Condensers. — Fig. 205 shows a section through a Baragwa- nath siphon condenser, illustrating the principles of a parallel-current baro- metric condenser. The cooling water enters the side of the condenser cham- ber at A and passes downward in a thin annular sheet around the hollow cone D. The exhaust steam enters at B and is given a downward direction by the goose neck C. It flows through the nozzle D and is condensed within the hollow cone of moving water, the combined mass including the entrained air discharging through the contracted throat E at high velocity into the tail pipe F. The water column in the tail pipe must be enough to overcome the pressure of the atmosphere; i.e., it should be 34 feet or more above the surface of the hot well, otherwise water would rise within this pipe to a height corre- sponding to that of the barometer, which is approximately 34 feet for a barometric pressure of 30 inches of mercury. This is not strictly true when the condenser is in full operation, as the injector effect of the I ^ u I pi [TO— i Ij^ i Fig. 205. Baragwanath Siphon Condenser. CONDENSERS 409 moving mass is sufficient to overcome several pounds pressure, and the tail pipe may be less than 34 feet, but to provide against any possibility of the water being drawn into the cylinder of the engine the length is made greater than 34 feet. The spray cone D is adjustable and admits of close regulation of the water supply without changing the annular form of the stream. The condensing water may be supplied under pressure or under suction. For lifts not greater than 15 feet no supply pump is necessary, the water being raised by the siphon action of the condenser. This condenser requires the same amount of cooling water per pound of steam as the standard jet condenser, and is capable of maintaining a vacuum of from 24 to 25 inches. A vacuum of 28£ inches has been recorded for a condenser of this general type. (Trans. A.S.M.E., 26-388.) An atmospheric relief valve G is provided in case the vacuum fails from any cause, which will permit the steam to escape to the atmosphere. The above type of condenser is adapted to very muddy cooling water, since no filtration is necessary beyond the removal of such solid matter as may clog up the annular space H. In the Armour Glue Works at Chicago condensers of this type are successfully maintaining a 90 per cent vacuum with cooling water at 60 degrees F., and the circulating water is practically liquid mud. Siphon Condensers, Discussion: Trans. A.S.M.E., Vol. 26, p. 388. Siphon Con- densers: Electrical World, June, 1897, p. 818; Hutton, The Mechanical Engineering of Power Plants, p. 106; Engr. U.S., Jan., 1906. 216. Size of Siphon Condensers. — The size of siphon is indicated by the diameter of the engine exhaust pipe. Table 51 gives the sizes of barometric condensers as manufactured by prominent makers. TABLE 51. SIZE OF SIPHON CONDENSERS. Steam to be Condensed. Size Usually- Furnished, Inches. Steam to be Condensed. Size Usually Pounds per Hour. Pounds per Minute. Pounds per Hour. Pounds per Minute. Furnished, Inches. 2,000 3,000 4,000 5,000 6,000 33 50 66 83 100 5 7 8 9 9 8,000 10,000 15,000 20,000 133 166 250 333 10 12 14 14 Vacuum 26 inches; barometer 30 inches. 410 STEAM POWER PLANT ENGINEERING The diameter of the throat may be closely approximated by the empirical formula Diam. in inches = 0.0077 VWw, (88) in which W = weight of steam to be condensed per hour. w = weight of water required to condense one pound of steam. The maximum width of the annular opening for the admission of water may be obtained from the empirical formula Width in inches = Ww , (89) 39,550 d in which d = diameter of the nozzle or bottom of the cone in inches. IF and was in (88) 217. Ejector Condenser. — Fig. 206 shows a section through a Schutte exhaust steam "induction" condenser, illustrating the prin- ciples of the ejector condenser in which the EXHAUST Fig. 206. Schutte Ejector Condenser. Fig. 207. Piping for Schutte Ejector Condenser. momentum of flowing water ejects the discharge without the aid of the circulating pump. Exhaust steam enters the ejector through the open- ing marked " exhaust," passes through a series of inclined orifices and CONDENSERS 411 nozzles at considerable velocity, and, meeting the cooling water in the inner annular chamber, is condensed. The cooling water is drawn in continuously through the opening marked " water," by virtue of the vacuum formed, and sufficient velocity is imparted to the jet to dis- charge the combined mass of condensed steam, cooling water, and air against the pressure of the atmosphere. Adjustment for capacity is effected by raising or lowering the ram R by means of the wheel H. An adjustable sleeve controls the avail- able area of the exhaust inlet by covering more or less openings in the combining tube. When the cooling water is supplied under pressure the openings marked " steam " and are blanked. When water is taken under suction and water under pressure is available for starting, is blanked and opening marked "steam" is connected with the pres- sure supply. When water is taken under high suction and live steam is used for starting, inlet marked "steam" is connected to live steam and an over- flow check valve is placed at 0. Fig. 207 gives an outline of the necessary piping for a condenser installation of this type. These condensers are made in all sizes conforming with exhaust pipe diameters of 1J to 20 inches. The same amount of cooling water is required as for jet condensing and vacua of 20 to 25 inches are readily obtained. Exhaust Steam Induction Condensers : Power, Dec, 1898, p. 14. Ejector Condenser: Hutton, Mechanical Engineering of Power Plants, p. Ill; Eng. News, Oct. 5, 1905, p. 360. 218. Barometric Condensers.*— Fig. 208 shows a section through a Weiss counter-current condenser, illustrating the principles of a barometric jet con- denser. The cooling water enters the upper part of the condensing chamber A through pipe iV and falls in cascades, as shown in the figure, to tail pipe B, from which it flows by gravity to the hot well. The exhaust steam enters chamber A through pipe D, * The author has been informed that the word " Barometric " in connection with jet condensers is the registered trade mark of the Alberger Condenser Company. Fig. 208. Weiss Counter-Current Condenser. 412 STEAM POWER PLANT ENGINEERING and, coming in contact with the cold-water spray, is condensed. The air is exhausted from the top of the condenser by a dry vacuum pump through pipe F. In flowing to the pump the air passes upwards through the water spray and its temperature is lowered to that of the injection water, thereby reducing the volume to be exhausted. Any moisture passing over with the air is separated at G before reaching the air pump, and flows out through the small barometric tube H. The cooling water is forced to the condenser chamber through pipe N by any positive displacement pump, the actual head pumped against Fig. 209. Section Through Condensing Chamber, Alberger Barometric Condenser. being the difference between the total height and that of a column of water corresponding to the degree of vacuum in the condenser. The main barometric tube or tail pipe B through which the water is dis- charged is 34 feet or more in length and is provided with a foot valve C. The counter-current principle permits a much higher temperature of hot well for the same degree of vacuum than does the parallel current, a hot- well temperature of 120 degrees and a vacuum of 27 inches being readily maintained. A small pipe K connecting the main condenser CONDENSERS 413 with the small barometric tube H insures at all times a sufficient quantity of water in the small auxiliary hot well to seal the tube. The water from this auxiliary hot well flows over a weir, as indicated, into a counter-weighted bucket M, the latter having a hole in the bot- tom which allows the normal flow to escape. But in case a sudden heavy overload is thrown on the engines, and the adjustment is for a light load, the temperature of the discharge will reach the boiling point and an abnormal quantity of water will flow down the small barometric tube. This will cause the water to flow into the bucket much faster than the opening in the bottom can dispose of it; as a result the bucket will increase in weight and will open up a free-air valve L which reduces the vacuum two or three inches and raises the boiling point without " dropping " the vacuum entirely. E is the atmospheric relief valve. Fig. 209 shows a section through the condensing chamber of an Alberger barometric condenser. In principles of operation the con- denser is similar to the Weiss, but differs considerably in details. Exhaust steam enters at A and divides into two streams, one flowing directly to the inner chamber D, the other through the annular space E. Cooling water enters through B and is broken up into a fine spray by the serrated cone F, which is hung upon a long spring, thus auto- matically adjusting itself to the quantity of water entering the con- denser. After condensing the exhaust steam in the inner cylinder the partly heated spray of cooling water in falling is brought in contact with the exhaust steam which enters through the annular space. This process permits of a high hot-well temperature without affecting the degree of vacuum. The air which is not entrained by the cooling water and carried down the tail pipe collects under the spray cone F and ascends through the tubular support of the cone into the air cooler. This air cooler is simply a small chamber in which the non- condensable gases are cooled by a small portion of the circulating water before they are withdrawn by the air pump. The circulating water used for the purpose is forced into the cooling chamber through pipe K and falls through serrated openings in the bottom to the condenser proper. The air enters the chamber through these same openings, and is withdrawn by the air pump. Surround- ing the cooler is a separating space of large capacity to allow the subsidence of any entrained moisture before the air reaches the vacuum pump. Fig. 236 shows a typical installation of an Alberger condenser in connection with a cooling tower, and Fig. 226 that of a Weiss condenser in the Northwestern Elevated R. R. Power Station, Chicago. 414 STEAM POWER PLANT ENGINEERING Fig. 210 shows a section through the condensing chamber of a Worthington barometric condenser. The drawing is self-explicit. HAND WHEEL FROM.ENGINE Fig. 210. TO TAIL PIPE Section Through Condensing Chamber, Worthington Barometric Condenser. Fig. 211 shows a section through a Tomlinson barometric condenser. The air pump instead of discharging into the atmosphere is made to deliver into the tail pipe where the vacuum is still sufficient to support the column of water below the point of delivery. The effect produced is that of a two-stage air pump, the tail pipe becoming the second stage. Suitable by-pass valves enable the air pump to be discharged into the atmosphere or to be cut out entirely. (Power, February, 1907, p. 94.) Fig. 211a shows the application of a centrifugal pump to the tail pipe of a barometric condenser. This permits of a very short tail pipe, as the pump takes the place of the barometric column. Counter-Current Condensers : Am. Elecn., Feb., 1905, p. 81; Power, March, 1905, p. 182, Jan., 1906, p. 44; Engr. U.S., Jan., 1906, p. 58; Hausbrand, . Evaporating and Condensing Apparatus, Chapter XX; Bulletin No. 6, Heisler Mfg. Co., St. Marys, Ohio. The Barometric Condenser : Power, Jan., 1907, p. 1. CONDENSERS 415 INJECTION Fig. 211. Tomlinson Type B Barometric Condenser. 416 STEAM POWER PLANT ENGINEERING As previously outlined, surface condensers may be divided into three general classes, (a) water cooled, (b) air cooled, and (c) evaporative. 219. Water-Cooled Surface Condensers. — Water - cooled surface condensers are by far the most extensive in use and only occasionally are the con- ditions such as to warrant the installation of the other class. They are ordinarily classified as (1) single-flow, (2) double- flow, and (3) multi-flow. Fig. 212 shows a sectional elevation through a Baragwa- nath vertical condenser, illus- Fig. 211a. Centrifugal Pump Applied to the Tail Pipe of a Barometric Condenser. DISCHARGE STEAM trating the single- flow type. It con- sists essentially of a cast - iron shell provided with two heads, into which a number of one-inch exhaust brass tubes are ex- panded. Exhaust steam fills the shell and flows around and between the tubes, while the cooling water is caused to circulate through the tubes by means of a circulating pump. The steam is con- densed by contact with the tubes and drops to the bottom tube sheet, from which it is exhausted by the air pump. The circulating water flows through the tubes in one direction only, hence the name " single flow." To allow for the unequal expansion of shell and tubes TO AIR PUMP Fig. 212. Baragwanath Surface Condenser. CONDENSERS 417 418 STEAM POWER PLANT ENGINEERING the two halves of the shell are provided with slightly thinner plates flanged outward, the flanges being bolted together with a spacing ring between them. This joint gives to the shell, in the direction of its length, a certain amount of elasticity which is sufficient to allow for the greatest possible elongation of the tubes without straining the tube ends and causing leakage. Fig. 213 shows a section through a Wheeler admiralty surface con- denser mounted on a combined air and circulating pump, illustrating the typical " double-flow " surface condenser. The condenser proper consists of a ribbed cast-iron chamber of rectangular section fitted with a number of small seamless drawn brass tubes through which the cooling water is forced by suitable means. The exhaust steam enters at the top and is prevented from impinging directly against the tubes by baffle plates, which serve also to distribute the steam more evenly over the cooling surface. The steam in passing between the tubes is condensed, and falls to the bottom of the chamber, from which it is removed, together with the entrained air, by a vacuum pump. The water chamber between the tube sheet and the head is divided into two compartments, as shown in the illustration, the partition being so arranged that the water flows first through the lower set of tubes and then through the upper set in the opposite direction. Thus the tem- perature of the cooling water increases as it rises, and reaches a maximum where the exhaust steam enters. Condensation begins as soon as the vapor enters the condenser, and the surfaces of the tubes are at once covered with a thin film of water flowing downwards from tube to tube. Fig. 214. Surface Condenser, C. H. Wheeler & Co. Fig. 214 gives the details of a C. H» Wheeler & Co.'s high- vacuum surface condenser. The condensing chamber is of the series- parallel type in which the water enters the top group of tubes, then passes to the middle section and finally through the bottom section- CONDENSERS 419 Connecting chambers are provided at the ends of the shell as illustrated. This construction of water chamber keeps the condenser completely filled with cooling water at all times. The inlet is at the bottom but the water is carried up through the annular chamber to the top of the tubes. Fig. 215. Weighton Multi-Flow Surface Condenser. Fig. 215 shows a section through a multi-flow surface condenser designed by Professor R. L. Weighton. The condenser has three compartments separated by two diaphragms inclined to the hori- zontal. Each compartment is fitted with a number of brass tubes three-fourths inch in diameter by four feet in length, spaced one and one-eighth inches between centers. The cooling water circulates through the tubes five times, giving an effective length of 20 feet. The notable features of the condenser are abolishment of steam space, and compartment drainage of condensed steam. Mere passages of such shape 420 STEAM POWER PLANT ENGINEERING and section as will insure distribution of the steam over the entire surface are used instead of the large steam space usually associated with surface condensers. Each compartment is separately drained to the air pump, so that the surfaces in the lower compartments are unimpeded in their condensing action by the condensed steam from the upper compartments flowing over them. Referring to Fig. 215: Exhaust steam enters the condenser at A and flows toward the hot well H. The greater part of the steam is condensed in the first section of the condenser, and the condensation is drained directly to the hot well. The balance of the condensation takes place in the remaining sections, the condensed steam being withdrawn from each section. The wet-air pump withdraws the condensed steam and non-condensable gases through opening P. Cool- ing water enters at / and leaves at 0. An exhaustive series of tests on a condenser of this type credit it with a much higher efficiency than the ordinary single or double-flow apparatus. (See " The Efficiency of Surface Condensers," Proc. Institute of Naval Architects, March, 1906; also Engineer, London, April 27, 1906.) 220. Cooling Water, Surface Condensers. — The amount of cooling water required per pound of steam in a surface condenser is dependent upon the vacuum, the temperature of the condensed steam, and the range in temperature of the cooling water; it may be closely approxi- mated from the formula W- A ~ Tl + 32 . (90) where \ = total heat of the exhaust steam above 32 degrees F. T x = temperature of the condensed steam. T = temperature of the injection water. T 2 = temperature of the discharge water. W= pounds of injection water necessary to condense one pound of steam. Example : Required the quantity of cooling water necessary to con- dense one pound of steam under the following conditions: Initial tem- perature of the cooling water 60 degrees F.; final temperature 100 degrees F.; vacuum 26 inches, referred to 30-inch barometer. Here X = 1120 B.T.U., T = 60, T 2 = 100. = 1120-110 + 32 _ 100 - 60 That is, the ratio of cooling water to condensed steam is 26.0 to 1. In turbine practice where vacua as high as one-half pound absolute are CONDENSERS 421 obtained, the ratio of cooling water to condensed steam is nearly twice this quantity. For example, if a vacuum of 28.92 inches is desired with the barometer at 29.92 and the range of the circulating water tem- perature is 70 to 50 degrees and the temperature of the hot well 80 degrees, the ratio will be 1106-80 + 32 W*= 70-50 = 52.9. In determining the amount of cooling water it is well to bear in mind that in the ordinary condenser of the single or double-flow type 150 140 130 120 110 100 90 80 TO 60 50 40 A New Type with Cores and Spray and Dry Air Pump B New Type without Cores and Spray Ordinary Pump C Old Type, Ordinary Pump \A B^ +3 <^.C !3 ft 3 Pi \ S3 \ Relation between Hot-Well Temperature and Vacuum in Surface Condensers \ \ \ \ \ \ 23 24 25 .26 27 Vacuum Referred to 30 Inch Barometer Fig. 216. the temperature of the condensed steam will be from 10 to 20 degrees lower than that corresponding to the degree of vacuum in the con- denser, and that the temperature of the condensing water at the dis- charge point will be from 5 to 10 degrees lower than the temperature due to the vacuum. With well-designed condensers of the multi-flow type the temper- ature of the hot well may be from 3 to 5 degrees higher than the tem- perature due to the vacuum, and the temperature of the condensing water at the discharge point may be equal to or slightly higher than that due to the vacuum. (Proc. Inst, of Naval Arch., March, 1906.) (See Fig. 216.) 221. Extent of Water-Cooling Surface. — Theoretically the opera- tion of a surface condenser is divided into two periods, (1) the period 422 STEAM POWER PLANT ENGINEERING of condensation during which the heat of vaporization at the observed pressure is removed and (2) the period of cooling during which the temperature of the condensed steam is reduced. In order to determine accurately the extent of cooling surface it would be necessary to cal- culate the heat transmission for each of the two periods. In practice, however, it is assumed that condensation and cooling take place simul- taneously, and that the mean temperature difference is a direct function of the temperature corresponding to the exhaust steam in the con- denser and that of the condensed steam and cooling water. The error in these assumptions has only a slight influence on the estimation of the cooling surface and is entirely lost sight of in the liberal factor allowed in practice. Let & = cooling surface in square feet. \ = total heat above 32 degrees of the exhaust steam at con- denser pressure. T = initial temperature of the circulating water. T 2 = final temperature of the circulating water. T x = final temperature of the condensed steam. Ti = temperature of the exhaust steam at condenser pressure. U = coefficient of heat transmission, B.T.U. per hour, per degree difference in temperature, per square foot of cooling sur- face. d = mean difference in temperature between T s and T 2> and T . W = weight of condensed steam per hour. d = T *~ T ° "(see equation (118), Chap. XII); and since the heat absorbed by the cooling water is equal to the heat given up by the steam, SUd = W{X-(T 1 -32)}. (91) s =W (X-T 1+ 32) . Ud Whitham (" Steam-Engine Design,' ' p. 283) uses the arithmetic mean T + T d = T s °— - — 2 instead of the mean as determined from (118). Equation 118 is based on the assumption that the fluid on each side of the tube is homogeneous, which is far from being true in the case of the air-steam mixture in a condenser, and for this reason many designers prefer to use the simpler arithmetic formula. The coefficient of heat transfer, U, as used in above equations refers CONDENSERS 423 tc the mean or average values for the entire surface since the actual heat transmission varies widely for different parts of the condenser; thus the actual value of U varies from over 1000 in the first few rows of the tubes (where the steam comes directly into contact with the cooling surface) to less than 50 in the bottom row (where the tubes are practi- Fig. 216a. Application of Weighton Dry-tube Surface Condenser to Vertical Marine Engine. cally submerged in water of condensation) and to 3 or less for the tubes surrounded only by air. Prof. Josse of the Royal Technical School, Charlottenburg, after an exhaustive investigation of the subject found that the actual value of U varied with (1) The material, thickness, shape and cleanliness of the tubes. (2) The velocity of the water through the tubes. 424 STEAM POWER PLANT ENGINEERING (3) The velocity of the steam against the tubes. (4) The percentage of air in the steam surrounding the tubes. (5) The extent of submersion of the steam side of the tubes. Some of the results of his investigations are shown in Figs. 216, b, c, and d. See also Power and ' ~ Engr., Feb. 2, 1909. The effect of thickness, ma- terial, etc., of condenser tubes is so small in the ultimate result and the choice and arrangement are so largely determined by practical consideration that they may be neglected. The value of U increases ap- proximately as the square root of the velocity of the water flow- ing within the tube, so that in- crease in water velocity effects a substantial increase in the heat transmission ; but the resistance encountered by the circulating water increases as the square of the velocity, and the power con- sumed in pumping the water in- creases as the third power of the velocity, so that a point is soon reached where the gain on the one hand may be offset by the loss on the other. A study of a number of instal- lations gave 12 3 4 5 6 7 Rate of Flow of Cooling Water- Feet per Second Fig. 216b. Old-style surface condenser, V = 30 to 240 ft. per min., average 90. Modern dry tube surface condenser, V = 120 to 360 ft. per min., average 240. From the curves in Figs. 126b and 126c it will be seen that air is an excellent heat-insulating material; hence, the greater the amount of air entrained with the steam the lower will be the coefficient of heat transmission. The necessity of removing the air as fast as it accumu- lates is at once apparent. In the older types of surface condensers the water of condensation from the upper rows of tubes is permitted to fall on the rows immedi- CONDENSERS 425 Heat Transference for Air 10 20 30 40 50 60 70 Mean Rate of Flow of Air —Feet per Second Fig. 216c. ately below, the water increasing in volume as it passes the successive banks of tubes until it completely envelops them. The coefficient U varies from 1000 or more in the upper row to less than 50 in the lower, giving a mean value of approximately 250 to 350 for the entire surface. In estimating the extent of cooling surface for a condenser of this type an average figure for plain brass tubes with water velocities of 50 to 100 feet per minute is U = 250. For a velocity of 100 to 240 feet per minute U may be taken 50 per cent greater than these figures. When the tubes are clean a much higher value may be taken, but a liberal factor is usually allowed for possible variation in the con- dition of operation. In the modern dry-tube surface condenser, designed along the lines of the one described in paragraph 221, in which the water of condensa- Heat Transference for Air tion is withdrawn as rapidly as it is formed, mean values of U = 800 to 900 are not unusual. In estimating the extent of cooling surface for condensers of this type an average value of U is 600 with water veloci- ties of 4 to 5 feet per second. Example: Standard Type of Surface Condenser : — Required the number of square feet of cooling surface per I.H.P. neces- sary to condense the steam from an engine operating under the following conditions : Engine uses 20 pounds of steam per I.H.P. -hour, vacuum 26 inches with barometer at 30; temperature of cooling water at 60 degrees. Here X = 1115 and T s = 126 (from steam tables), T = 60, T 1 = T s - 10 = 116. w Q go t K £4 K a* A ,> <* e o) V f^ *f 5.25 **• perSecN 25 20 15 10 5 Air Pressure "-Inches of Mercury Fig. 216d. 426 STEAM POWER PLANT ENGINEERING In this type of condenser average practice gives a temperature difference of approximately 10 degrees between the temperature of the hot well and that corresponding to the degree of vacuum. T 2 = T s - 15 - 101. Any value may be fixed upon for T 2 greater than T and less than T s . The nearer T 2 is to T the greater must be the quantity of circulating water per unit of time for a given rate of condensation. On the other hand, the nearer T 2 is to T s the less is the mean temperature difference d and hence the greater must be the cooling surface for a given coeffi- cient of heat transmission. When water is cheap and the head pumped against is small T 2 should be given a lower value than when water is costly and the discharge head is large. Average engine practice, with conditions as stated, gives T 2 a value of approximately 15 degrees less than that corresponding to the degree of vacuum. The logarithmic mean is, equation (118), 101 - 60 *- 126 - 16 = 42A ge 126 - 101 The arithmetic mean gives , ™ 60 + 101 ._ a = 126 » = 45.5. Substitute the value of d in equation (92) and assume U = 250, the figure commonly used for this type of condenser. _ _ 20(1115- 116 + 32) _ * " 250 X 42.4 ~~ iy4 ' or say two square feet per I.H.P. of engine. Surface condensers of this type are ordinarily rated on a basis of two square feet per I.H.P. Example: Dry-tube Multi-flow Surface Condenser: — Required the number of square feet of cooling surface per kilowatt necessary to condense the steam from a steam turbine operating under the following conditions: Turbine uses 15 pounds of steam per kw.-hour; vacuum 28.5 inches, referred to 30-inch barometer; temperature of cooling water 70 degrees. Here i = .9X 1100 = 990. The total heat of dry steam corresponding to an absolute pressure of 1.5 inches is 1100, but in the case of high vacuum turbine practice the steam entering the condenser is far from being dry, the quality varying from 0.80 to 0.95, depending upon the quality of the steam at admission. An average figure is 0.9. T s = 92, T = 70, T t = T s - 4 = 88. CONDENSERS 427 In this type of condenser the hot -well temperature varies from T x = T s to T x = T s - 8. T ± s In this type 7\ varies from T 2 = T s = 87. to 7\, = 10. 87 - 70 log 90 - 70 92 - 87 11.5. = 2.02, 70 -I- 87 Arithmetic mean gives d = 92 ■= — = 13.5. Substitute the value of d in equation (92) and assume U = 600, the figure commonly used for this type of condenser. _ 15 (990 - 88 + 32) 600X11.5 or say 2 square feet per kw. of generator. There is no standard rating of surface condenser for steam turbine work because of the wide variation in operating conditions. A study of a number of modern installations gives 1 to 2.5 sq. ft. per kw. for large turbo-generators using dry tube surface condensers. 2.5 to 4 sq. ft. per kw. for small turbo-generators using standard surface condensers. Professor Weighton found from his experiments that a surface con- denser constructed on the lines of the one described in paragraph 221 in conjunction with dry-air pumps, was capable of condensing 20 pounds of steam per square foot of surface per hour and maintained a vacuum of 28^ inches (referred to a 30-inch barometer), and this with a cooling- water consumption of 24 pounds per pound of condensed steam; with an inlet temperature of 50 degrees F. a condensation of 35 pounds of steam per hour per square foot of cooling surface was effected at a ratio of 28 pounds of cooling water per pound of steam, the vacuum remaining 28^ inches. See Fig. 216. (Engineering Record, May 19, 1906, p. 615.) EXAMPLES OF MODERN CONDENSER PROPORTIONS. Name of Station. Commonwealth Edison Co.: Northwest Station Quarry Street Fisk Street *59th St., Interborough, N. Y Metropolitan St. Ry., Kansas City Size of Turbo- Generators. 20,000 14,000 12.000 15.000 10.000 Sq. Ft. of Condenser Surface. 32,000 25,000 25,000 25.000 22,000 Sq. Ft. of Surface per Kw. 1.60 1.79 2.08 1.67 2.20 * Combined Engine and Low-pressure Turbine. Surface Condenser Air Pumps. — See paragraphs 284-291. 428 STEAM POWER PLANT ENGINEERING TABLE 52. SQUARE FEET OF COOLING SURFACE NECESSARY TO CONDENSE AND COOL ONE POUND OF STEAM PER MINUTE. (Barometer 29.92.) Temp. Vacuum 24". Temperature of Steam 141°. Temp, of In- jection Water. Vacuum 25". Temperature of Steam 134 of In- jection Water. Temperature of Hot Well. Temperature of Hot Well. 110 115 120 125 130 105 110 115 120 125 40 50 60 70 80 90 3.1 3.3 3.7 3.9 4.1 4.4 3.3 3.5 3.8 4.1 4.4 4.8 3.5 3.7 4.0 4.3 4.7 5.1 3.7 4.0 4.2 4.6 5.0 5.5 4.0 4.3 4.6 5.0 5.5 6.0 40 50 60 70 80 90 3.3 3.5 3.8 4.1 4.5 5.1 3.5 3.7 4.0 4.4 4.8 5.4 3.7 4.0 4.3 4.7 5.2 5.8 4.0 4.3 4.6 5.1 5.6 6.2 4.4 4.7 5.1 5.5 6.1 6.9 Temp. Vacuum 26". Temperature of Steam 125°. Temp. of In- jection Water. Vacuum 27". Temperature of Steam 114°. of In- jection Water. Temperature of Hot Well. Temperature of Hot Well. 100 105 110 115 90 95 100 105 40 50 60 70 80 3.6 3.8 4.2 4.6 5.1 3.9 4.2 4.6 5.1 5.6 4.2 4.6 5.0 5.4 6.1 4.6 5.0 5.4 6.0 6.7 40 50 60 70 80 4.1 4.4 4.8 5.4 4.4 4.7 5.2 5.8 6.7 4.7 5.1 5.G 6.3 7.2 5.1 5.6 6.2 7.0 8.0 Temp. Vacuum 28". Temperature of Steam 100°. Temp. of In- jection Water. Vacuum 29". Temperature of Steam 77 of In- jection Water. Temperature of Hot Well. Temperature of Hot Well. 75 80 85 90 60 65 70 40 4.6 5.0 5.5 4.9 5.4 6.1 5.3 5.8 6.6 7.7 5.8 6.4 7.3 8.5 35 40 45 50 6.3 6.8 7.4 6.9 7.4 8.0 8.9 7.7 8.3 9.0 9.9 50 60 70 Final temperature of injection water assumed to be 10 degrees lower than that of the hot-well. 222. Dry-Air Surface Condensers (Forced Circulation). — Where water is very scarce and the feed supply is reclaimed by condensing the exhaust steam, water-cooled condensers may be prohibitive in cost of operation, even when combined with cooling tower or other water-cool- CONDENSERS 429 ing device, since the latter involves a loss of water approximately equivalent to the amount of steam condensed, due to evaporation. Under these conditions air cooling has been successfully adopted. In the city of Kalgoorlie, West Australia, an electric station of 2000- horse-power capacity is equipped with air-cooled surface condensers. The condensers have been in use five years (1906), and have given excellent service with very little expense and maintenance. The con- denser consists of a large number of narrow chambers constructed of thin corrugated sheet-steel plates spaced J inch between centers. Each chamber has 1345 square inches of cooling surface. Fifty-one of these chambers are grouped into a compartment and 15 compartments constitute a section. Each section is equipped with three motor-driven fans 7 feet in diameter and running normally at 320 r.p.m. In all there are six sections, giving a total cooling surface of 45,000 square feet. The steam consumption of the main engines is 16 to 16.5 pounds per I.H.P. hour at rated load. At full load the fans require 130 kilo- watts, or approximately 10 per cent of the station output. The average vacuum obtained is about 18 inches throughout the year and ranges from inches on very hot days to 22 inches in cooler weather. The following figures, based on actual observation, show the effect of tem- perature of the external air on the vacuum when condensing 32,000 pounds of steam per hour (the rated capacity of the condenser) . Temperature Ex- Vacuum, Inches Temperature Ex- Vacuum, Inches ternal Air, (referred to 30-Inch ternal Air, (referred to 30-Inch Degrees F. Barometer) . Degrees F. Barometer) . 42.8 22 96.8 9.6 50 21.2 100.4 7.6 60.8 20 107.6 3.6 68 18.4 113 78.8 16 Air-Cooled Surface Condensers : Engineering News, Oct., 1902, p. 271 ; ibid., Vol. 49, p. 203. 223. Quantity of Air for Cooling (Dry-Air Condenser). — The volume of air, under atmospheric conditions, necessary to condense steam to any given temperature may be determined as follows: Let X = total heat, above 32 degrees F. of the steam at condenser pressure. T s = temperature of the vapor in the condenser. T x = temperature of the condensed steam. t = temperature of the air entering condenser. t x — temperature of the air leaving condenser. V = volume of air in cubic feet necessary to condense and cool one pound of steam. 480 STEAM POWER PLANT ENGINEERING B = specific weight of air under atmospheric conditions. C = specific heat of air under atmospheric conditions. d = mean temperature difference between the air and steam. S = cooling surface in square feet. U = coefficient of heat transmission, B.T.U. per square foot per degree difference in temperature per hour. Since the heat absorbed by the air must be equal to the heat given up by the steam, neglecting radiation we have VBC (t.-t) = \-T\ + 32, (93) from which V = ^,+32 . (94) For practical purposes C may be taken as the specific heat of dry air, the error due to this assumption being negligible even if the air is saturated with moisture. Example : How many cubic feet of air are necessary to condense and cool one pound of steam under the following conditions : Vacuum 20 inches; temperature of entering air, leaving air, and condensed steam, 60, 110, and 140 degrees F. respectively? Here A = 1131 (from steam tables). t x = 110, T 1 = 140, t = 60, C = 0.2377, B = 0.075. Substituting these values in equation (94), jr 1131-140 + 32 11W1 u- * + * 1 = 0.075X0.2377(110-60) = 115 ° ° UblC feet ° f air neCGSSary t0 condense one pound of steam under the given conditions. The proper area of cooling surface depends upon the value of the coefficient of heat transmission, which varies with the velocity and humidity of the air and character of the cooling surface. Accurate data are not available on this point. A few experiments made at the Armour Institute of Technology gave values of U = 10 to 25 B.T.U. per hour, per square foot, per degree difference in temperature for air velocities of 500 to 4000 feet per minute for corrugated steel sheeting J inch thick. Hence, sub- stituting in equations (94) and (92) we get, for the above example,. S = 1.5 square feet of cooling surface per pound of steam condensed per hour for air velocity of 4000 feet per minute, and S = 3.7 square feet for a velocity of 500 feet per minute. 224. Saturated-Air Surface Condensers (Natural Draft). — Fig. 217 shows vertical and horizontal sections of a Pennel saturated-air surface CONDENSERS 431 condenser. The apparatus consists of an upright cylindrical shell containing a number of vertical 4-inch steel tubes through which air is drawn by natural draft. A centrifugal pump circulates about one- half gallon of water per horse power per minute from a cistern below the condenser. The water flowing over the upper tube sheet and then descending the tubes by gravity forms a film over their entire interior surface. Horizontal Section. Section on AB. Fig. 217. Pennel Saturated- Air Surface Condenser. The condensing action is as follows: The current of exhaust steam entering the side of the shell at A is caused by suitable baffle plates to circulate among the tubes, and in condensing gives up its latent heat to the water film, which wholly or partially evaporates, saturating the ascending current of air at its own temperature. The upward current of hot vapor-laden air carries off the heat into the atmosphere. The cooling water which is not evaporated and lost to the atmosphere falls into the cistern below to be again taken up by the circulating pump, the water level in the cistern being kept constant by a float governing a valve on the supply pipe. The non-condensable gases collect at C, where they are removed by the dry -air pump, while the condensed steam is drawn off from the bottom tube sheet by the vacuum pump and discharged into the hot well. An excellent feature of this device is that the film of water on the cooling surface is secured without inter- ference with the ascending air currents and also without the use of sprays through small orifices likely to become clogged with rust or sediment. Where the recovery of the condensed steam is essential and a high vacuum of secondary importance, condensers of this type have proved to be good investments on account of the low first cost. 432 STEAM POWER PLANT ENGINEERING Table 53 gives the results of a test of a condenser of this type, taking steam from a 30 x 58 x 48 engine running at 45 r.p.m. (Power, December, 1903, p. 672; West Elect, May 19, 1900, p. 323.) TABLE 53. TEST OF PENNEL SATURATED-AIR SURFACE CONDENSER. Duration of trial .9 hours Average steam pressure at engine by gauge 139.8 pounds Average vacuum, mercury column 17.5 inches Average temperature in condenser 123.7 degrees F. Average temperature of circulating water 116.4 do Average temperature of city water 52 do Average temperature of outside air . . 62 do Average temperature of saturated air 106 do Average draft in stack of condenser 1.1 inches Average humidity of outside air 67 per cent Average amount of steam condensed per hour 7950 pounds Average amount of circulating water used per hour 114,660 pounds Average amount of city water used per hour 3462 pounds Pounds of city water per pound of steam 2.3 Pounds of circulating water per pound of steam 14.4 Average horse power of engine 569.7 Steam, pounds per I.H.P. per hour 13.95 Horse power required to run air pumps 10 . 5 Horse power required to run circulating pumps 3.0 Condensing surface, square feet 3900 Pounds of steam condensed per square foot surface per hour 2038 Barometer 28.58 inches Vapor tension corresponding to 123.7 degrees 3.82 inches Per cent of main engine steam used by auxiliaries 2.38 218 illustrates the Pennel a b Fig. " flask " type of atmospheric condenser. The exhaust steam enters below and follows the zig- zag course bounded by the inter- nal stay channels, condensing as it goes and driving before it the non-condensable gases to the out- let at the top. The condensed steam gravitates to the bottom and thence to the hot well. The I 4^4" V v> top of the flask is trough shaped " and causes the cooling water to Fig. 218. Pennel Flask Type of Saturated- fl()W down the gides of the flagk Air Surface Condenser. . in a thin stream. The portion of the cooling water not evaporated collects at the bottom of the flask and flows to the cooling-water reservoir. 3E CONDENSERS 433 225. Evaporative Surface Condensers. — An evaporative surface con- denser consists of a number of copper, brass, wrought- or cast-iron tubes arranged horizontally or vertically and connected to manifolds or chambers at each end. The exhaust steam passes through the tubes and a thin film of water is allowed to flow over the external surfaces. The cooling effect is brought about by the evaporation of part of the circulating water, and the general principle of operation is the same as that of the saturated-air condenser described above. Evaporation is sometimes hastened by constructing a flue over the tubes, thereby creating a natural draft, or by means of fans. With horizontal cast-iron tubes and natural draft, vacua from 23 to 27 inches are readily maintained with a cooling surface of approximately eight- tenths square foot per pound of steam condensed per hour. With vertical brass tubes and fan draft 8 pounds of steam per hour per square foot of cooling surface is not an unusual figure. The amount of cooling water evaporated per pound of steam varies from eight- tenths to one pound, depending upon the draft. The power necessary to operate the pumps and fans varies from 1 to 4 per cent of the total output of the plant. For an interesting discussion of evaporative condensers the reader is referred to the admirable article by Oldham in the Proceedings of the Institute of Mechanical Engineers, 1899, and reproduced as a serial in Engineering (London), April 28 to June 30, 1899. The following test of a vertical cast iron tube evaporative surface condenser (Table 54) will give some idea of the performance of this type of condenser. This condenser consisted of two rows of 4-inch vertical cast-iron pipes connected at the top by U bends and at the bottom by cast-iron manifolds. A perforated iron trough dis- tributes the water over the center of the bend and causes it to flow in a thin stream over the surface of the tubes. A wet-air pump is used for withdrawing the condensed steam and air. No fan is used for hastening evaporation.* Evaporative Condensers: Engr.,Lond.,May 5, 1899, pp. 432, 442, 447 ; Engineering, May 19, 1899, p. 661, June 2, 1899, p. 721, June 30, 1899, p. 861; Trans. A.S.M.E. 14-696; Power, Sept., 1904, p. 542; Prac. Engr. U.S., June, 1910, p. 346. 226. Location and Arrangement of Condensers. — In the modern power house one sees two general arrangements of condensers and auxiliaries: 1. The independent or subdivided system/in which each engine or turbine is provided with its own condenser, air and circulating pumps. 2. The central system, in which the condensers and auxiliaries are grouped together. Ordinarily one condenser suffices for all engines. * See end of paragraph 236 for evaporated surface condenser calculations. 434 STEAM POWER PLANT ENGINEERING TABLE 54. TEST OF A CAST-IRON, VERTICAL-TUBE, EVAPORATIVE SURFACE CONDENSER, NATURAL DRAFT. Date Weather Barometer Temperature of air Cooling surface, external Duration of trial, minutes Weight of steam condensed, pounds Boiler pressure Weight of water in circulation Weight of fresh water added Vacuum in condenser Initial temperature of circulating water Final temperature of circulating water Temperature of " make up " water Temperature of water in hot well Weight of steam condensed per hour, pounds . . . Weight of water circulated per hour, pounds Weight of " make-up " water added per hour. . . Weight of steam condensed per square foot of cooling surface per hour Weight of "make-up" water per pound of steam condensed, pounds Sept. 12 Sept. 13 Wet Fine 29.8 29.5 ? 60 272 272 99 115 800 800 60 60 1830 1830 600 640 23.36 24.1 117.5 113.9 128.4 125 58 58 136.5 131.8 485 427 6786 ? 364 334 1.8 1.54 0.75 0.80 226a. The Independent System. — The condenser is usually placed close to and below the engine so that all condensation may gravitate into it. Figs. 219 and 221 show an application of this system with jet condensers. Here each condenser receives its supply of cooling water from a main injection pipe and discharges into a main overflow pipe. The exhaust pipe leading to the condenser is by-passed through a suitable atmospheric relief valve to a main free exhaust header so that the engine may operate non-condensing in case the vacuum breaks or the condenser is cut out. The chief feature of this arrangement is its flexibility, as each unit is complete in itself and independent of the others. By far the greater number of central stations are equipped with independent condensers. Occasionally a jet condenser is located on the same level with the engine or even above it, Fig. 222, but such a location should be avoided if possible, as it usually necessitates a larger number of bends and joints in the exhaust pipes than the basement arrangement, and increases the possibility of air leakage. If the exhaust pipe does not drain directly into the condenser, the lowest point in the piping should always be provided with a drip which should be opened when the engine is shut down, as condensation and leakage are apt to fill the pipe with water if the engine stands for any length of time. The end CONDENSERS 435 DISCHARGE -*- ATMOSPHERIC RELIEF VALVE SJijigSjUttM^ ' Fig. 219. Jet Condenser located below Engine-Room Floor. Fig. 220. Surface Condenser located below Engine-Room Floor. 436 Fig. 221. Surface Condenser, Installed in the Suction Line of a Pumping Engine. n — i r^-r Fig. 222. Jet Condenser located above Engine-Room Floor. Fig 223 Typical Arrangement, Westinghouse-Leblanc Condenser and Curtis Turbine -ALL FOREIGN SOB STAN CES SUCH AS LEAVES, «&TIC«S STRAW.ETC. TSfcSJTSfe CONDENSERS 487 & 6 to a I 438 STEAM POWER PLANT ENGINEERING of the drip should be connected so that water cannot be drawn back through the drip pipe and into the engine cylinder. The length of exhaust pipe and particularly the number of bends between engine and condenser should be kept as small as possible, otherwise the engine may not derive the full benefit of the vacuum in the condenser. A case is recorded where the exhaust piping and appurtenances in con- nection with a 5000-horse-power engine caused a drop of several inches in vacuum between condenser and exhaust opening of the low-pressure cylinder. {National Engineer, December, 1906, p. 10.) The wet-air pump must always be located below the condenser chamber so that the condensation may gravitate to it. Fig. 225. Plan of Piping for Engine and Condenser, Des Moines City Ry. Co. Fig. 252 shows a surface condenser installation in connection with a vacuum or primary heater. Fig. 236 shows an application of a barometric condenser to a vertical engine installation. Fig. 220 shows the arrangement of a surface condenser with com- bined air and circulating pump in connection with a horizontal cross CONDENSERS 439 compound engine. The condenser and appurtenances are placed below the engine, thereby permitting the condenser to be closely connected to the engine. Fig. 221 shows the arrangement of a surface condenser in connec- tion with a pumping engine. The condenser is placed in series with the pump suction. 227. Central Systems. — In the central condensing systems the con- denser is located at any convenient point and the exhaust from all the engines piped to it. Any arrangement of condenser and auxiliary machinery may be adopted which will favor the lowest cost of installa- tion and expense of operation. Except where continuity of opera- tion is absolutely essential, only one circulating pump and one air pump are installed. This reduces the number of auxiliary pumps and appliances to a minimum, with a consequent decrease in first cost and maintenance. With properly designed exhaust piping the condenser may be located at a considerable distance from the engine without undue loss of vacuum. At the Cambria Steel Works, Johnstown, Pa., the maximum drop between condenser and engine is only three-quarters of an inch and the distance between them is about 1000 feet. Central condensers have found great favor in power plants in which the individual units are subjected to extreme variations in load, as in rolling mills. At the works of the Illinois Steel Company, South Chicago, 111., one condenser takes care of the steam from 15,000 horse power of engines in the rail mill, and another condenses the steam from the 15,000 horse power of engines in the Bessemer steel mill. A notable installation of this system in connection with street-railway work is in the power house of the Northwestern Elevated Company, Chicago, where a single condenser takes care of the exhaust steam of five engines, 11,000 horse power in all. Fig. 226 shows the general arrangement of this installation. For a comparison of the advantages and disadvantages of the inde- pendent and central systems see Engineering Magazine, October, 1900, p. 56, Engineering, London, June 23, 1899, p. 615, and Engineering, July 17, 1903. Centralization of Steam-Condensing Plant : Eng. Mag., Oct., 1900, p. 56 ; Iron Age, Jan. 7, 1904; Revue Technique, Feb. 25, 1903. Five Thousand H.P. Surface-Condensing Plant : Engr., Lond., May 23, 1903. Aurora & Elgin R.R. Condenser Plant : Engr. Rec, Vol. 47, p. 153. Condensing Apparatus of Manhattan Elevated Power Plant: Power, Aug., 1903, p. 411. Interborough R.R. Condenser Plant : St. Ry. Jour., Oct. 8, 1904. New York Rapid Transit Condenser Plant : Power, June, 1903, p. 283. Worthington Surface Condensers for Metropolitan Power Station : Power, June, 1901, p. 15. 440 STEAM POWER PLANT ENGINEERING ggsggga HMdfl j IS3& B ■*i i cd * =1 =dc ==f -*= CONDENSERS 441 228. High-vacuum Systems. — The average reciprocating engine gives its best commercial economy at a vacuum of approximately 26 inches (referred to a 30" barometer), and the ordinary standard jet or surface condenser has been designed to meet this requirement. At the time of the introduction of the steam turbine it was discovered that a very high vacuum would improve turbine economies to an extent hitherto impossible when applied to reciprocating engines. This con- dition naturally created an era of development among the condenser Fig. 226a. Condenser Installation, Quincy Point Power Plant of the Old Colony Street Railway Company. designers. It became evident at once that the old types that were capable of creating a 26" or 27" vacuum would require considerable modification to maintain a vacuum of 28" or 29". The principal improvement adopted by practically all manufacturers has been to apply a separate dry vacuum pump for the removal of air and non- condensable vapors. Surface Condensers. — Fig. 227 shows the arrangement advocated by the H. R. Worthington Company. The equipment comprises a surface condenser, a steam-driven centrifugal pump for circulating the cooling water, a steam-driven rotative dry-air pump, and a motor- 442 STEAM POWER PLANT ENGINEERING CONDENSERS 443 driven centrifugal hot-well pump. The surface condenser is piped direct to the turbine exhaust, only a corrugated copper expansion joint and a tee intervening. A tubular water vapor cooler, which is in reality a small surface condenser, is inserted in the circulating water line between the pump suction and condenser, and serves to arrest all the condensable vapor and thus reduces the volume to be handled by the air pump. All condensation, including that from the air cooler, collects in the hot well, from which it is pumped by a motor-driven circulating pump direct to heater or boiler. Cooling water is handled by a cen- trifugal pump having both suction and delivery pipes water-sealed, so that the work done by the pump is virtually that of overcoming the Exhaust Steam Inlet Water Outlet "Fig. 228. High- Vacuum System, C. H. Wheeler Co. fluid friction in the condenser and piping. All valves and stuffing boxes are water-sealed to prevent any possible leakage of air, and the condenser pump cylinder is especially designed to avoid vapor binding. This makes it possible to maintain a vacuum of one-half pound absolute with cooling water at 60 degrees F. In the high-vacuum condenser installation of the Chicago Edison Company the hot-well pump, dry-air pump, and the circulating pump are direct connected to a single- cylinder Corliss engine. Fig. 228 shows the general arrangement of the C. H. Wheeler Com- pany's high-vacuum condensing outfit. The condensing chamber is shown in section in Fig. 214 and is described in paragraph 219. The wet-air pump is illustrated in Fig. 287 and is described in paragraph 288. No dry-air pump is needed, and the makers, guarantee a vacuum within 4U STEAM POWER PLANT ENGINEERING two inches of absolute under full-load conditions of operating steam turbines. Fig. 229 shows a section through a Parsons " vacuum augmenter v for increasing the vacuum in a surface condenser. A pipe is led from the bottom of the main condenser to an auxiliary or augmenter having about one-twentieth of the cooling surface of the main condenser. At the point indicated a small steam jet is provided which acts as an ejector and draws out the air and vapor from the condenser and delivers it to the air pump. The water seal prevents the air and vapor from returning to the condenser. With this arrangement, according to tests conducted by Mr. Parsons, if there is a vacuum of 21\ or 28 inches in the condenser, there may be only 26 at the air pump, which, therefore, may be of small Fig. 229. Parsons Vacuum Augmenter. size, the jet compressing the air and vapor from the condenser to about one-half of its original volume. The steam jet uses about one and one- half per cent of the steam used by the turbine at full load. Jet Condensers. — Fig. 229a gives the general details of a Westing- house-Leblanc multi-jet condenser which, under commercial conditions, has realized vacua within 99 per cent of the ideal. The most striking feature of this system lies in its compactness and simplicity, a 1500- kw equipment being less than 9 feet in height. Referring to Fig. 229a, exhaust steam enters the condenser chamber at the upper left- hand opening and meets the cooling water as it is forced through spray nozzle C. The condensed steam and injection water fall to the bottom of the condenser and are removed by centrifugal pump M. The non- condensable vapors are withdrawn by valveless rotary air pumps P, through suction opening 0. Referring to section N-N through the CONDENSERS 445 air pump it will be seen that this pump consists primarily of a reverse Pelton turbine wheel in conjunction with an ejector. Sealing water is introduced through the branch indicated by dotted outline, into the central chamber G, from which it passes through port H. It is then caught up by the blades P of the Pelton wheel, which is rotated at a suitable speed, and ejected into the discharge cone in the form of thin SECTION M.-M, THROUGH WATER PUMP. Fig. 229a. Westinghouse-Leblanc Multi-jet High- vacuum Condenser System. sheets having a high velocity. These sheets of water meet the sides of the discharge cone and thus form a series of water pistons, each of which entraps a small pocket of air and forces it out against the atmos- pheric pressure. In passing through the air pump the sealing water receives practically no increase in temperature, hence the same water may be used over and over again. The air pump rotor and main pump runner are enclosed in a common casing mounted on the same shaft. 446 STEAM POWER PLANT ENGINEERING This arrangement makes the plant very compact and requires the use of only one motor to drive both pumps. There is a clear passage throurh the condenser and pump, so that should the pump stop for any reason Exhaust from Turbine Fig. 229b. Tomlinson Type C High- vacuum Jet Condenser. air rushes into the condenser through the air pump and immediately breaks the vacuum. In starting up the condenser, steam is turned into auxiliary nozzle L, section N-N , for a few moments, thus creating suf- ficient vacuum to start the regular flow of water through the air pump. Fig. 229c. Section through Condensing Chamber of Kbrting Multi-jet Condenser. Chamber Capable of Maintaining a Vacuum of 95 per Cent of the Ideal without the Use of Air Pumps. The pumps require from H to 3 per cent of the power generated by the main engines. Fig. 223 shows an application of a Westinghouse- Leblanc condenser to a Curtis turbine. CONDENSERS 447 229. Power Consumption of Condenser Auxiliaries. — In estimating the cost of producing vacua with the different types of auxiliaries, steam driven, electrically driven, or belted, the power consumption is most conveniently expressed in terms of the equivalent heat consumption of the auxiliary in question and not the indicated or developed power. For example, suppose a power plant has a number of 1200-LH.P. engines direct connected to 800 -kilowatt generators and that the engines use 20 pounds of steam per LH.P. hour at rated load; furthermore suppose the engine driving the air pump (jet condenser) to indicate 24 horse power. Now, it is manifestly incorrect to say that the power 24 consumption of the air pump is equivalent to = 2 per cent of the main engine power unless the engine driving the air pump uses 20 pounds of steam per I.H.P. As a matter of fact the small engine proba- bly uses 30 to 40 pounds or more of steam per I.H.P. hour, and the true power consumption is 24 X 30 1200 X 20 = 3 per cent, or more. If the exhaust steam is piped to the condenser, then all of this 3 per cent or more should be charged against the condenser; if the steam is piped to a heater, then only the difference between the heat enter- ing the small engine and that given up to the feed water should be charged against it. For example, suppose the engine in the preceding examples uses 30 pounds of steam per I.H.P. hour when running condensing and 40 pounds when operating non-condensing. Let the initial steam pressure be 150 pounds and feed-water temperature 120 degrees F. when the air pump is running condensing. If the boiler feed is not taken from the hot well, the heat in the exhaust steam is lost so far as the economy of the plant is concerned, and the heat con- sumption per I.H.P. hour is 30(1193.6 - (120 - 32)} = 33,168 B.T.U. This represents the cost, in heat units, of producing the vacuum, and is equivalent to 3 per cent of the main engine output. If the air pump runs non-condensing and the exhaust steam is piped to the heater, each pound of exhaust steam gives up approximately 950 B.T.U. per hour to the feed water and the temperature of the latter is raised from 120 to 180 degrees F. The heat entering the air pump is 40(1193.6 -(120 -32)} = 44,224 B.T.U. per I.H.P. hour. But 40 X 950 = 38,000 B.T.U. are returned to the feed water. Hence 44,424 — 38,000 = 6224 is the net heat consumption of the air pumps per I.H.P. hour. This corresponds to approximately 0.55 per cent of the main engine output. 448 STEAM POWER PLANT ENGINEERING In the preceding example suppose the air pump to be motor driven and that it requires 20 electrical horse power per hour. This will be 20 the equivalent of — = 26.2 I.H.P. of the main engine on the 0.85 X 0.90 & assumption that the efficiency of the small motor is 85 per cent and that of the engine and generator combined 90 per cent. The power required by the air pump will be 26.2 4- 1200 = 2.2 per cent of the total output. 8 1 1 1 1 1 1 1 1 1 1 1 1 1 1 1 1 1 a d Ph "3 a 3 ■O RELATION Or POWER CONSUMPTION OF AUXILIARIES TO STATION OUTPUT Citizens Light, Heat and Power Co. Johnsto-wn, Pa. 7 \ 6 \ \ 5 \ Auxiliary Input % of Turbine Output I.H.P. " » » Generator " E.H.P. \ _- 1 - K i *•. «,_ 3 Tot aJ Stp --. 3 -Q \ T^m"" der 2 To Circulat OS L Pi S ^i =s 1- A rl *ump Full .Load 100 300 400 500 Load E.H.P. Fig. 230. 700 In practice the auxiliaries use the equivalent of from 1 to 15 per cent of the main engine or turbine steam, depending upon the size of the plant, character and number of auxiliaries, and the conditions of operation. 11 1 1 ! 1 1 1 1 1 1 1 1 1 1 1 POWER CONSUMPTION OF AUXILIARIES 12 u 2000 K. W. Curtis Turbine La. Purchase Exposition 3 10 3 Q V C^y Pj 8 l^pu to -it xu 9 h 2rjes 6 c ret ila H ' n i 'P u«ip « 4 2 H ->t A Ve LP n _Air_l tar ±£z r~ 1000 1500 LoadiCJV. 2000 Fig. 231. 2500 Table 55 gives the power consumption of the condenser auxiliaries in a number of installations. Fig. 230 shows the relation between the power consumption of the auxiliaries and the total output of the CONDENSERS 449 i o 0? S c bJ s I I or 3 ° ^ °iJ ^T3 ©13 0JD M org bflgL 5§c3gO§ ^ 02 ^ •^88^ •^SUl'ESV lO padranj p^9H NOW50 TJH CM r-H -* CO 00 O CO •sduinj HMONiOrHiO ,_, £q pasfi aa.vioj N(N(M(N(NrtH 00 1—1 -* I^^ox jo iuao J9j ^ J .J ^ P-< P-< Ph Ah" £ & & £ hh' hr! HH w 5=' •sduinj £q r- HHH •H pauinsuoo J9M0J lO "^ i-i t- CM r- OS O CM O C lO CO c i 6 +3 Met) Opei Pu g-rST5*C 5 T 5 T3 0) s GO W # d HP 5ri a> i 1 - c ^2 ± v ) ) ! a o S-! c3 a o So « - a t o 5 i ^2 •s C3 . B O 1 r-i u GO • f"\ *■ L a c. ; .a CD H C Q e i - CO S- to a i C5 - DC 450 STEAM POWER PLANT ENGINEERING station at different loads for a Parsons steam turbine installation, and Fig. 231 shows a similar relation for a 2000-kilowatt Curtis turbine. (J. R. Bibbins, Power, January, 1905.) Steam vs. Electric Auxiliaries : Engr. U.S., 1902, p. 113; Power, Feb., 1905, p. 90, Sept., 1906, p. 502; St. Ry. Review, March, 1898, p. 184, July, 1899, p. 458. Centralized Control of Auxiliaries : Engr. U.S., Nov. 1, 1906, p. 782. 230. Cost of Condensers. — The following figures give an idea of the relative costs of the different types of condensers and auxiliaries for a 1000-I.H.P. plant using 20 pounds of steam per I.H.P. hour at rated load, or a total of 20,000 pounds per hour. Vacuum to be maintained, 26 inches, unless otherwise stated; temperature of cooling water, 70 degrees F.; hot-well temperature, 105 to 120 degrees F.; dis- tance between engine exhaust opening and mean level of intake well, 10 feet. Siphon Condensers. 1 16" siphon condenser with 6" centrifugal pump driven by 6" by 6" vertical engine $800 Jet Condensers. 1 14" by 22" by 24" jet condenser with single horizontal direct-acting pump 1335 1 16" by 24" by 18" jet condenser with single vertical direct-acting pump 1620 1 14" by 24" by 18" jet condenser with single vertical fly wheel vacuum pump 1770 1 12" by 17" by 22" by 25" jet condenser, single horizontal direct- acting compound pump 2200 Barometric Condensers. 1 barometric condenser; 10" by 16" by 12" horizontal single-cylinder rotative dry-air pump; 8" horizontal volute centrifugal pump direct connected to 23-horse-power high-speed engine 2500 1 barometric condenser; 16" by 16" dry-air pump direct connected to 9" by 16" steam engine; positive rotary pump, for circulating cooling water, belted to above engine 4300 Surface Condensers. 1 surface condenser, 1025 square feet cooling surface, mounted over 1\" by 14" by 14" by 12" combined air and circulating pump 2100 1 surface condenser, 1025 square feet cooling surface, with 1\" by 12" by 12" horizontal air pump, direct acting, and 6" centrifugal pump driven by 5" by 5" engine 2300 1 surface condenser, 1025 square feet cooling surface; 5" by 12" by 10" Edwards single-cylinder air pump and 6" centrifugal pump driven by a 5" by 5" engine ; maximum 28", referred to 30" barometer . . 2850 1 surface condenser, 1025 square feet cooling surface; 6" by 8" rotative dry-air pump; 6" by 6" Edwards wet-air pump and 6" centrifugal pump driven by 5" by 5" engine; maximum vacuum 29", referred to 30" barometer (temp, cooling water 50 deg. F.) 3500 CONDENSERS 451 In general the cost of complete condensing equipments installed and ready for operation will approximate as follows: Cost per Kilowatt of Main Generating Unit. Siphon condensers without air pump $2.00 to $ 3.00 Jet condensers 3.00 to 4.50 Barometric condensers with dry-air pump 4.00 to 6.00 Surface condensers for 26-inch vacuum 3.50 to 5.00 High- vacuum surface condensers 3.50 to 10.00 1 1 II 1 1 1 1 1 II 1 1 1 1 RELATIVE COST OP HIGH VACUUM Condens'ng Apparatus | > - 28 be .a -3 a o- -"3 y Comprising ^Surface Condeasers>, Dry Air Pumps, Circulating Pumps, Hot Well Pumps. 27 55 26 / Piping, Valves, etc. / / / OS, / r iceut of( '08 t C f Appa ra US f( '' 6 [Vacuum 100 120 140 160 Fig. 232. 200 The curve in Fig. 232 shows the relative costs of complete surface condensing plants for steam turbines to maintain the vacua indicated. It will be noted how much more expensive a high-vacuum plant is than one designed for moderate vacua. Thus a 27- inch plant costs 25 per cent more than a 26 -inch plant, and a 28.5 -inch plant costs twice as much. (J. R. Bibbins, Power, January, 1905.) The real cost of a condensing plant, however, is not limited to the cost of condensing auxiliaries and piping, but should include all other costs necessitated by the use of the condensing plant, including cost of extra building space, foundations and the like, and the attending fixed charges. 231. Most Economical Vacuum.* — The load factor, or the ratio of the actual yearly load to the rated yearly capacity, has a marked influence on the degree of vacuum best suited for a given installation, since the fixed charges go on whether the plant is running or not, while the gain due to the higher vacuum is realized only when the engines are operating. The higher the load factor the greater is the amount of power produced, the longer does the apparatus operate at best effi- ciency, the lower the ratio of fixed charges to total operating expenses, and consequently the lower the cost of power per unit. The load factor for electric-lighting stations is invariably low and seldom exceeds 25 per cent, with an average not far from 18 per cent. In street-railway work it is higher and averages about 30 per cent. In manufacturing plants the load factor varies considerably, but as a rule is somewhat higher than in either of the above cases. Tables 56 and 57 {Power, December, 1906, p. 769) show the most economical vacua for different load factors for plants of 1000 kilowatts capacity with * See also, Elec'n, Lond , Jan. 14, 1910. 452 STEAM POWER PLANT ENGINEERING conditions as stated. From the tables it would seem at first glance that, except where coal is expensive, all the plants with low factors, 10 per cent and under, ought to be run non-condensing. This is true for " one-engine " installations, but not necessarily so where there are a number of engines or turbines. In the latter case higher economy may be effected by providing only a portion of the engines with condensing equipment. The engine carrying the continuous or day load should operate condensing, and the non-condensing engine should carry the peak load. In order that any of the units may be used for the day work, all engines could be connected to the condenser, but only those carrying the day load should be operated condensing. Each installation, of course, must be considered separately and due weight given to the various factors entering into the problem. For an excellent article on the subject see " Condensers for Steam Engines and Turbines," Power, December, 1906, p. 769, and the Engineer, London, April 13,' 1906, p. 381. 232. Choice of Condensers. — The proper selection of a condenser for a proposed installation depends upon the conditions under which the plant is to be operated. When there is a plentiful and cheap sup- ply of good condensing water suitable for boiler feed, and extremely high vacua are not essential, some type of jet condenser will generally be found most desirable. If overhead room permits, a siphon or baro- metric condenser will probably be most suitable and least expensive. TABLE 56. MOST ECONOMICAL VACUUM FOR STEAM TURBINES. Vacuum referred to 30-Inch Barometer. Cost of Coal, Dollars per Ton. Load Factor, per Cent. $1.50 $2.00 $2.50 $3.00 $3.50 A B A B A B A B A B 5 10 N.C. 20 24 26.5 27.5 28 N.C. N.C. 17 20 24 27.6 N.C. 23 26.5 27.3 27.8 28.2 N.C. N.C. 20 23 27 27.9 18 25 27 27.6 28 28.3 N.C. N.C. 22 25.5 27.6 28 20 26.5 27.5 27.8 28.1 28.4 N.C. 20 24 27 27.8 28 22 27 27.7 27.9 28.2 28.5 N.C. 22 15 25.8 20 27.5 30 28 50 28 A. Surface-condensing plant; cost $6 per kilowatt of main generator. Fixed charges 12 per cent. Cost of water not included. Rated capacity of generator, 1000 kilowatts. B. Surface-condensing plant, including cooling towers and extra cost of land, etc.; cost $10 per kilowatt for 26-inch plant, increasing to $14 per kilowatt for 28.5-inch plant. Fixed charges 12 per cent. No charge for water. Rated capacity of generator, 1000 kilowatts. CONDENSERS 453 TABLE 57. MOST ECONOMICAL VACUUM FOR RECIPROCATING ENGINES. Vacuum referred to 30-Inch Barometer. Cost of Coal, Dollars per Ton. Load Factor, per Cent. $1.50 $2.00 $2.50 $3.00 $3.50 A B A B A B A B A B 10 15 N.C. 16 22.5 24 25.5 N.C. N.C. N.C. 16 22 15 20 23 24.5 26.7 N.C. N.C. N.C. 21 23.5 18 22 23.5 25.5 27.2 N.C. N.C. 20 22 23.5 20 22.5 24.5 26.4 27.5 N.C. 16 21 23 26.3 22 24 25 26.8 27.7 N.C. 20 20 22 30 24 50 27 A. Surface-condensing plant; cost $7 per kilowatt of main generator. Fixed charge 12 per cent. Cost of water not included. Rated capacity of generator, 1000 kilowatts. B. Surface-condensing plant, including cooling towers and extra cost of land, etc.; cost $11 per kilowatt for 26-inch plant, increasing to $13 per kilowatt for 27.5-inch plant. Other conditions as in A. Where there is a plentiful supply of good water for boiler feed but the water which must be used for cooling purposes is very dirty the siphon condenser is preferable to the barometric form. A surface con- denser may be used in the latter case if the condensing water is not so dirty as to seriously impair the efficiency by coating the tubes with sediment, and boiler feed water is scarce. The air-cooled surface condenser is employed only where water of any kind is scarce. For very high vacua in connection with steam turbine work the sur- face condenser is almost universally adopted, although the barometric condenser in connection with dry-air pumps is finding favor with many engineers. In selecting the type of condenser and auxiliaries due weight must be given to the load factor, cost of coal, water, land, building, interest, depreciation and the like, as outlined in the preceding paragraph. 233. Water-Cooling Systems. — When an ample supply of cooling water is unobtainable for natural or economic reasons, the circulating water may be used over and over again by employing suitable cooling devices. The three most common in practice are 1. The simple cooling pond or tank. 2. The spray fountain. 3. The cooling tower. 454 STEAM POWER PLANT ENGINEERING 233a. Cooling Pond. — The water is cooled partly by radiation and conduction but principally by evaporation. The air is seldom saturated normally, and its capacity for absorbing moisture is increased on account of its temperature being raised by contact with the warm water and by radiation. The cooling action is independent of the depth of water and varies directly as the surface, the amount of heat dissipated for each square foot depending upon the temperature of the water, the rela- tive humidity, and the velocity of the air currents. Results of tests are very discordant. Box in his treatise on Heat states that the pond surface should approximate 210 square feet per nominal horse power for an engine working twenty-four hours a day. (Treatise on Heat, Box, p. 152.) If the engine works only twelve hours per day, the area may be reduced to 105 square feet per horse power, because the water will cool during the night, but in that case the depth should be such as to give a capacity of 300 cubic feet per horse power. These figures are based on a reduc- tion in temperature of 122 to 82 degrees F., with air at 52 degrees F. and humidity 85 per cent, the steam consumption per nominal horse power being taken at 62.5 pounds. Box gives the following formula for the rate of evaporation in per- fectly calm air: E = (243 + 3.7 (V-v), (95) in which E = evaporation in grains per square foot per hour. t = temperature of the water, degrees F. V = maximum vapor tension in inches of mercury at temperature t. v = actual vapor tension. Evaporation is greatly affected by the force of the wind and varies from 2 to 12 times the amount determined from equation (95). Example: How many pounds of water will be evaporated per square foot per hour from a pond with the temperature of the water and air 80 degrees F.; air perfectly calm; barometric pressure 29.5 inches and relative humidity 70 per cent? The maximum vapor tension at temperature of 80 degrees is 1.02 inches of mercury. The actual vapor tension will be 1.02 X .70 ( = relative humidity) = .714. Substitute these values in (95). E = (243 + 3.7 X 80) (1.02-0.714) = 165 grains per square foot per hour. = .023 pound per square foot per hour. CONDENSERS 455 If the temperature of the water were 130 degrees F. and that of the surrounding air 80 degrees F., humidity 70 per cent, the evaporation would be E = (243 + 3.7 X 80) (4.5-0.714) = 2040 grains per square foot per hour = 0.291 pound per square foot per hour. Here 4.5 = maximum vapor tension, corresponding to a temperature of 130 degrees. 233b. Spray Fountain. — From equation (95) we see that even under the most favorable circumstances an enormous pond surface is necessary. To facilitate evaporation with a view toward reducing the size of the pond, the hot circulating water is sometimes distributed through pipes and discharged through nozzles, falling to the surface of the pond in a spray. The following gives some interesting data con- 100 9 GO 2.40 / \ 1 | 1 1 Water at Nozzles "\ J I [— ^y \-T ^J / iN f Water in Pond A r \ , / \ r 1 \> U' w / \ 7 iL A / — J.' ir 1/ \, j \ / / y \ z_ L t 10 13 1G 1 9 22 25 28 L i t r i 13 1G 19 22 25 28 § 90 ^ 80 A A fs |100 *80 J» W 60 CD I 50 « 40 $T i h 8U *70 Jh r\ A \ \ A I J v. y ' / J \ J \ / / \ \ ^ i a 00 3 N \ f v \ s V > v J V V> 3 i0 (2 30 i V V 1 4 7 10 13 16 19 22 25 28 Day o£ the Month September, 1901 4 7 10 13 16 19 22 25 Day -of the Month January, 1905 Fig. 233. Curves Showing Performance of Spray Fountain; Chattanooga Electric Company's Power Plant. cerning the spray fountain installation at the power plant of the Chat- tanooga Electric Company, Chattanooga, Tenn. {Street Railway Review, March 15, 1905.) Adjoining the power house a pond 150 x 300 feet was excavated to a depth of 4 feet, the level of the water being 8 feet below the condensers. 456 STEAM POWER PLANT ENGINEERING Circulating water returned from the condensers is distributed through a set of pipes provided with 42 nozzles through which the water is dis- charged upwards. The rectangle denned by the center lines of the outermost pipes is 98 feet by 125 feet. The pipes are supported on brick piers spaced at intervals of about 20 feet in each direction. The discharge opening of the nozzles is 1J inches in diameter, and the interior is provided with a spiral core so that in its passage the water is given a rotary motion, the effect of which is to greatly increase the spraying action. The nozzles, except on the extreme outer lines of piping, are placed in pairs with the axes in a vertical plane at right angles to the center line of the supply pipe, the axis of each nozzle making an angle of 30 degrees with a vertical plane through the center of the supply pipe. The effect of each pair of nozzles is to throw a mass of spray to the height of about 15 feet, which in falling covers an area of 15 x 30 feet. A dike extending nearly across the pond near one end provides a canal through which the water is conducted to the suction chamber, the object being to draw the supply from distant parts of the pond to give greater time for cooling. The " make-up " water is supplied by wells. The operation of the cooling pond for a warm month and for a cold month is shown in Fig. 233. Readings were taken at three-hour inter- vals. The pond supplies the circulating water for three 2000-square- feet Worthington surface condensers. 234. Cooling Towers. — A cooling tower consists of a wooden or sheet-iron housing open at the top and bottom and so arranged that the heated cooling water may be elevated to the top and dis- tributed in such a manner that it falls in thin sheets or sprays into a reservoir at the bottom, air at the same time being drawn in at the bottom by natural draft or forced in by a fan. The water gives up its heat to the ascending current of air by evaporation and conduction, the latter, however, being a relatively small factor. If the air supply is dependent entirely upon convection, the system is known as the natural- draft or flue cooling tower; if the air is forced into the tower by fans, it is called a fan cooling tower. The different types vary principally in the method of water distribution. Fig. 234 illustrates the Barnard cooling tower, in which the falling water is broken up by vertically suspended galvanized iron wire cloth mats, causing it to trickle in thin sheets to the bottom. A similar result is brought about in the Worth- ington tower, Fig. 235, by pieces of terra-cotta pipe 6 inches in diame- ter and two feet long placed on ends in rows. In the Alberger cooling tower the water trickles down the sides of swamp-cypress boards arranged in honeycomb fashion. In the Jennison cooling tower the CONDENSERS 457 water is divided into a rain of drops, constantly retarded in their fall by a series of perforated 4x4 inch galvanized iron trays arranged in horizontal rows and staggered vertically. With the best forms of cooling towers, under average conditions, the DISTRIBUTING TROUGH DISCHARGE FROM TOWER Fig. 234. Barnard- Wheeler Cooling Tower. temperature of the circulating water may readily be reduced from 40 to 50 degrees with a loss not exceeding 3 or 4 per cent of the total quantity of water passing through the tower. The power consumed by the fan in a forced-draft apparatus averages 2 per cent of that 458 STEAM POWER PLANT ENGINEERING developed by the main engines, for the maximum requirements during summer months, and 1 J per cent during the winter. The location of the tower may be on the engine-room floor, on top of TOWD\ < ■ CC « HOT WATER. COLD WATER,. Suction tank Fig. 235. Worthington Cooling Tower. the building, or in the yard, the latter being the most adaptable. It may be any reasonable distance from the engine and condenser. Fig. 236 shows a typical installation of Alberger condenser and cooling towers. COiNDENSERS 459 460 STEAM POWER PLANT ENGINEERING 235. Parallel Comparison of Fan and Natural-Draft Cooling Towers. Fan. Natural Draft. Size. Small, the forced draft providing Large, draft being necessarily small, sufficient air velocity to effect evapo- a larger area must be provided to ration. perform same work. Height limited, because loss from back Height is an advantage because the pressure increases with the height. tower operates on the principles of a Tower usually short and of large area. chimney. Power Consumption. One per cent of station output and None. upwards, depending upon the type of auxiliaries and the conditions of operation. Location. Inside or outside. Can operate in any Outside only, unless exceptionally good location where sufficient head room draft is obtainable. and air supply are available. Preferably in the open where advan- Especially adapted to inconvenient lo- tage may be taken of prevailing cations, as roofs, upper decks, boiler winds. floors, etc. Conditions of Atmosphere. Comparatively little affected by tern- Largely affected by temperature and perature, considerably by humidity, humidity and wind. Draft increased and none by winds. by steady winds. Conditions of Operation. More especially adapted for heavy con- Especially adapted for light summer tinuous duty the year round, as in and heavy winter duty, as in electric- rail-plants or mills. lighting plants. First Cost and Cost of Operation. First cost greater on account of First cost small by reason of simplicity mechanical construction and neces- and construction. sary auxiliaries. First cost largely dependent upon ma- Cost of operation dependent upon type terials used in interior construction. of auxiliary and conditions of oper- Cost of operation limited to fixed ation. charges. 236. Water-cooling Calculations. — Air is said to be completely saturated when it contains all the water vapor it can hold without CONDENSERS 461 causing precipitation. If the vapor content is less than that corre- sponding to complete saturation the air will tend to become saturated by absorbing moisture from surrounding objects. The drier the air the greater will be its affinity for moisture. The necessary latent heat for vaporization is supplied directly by the water producing the vapor or by the surrounding objects in contact with the water. Thus, in the open cooling-tower the water vapor is absorbed from the circulating water, and the heat necessary to effect this vaporization is given up by the water, with a resultant reduction in temperature of the water itself; and in the evaporative surface condenser the vapor is absorbed from the water spray in contact with the tubes, the heat required to effect this vaporization being given up by the steam within the condenser chambers, resulting in condensation of the steam. If the air coming in contact with the water is very dry and at a high temperature the vaporization of the water may be rapid enough to cool the remaining water to a temperature much lower than that of the air. In this case practically all of the cooling is effected by evaporation. But when the air is at a low temperature and high relative humidity a considerable amount of heat may be carried away by the air by conduction. The quantity of air and water necessary to produce a given cooling effect may be determined as follows: Let H = total amount of heat to be abstracted, B.T.U. per hour. W = weight of water to be cooled, lbs. per hour. t e = temperature of water entering cooling device. ti = temperature of water leaving cooling device. t = temperature of air entering cooling device, °F.;T = £ + 460. U = temperature of- air leaving cooling device, ° F. ; T 2 = t 2 + 460. p = ordinary atmospheric pressure = 29.92 in. of mercury. p a = observed atmospheric pressure, in. of mercury. p = elastic force of vapor at temperature t , in. of mercury. p 2 = elastic force of vapor at temperature t 2 , in. of mercury. V = volume of air entering the cooling device, cu. ft. per hour, atmospheric conditions. V 2 = volume of air discharged from the cooling device at tem- perature t. d = density of dry air, at pressure p and temperature t . h = weight of moisture in 1 cu. ft. of saturated air at tempera- ture t , pounds. h 2 = weight of moisture in 1 cu. ft. of saturated air at tempera- ture t 2 , pounds. z = relative humidity of the air entering the cooling device. z 2 ==Telative humidity of the air leaving the cooling device. 462 STEAM POWER PLANT ENGINEERING a < is 06 fe ?Sj3 «5 O Q^ H CQ OO O ceos •^ 9inj'Bjadui9x J3d 99.1§9(l 9U0 P9UIJBAY jiy pa^jn^Bg ^99j oiqno •fi\ia J9d 99lS9(I 9XIO P9UIJBA1 jjv &xi %&& o?qno ONNIN(N(NONO(NN(NO0©XXX©CNl>X"5.-lXXiO'H©©»0'tfOO«0 Or-iT-iO(N©©(NCCfO©ce^H©05iO-<*' •J 99lS9G[ J9d JTV pg^BiniBg aooj oiqiio 9UQ £q pgqjosqv £L\La *J 99J39CI J9d Jiy Ai(J aooj oiqno 9UQ £q pgqaosqy 'fl'L'Q. 'f umnTOo in sb 9jnss9it^l>t^t>t^©©©©©©CO oboooooooooooooooooooo <© T j<,_i_,c< l t^00'Ht^I>T* l O©-*i-HO300iC(N©t^'*i-iOil>TfcOC ©0©0300000000t^t^l>l>CO©CD©iO>OOkO , *TfH-*T}iT}. CNCN-Hl-l'-ll-ll-l^.-ltH.-li-li-ll-l.-lTH.-l.-I^H.-l.-I^HrH.-lrH ©©©0©©©©0©0©©©0©©©©0©©0©© ©-< X t> OS © © "* *£> ©© X © »0 © X »-0 i-O CO © -H lO © ^TH^Xi-lrHTtO»OC0^©©©©-*r-l^i-l^1^Me0© 7-1 -H »0 >-i X © T* 00 ON <-l OS-*©©IXN©X©iO^COeoTtX©©XCS^l-OXI>r^©'-Hi-HI>OOCO(N^«OCN>Ot^^H'*t^eOeO©CN-- OOOOOOHHHHiMcO^TfOOOHONHCOHlNOC ©©©©©©0©©©©©©©©©rHrt(M-H©^-H^t>^-i|>CN©©CO©«0(M©CNiO© I>CCOO'*(MiOX©CN|©^H©©-^-*COt^»-i^HOOCOTf < i*(N CO'HCO'-OXiO>0C000(NO©00l>i0C0©00i0t^t^-t^©©©©©©»o»oio^<^^*©i-l«l^I>©00t^©CN©C0©i0CN©OcN©iOTj»Hr-(OOCO'*'*CN OrH(NCC'*OG000'-H(N©(M00©©'-H©'*^ , *©i0^l000 t~ OOOOOOOOHrtrHNNNMiacOOCOWOOiSOCO 0©©0©©©©©©©0©0©0©©^h—"-i.cO , * , *©'-<©Tt<'<*t^C«3©00LO©© CO"#CN©X©»0 - <#COeNOX©»-OCO©©.t~r-r-t>'t^©©©©»oioiOT}HTt> § d^S TjHO-^»OCC©C000CCC^^©G0©O©L0©!MC0©©t^© 00XXt»©»-OCOCO'-l'-lX''tf©X'-iCO'-it>©X<-OOCN©Ttt^©'0©©t-©t-'-H05©t-©X'#© o ^^1-l-HCNl^^cv^'*©l>©X©CNC0 - <*C0XO't-X©©©'-" _ 8320 20 X 500 10,000 ' that is, the equivalent of 83.2 per cent of the steam used by the engines is evaporated in the cooling tower, or the make-up water is more than supplied by the condensed steam. Example: Evaporation Surface Condenser. — How many cubic feet of air and how many pounds of water spray must be forced through an evaporative surface condenser of the fan type in order to condense 1000 pounds of steam per hour and maintain a vacuum of 25 inches, barometer 29? (Atmospheric air 80° F., relative humidity 70%.) The air and vapor issue from the discharge pipe under pressure of four inches of water, temperature 120° F., relative humidity 98%. The absolute pressure in the condenser is 29.0 — 25.0 = 4 inches of mercury. The total heat to be withdrawn in order to cool and condense 1000 lbs. of steam per hour at absolute pressure of 4 inches to 120° F. is 1000 [1114.8 - (120. + 32)] = 1,026,000 B.T.U. Neglecting radiation and leakage losses, this is the heat to be ab- stracted per hour by the air and water spray. The pressure of the dry air in the mixture entering the condenser is, equation (96), p 1 = 29.0 - 0.7 X 1.029. = 28.28. The pressure of dry air in the mixture leaving the condenser is, equation (96a), p 3 = (29.0 + 0.294) - 0.98 X 3.438 = 25.925 (0.294 is the value in inches of mercury of four inches of water-fan pressure). 1 CONDENSERS 467 Let V = volume of atmospheric air entering the condenser. The vol- ume leaving the condenser will be, equation (98), _ 28.280 460 + 120 _ V °~ 25.925' 460+80 " ' V °' The weight of vapor in the condenser discharge is, equation (98a), w 2 = 1.172 V X 0.004888 X 0.98 = 0.005615 V lbs. The weight of vapor in the mixture entering the condenser is, equa- tion (97a), Wq = 0.00157 X 0.7 V = 0.001099 V lbs. The amount evaporated therefore is w 3 = 0.005615 7 - 0.001099 V = 0.004516 V lbs. The weight of dry air entering the condenser is, equation (97), 90 90 W = 29321 007362 ^ = 0.06958 y lbs. The heat absorbed by the dry air in being heated from 80° to 120° F. is, equation (100), H = Cw (t 2 - t ) = 0.02375 X 0.06958 V (120 - 80) = 0.658 V B.T.U. Heat required to superheat w lbs. of vapor from 80° to 120° F. is, equation (100a), ^ = 0.001099 V X 0.46 (120 - 80) = 0.02022 V B.T.U. Heat absorbed by the evaporation of w 3 lbs. of water is, equation (101), H e = 0.004516 y X 1046.7 = 4.720 V B.T.U. (Here the latent heat is taken at the lower temperature, it being the original temperature of the liquid.) Total heat absorbed by the entering air and spray is H t = 0.658 7 + 4.720 V + 0.020 V - 5.398 y . But this represents also the heat given up by the steam, or 5.398 y = 1,026,000. From which V = 190,500 cu. ft. of atmospheric air necessary to con- dense and cool 1000 pounds of steam under the given conditions. The water spray to be injected per hour is 0.004516 y = 0.004516 X 190,500 = 860 pounds. 468 STEAM POWER PLANT ENGINEERING 236a. Hygrometry. — The degree of saturation, or relative humidity, is ordinarily determined from the difference in reading of a wet and a dry bulb thermometer, thus : If the air is saturated with aqueous vapor no evaporation takes place from the wet bulb and the two thermometers give identical readings; but if it is unsaturated, evaporation occurs. The wet-bulb thermometer is thus cooled and its readings are lower than those of the dry bulb. The difference in reading is a function of the relative humidity, and the latter may be calculated from the following modification of Apjohns' formula: If the thermometer reads above 32° F. j, / dP \100 htes { v "~2m) TT (103a) If it reads below 32° F. h in which h = relative humidity, per cent. d = difference in reading of the wet and dry thermometers, degrees F. P = barometric pressure, inches of mercury. P w — maximum tension of aqueous vapor corresponding to the temperature of the wet thermometer, inches of mercury. (This may be taken directly from the Steam Tables.) P t = maximum tension of aqueous vapor corresponding to the tem- perature of the dry thermometer, inches of mercury. Example: Determine the relative humidity when the dry bulb reads 70° F., wet bulb 60° F., barometer 28.0. From the Steam Tables we find P w = 0.522; P t = 0.739. Whence eoo 10 X 28\ 100 • 522 "-264(r)a739 = 56 - 5percent - Tables giving the relative humidity in terms of the temperature difference are published in most engineering handbooks and the above calculations are unnecessary. These tables, however, are based on a fixed barometer pressure, whereas the formula takes the actual pressure into consideration. 237. Test of Cooling Tower (Wheeler Condenser Company), — The following gives the results of a test made on the cooling-tower plant of the A. F. Brown Company at Elizabethport, N. J. The tower is work- ing in connection with a Wheeler surface condenser of 280 square feet of cooling surface, mounted over a 10, 12X12 combined air and circulating pumn. („. ! CONDENSERS 469 Observations made on June 24, 1904. Temperature of air 81 degrees Hygrometer 69 degrees Temperature of air at top of lower 89 degrees Temperature of water in troughs 105 degrees Temperature of water in tank 83 cleg Revolutions of fun, 239 r.p.m.. air pressure \ inch water Velocity of air out of tower 822 feet per minute Gallons of water passing over mats 385 per minute Vacuum 20 inches Temperature of air-pump discharge 87 degrees Observations made June 28, J 904, 9 a.m Temperature of air 70 degrees Hygrometer 59 degrees Temperature of air at top of tower 81 degrees Temperature of water in troughs 96 degrees Temperature of water in tank 78 degrees Revolutions of fan, 232 r.p.m., air pressure jj inch water Velocity of air out of tower 080 feet per minute Gallons of water passing over mats 400 per minute Vacuum 25.5 inches Temperature of air-pump discharge 90 degrees Observations made June 28, 1904, 3 P.M. Temperature of air 74 degrees Hygrometer 57 degrees Temperature of air at top of tower 83 degrees Temperature of water in troughs 99 degrees Temperature of water in tank 80 degrees Revolutions of fan. 237 r.p.m , air pressure ^ inch water Velocity of air out of tower 709 feet per minute Gallons of water passing over mats 470 per minute Vacuum ... 2.5.5 inches Temperature of air-pump discharge 92 degrees Observations made June 29, 1904. Temperature of air 78 degrees Hygrometer 71 degrees Temperature of air at top of tower 80 degrees Temperature of water in troughs 108 degrees Temperature of water in tank 82 degrees Revolutions of fan, 211 r.p.m., air pressure § inch Velocity of air out of tower 772 feet per minute Gallons of water passing over mats 430 per minute Vacuum 25.5 inches Temperature of air-pump discharge 93 degrees Specifications for condensers — See paragraph 414. RESULTS OF TEST OF NATURAL-DRAFT TOWER, DETROIT. Complete Five-Fifths Surface Installed. Proc. A.S.M.E.. Mid-Nov., 1909, p. 1205. Engines: Two 400-i.h.p. 300-kw. Macintosh & Seymour tandem-compound engines, overhung generators. Condensers: Worthington surface (admiralty type) 1000-sq. ft. reciprocating wet- air pump and circulating pump Tower: Wood-mat construction, 24,500 sq. ft evaporating surface, exclusive Of shell Test- March 15 to 10, 1901, 4 p.m. to 4 p m , 24 hr. 470 STEAM POWER PLANT ENGINEERING Weather; Load: Steam : Water: Results: Cooling: Evaporati Tower: A.M. 30.22 18.5 76 P.M. AVERAGE. 30.07; 30.14 30.27 25; 30 25 82; 58 72 Barometer (abs.), min Temperature air, deg Relative humidity, per cent 600 kw. max. to 50 kw. min. Average 244.9 kw. Engine efficiency = 92.5 = 875 i.h.p. max. Average . .354.8 i.h.p. Weight of condensed steam per hr., lb 5910.6 Temperature exhaust steam, deg. F 134 . 38 Temperature condensed steam, deg. F 108 . 78 Weight of steam per hour, max. load, lb 13,500 Vacuum (abs.) 25 to 19, average about 22 Vacuum corresponding to temperature exhaust steam. . . 25 Vacuum possible with good condenser (10 deg. difference) 28 Circulated per hr., lb 293,536 Temperature hot well, average, deg. F 87.50 Temperature cold well, average, deg. F 71.27 Vaporization loss per hr., lb 5970 Condenser surface per kw., sq. ft 2.66 Steam per kw. hr., lb 24 . 3 Steam per i.h.p. hr., lb 16.66 Circulating water per lb. of steam, lb 49 . 6 Steam per sq. ft. condenser surface per hr., lb 3.7 Circulating water per sq. ft. tower surface, lb 12 Difference in temperature between exhaust steam and discharge, deg. F 47 Max. 20 deg., min. 3 deg. -5 deg. Average 16.23 Heat dissipated per hr., B.T.U 4,769,000 Heat per sq. ft. tower surface, B.T.U 195 Heat per sq. ft. per 1000 lb. water, B.T.U 0.665 Circulating water, per cent 2 .03 Engine steam, per cent 101 Surface per kw. (average load 245 kw.), sq. ft 100 Surface per kw. (max. load 600 kw.), sq. ft 40.8 Surface per 1000 lb. steam max. load, sq. ft 1820 Surface per 1000 lb. steam average load, sq. ft 4140 Surface per 1000 lb. circulating water per deg. max. cool- ing, sq . ft 4.17 Temperature, Deg. Fahr. Quantities. Time. Air. Hot Well.* Cold Well. Water Cool- ing. Total Heat Head.t Tower Water, Lb. per Hr. Heat Dissi- pated, B.t.u. Lb. per Hr. Heat per Sq. Ft. Cool- ing Surface, B.t.ii.perHr. Circulating Water per Sq. Ft., Lb. per Hr. Load, Kw. 1 2 3 4 5 6 7 8 9 10 11 12noon 34 102 89 13 68 375,000 4,880,000 332 25 270 1.30 35 106.5 90 16.5 71.5 ♦ (375,000 1370,200 6,108,000 415 24.8 )315 1290 2.30 35 106.5 87.5 19 71.5 375,000 7,120,000 484 25 315 3.30 35 113 88.5 24.5 78 375,000 9,000,000 613 25 350 4.30 32.5 100 84 16 67.5 399,000 6,384,000 434 26.6 365 5.00 28.5 103.5 88 15.5 75 445,500 6,900,000 470 29.7 485 6.00 26 125 94 31 99 417,000 12,930,000 880 27.8 655 7.00 24 121 94 27 97 427,000 11,532,000 785 27.4 570 8.00 24 123 94.5 28.5 99 427,000 12,174,000 827 27.4 600 * Assuming a more efficient condenser, say 10 deg. difference, the probable vacuum would be 26 deg. to 27.5 deg. This condenser actually operated at 40 deg. to 50 deg. difference. t Total heat head = air heating + lost head. J Difference due to rapid change in load. CHAPTER XII. FEED-WATER PURIFIERS AND HEATERS. 238. General. — All natural waters contain more or less foreign matter either in suspension or solution. Waters containing carbonates and sulphates of magnesia and lime, soluble salts of silica, iron, and alumina, and suspended matter, tend to form scale in the boiler and reduce its steam-generating capacity and economy. The loss due to this cause is often overestimated but is of secondary importance to the danger due to retarded heat transmission which overheats and weakens the plates and tubes. Table 59 gives the results of a number of tests made on loco- motive boiler tubes with different thicknesses and characters of scale. The diversity of the results indicates the futility of bas- ing the decrease in conductivity on the thickness of the scale. For example, test No. 1 shows a decrease in conductivity of 9.1 per cent for a scale .02 inch thick, while No. 16 shows a decrease of only 6.75 per cent for a scale over 6.5 times as thick. The scale in each case was even, hard, and dense. Again, No. 8 with a very soft scale .042 inch thick gives a decrease in conductivity of 9.54 per cent, whereas No. 14, also very soft but twice as thick, gives a decrease of only 4.95 per cent. No doubt the heat transmission is a function of the chemical as well as the physical properties, but further experiments are necessary before any specific conclusion can be drawn. Waters containing acids, organic matter, and magnesium chloride and sulphate tend to corrode the boiler, and those containing sodium carbonate, organic matter, and alkalies induce priming. Even distilled water, as obtained from a surface condenser, is a solvent of iron to a certain extent and causes corrosion and pitting. Table 60 gives some idea of the character and extent of impurities in water from various localities, with an analysis of the scale produced by the water and the trouble in the boiler arising from its use. It is impossible to judge the quality of feed water merely by the grains of solids per gallon, since a large amount of soluble salt such as sodium chloride will not be as deleterious as a very small amount of calcium sulphate. 471 172 STEAM POWER PLANT ENGINEERING TABLE 59. INFLUENCE OF SCALE ON HEAT TRANSMISSION. (Locomotive Boiler Tubes.) No. Thickness of Scale, Inches. Character of Scale. Decrease in Con- ductivity due to Scale. Per cent. 1 .02 .02 .033 .033 .038 .04 .04 .042 .047 .065 .07 .07 .085 .089 .11 .13 Hard, dense Hard Soft Very hard Medium Soft, porous Hard, dense Very soft Hard Medium Soft Hard Soft, porous Very soft Hard, porous Hard, dense 9 1 2 2 02 3 4 3 4 3.5 5 4.03 6 6.82 7 3.07 8 9.54 9 2 75 10 2.39 11 2.38 12 4.43 13 19.0 14 4.95 15 16.73 16 6.75 From tests conducted at the University of Illinois, Railroad Gazette, Jan. 27, 1899, June 14, 1901. See also Engineering Record, Jan. 14, 1905, p. 53; Power, February, 1903, p. 70; Street Railway Review, July 15, 1901, p. 415. The following is a rough rating according to the number of grains of incrusting solids per United States gallon: Less than 8 grains very good. 12 to 15 grains good. 15 to 20 grains fair. 20 to 30 grains bad. Over 30 grains very bad. This applies to calcium carbonate, magnesium carbonate, and mag- nesium chloride. For water containing sulphate of calcium and mag- nesium, divide the first column by 4 for the same rating. On account of the great variety of possible impurities the proper treatment to be adopted can be determined only by chemical analysis of the feed water in each case. Table 61, compiled by the Hartford Steam Boiler Inspection and Insurance Company, shows the number of boilers inspected by that company during the year 1907 and the number found defective from various causes. FEED-WATER PURIFIERS AND HEATERS 473 m GO >H < s5 ■d 03 S !3< CQ O 0) ifi £ ^ « 1 ^ "5 J -g ^ 03 S § O 00 iO CM •>* »0 „, © tJh CO ONOOWN "U5 00O) IflrlCOHOO ^(OlOO 00;i-iOr-l tHl 1—1 ©iOOOOOOOSOOi-hcM (OOHIOOOMONO ©OCMOO-^t-hcM©© MHHOMOOO^M NOON!Dh(NI>OOH MONCO(NffiO>OCO lOtOHiOOS J?< "*" >0 00 03 <*0-H(MN ..0(0 00 CO OS CO CO OS £ N CO lO -* © n as © 5$cooo .2 O t a el 1 ^ S d<+H o o .S © °§ ■ coOo ■A c CO © .sag 8.3.2 . Cj CO CO a a © o %^ 03 ° © fl d d £ o *a 03 03 «.d " d d a a -2 d^ .d O d d Qrs C Q,,0 - •— • - — • ?3 CO o d 03 O o^-OJ £ OQ § © .5 d NOCOtO " 00 CO .03.. «3 (M CS OS > m N CM ■ CO • -0 T5 CD t-i T— 1 5J c3 i xT o3 O CO OO rt< -* iO CO o • gJeMoot^oooot^r- iO • 08 s. S CO (M 00 cs o ■* io i-H • d rHio © - © M(NQ0(O ^OOOCMM • CQ ft N^»^ r ^O)NNO ■ -Q -2 (O O •* OO CR SO Oj 00 • ffi co n m fh oo to CO • o3 w •* H CO IO CO CO CO o • r-l IO £*ri O CO CM 00 o^OS IO IO © !3 O CM O "* 53 "* CO >o IO kT o3 >rd OS CO OS >0 r~\ i-t CO i— I CM t - ' i— I i-H CO CM « ol CM CM O CO £ -* CO ■*0r-l05 5s OS CO . 03 Tt(HO)O r ^iOCO CM CM HNH o • OS • Ph H --T *-< ccooococn ^ © © i>. eo -* ? © deOi-HCNCO V CM t~ CM CM CM J3CLO • • • -J 3 • . ■ • • • °3 G r -' c ^' :0, - ,, *- 00T -"= , O»O 1 1 w a > y-i CM CO t-l r-t co co kH 11.18 10.44 40.96 Trace 22.60 00 CO CM • 1—1 h-1 d m .3 a -* O CM CM ^ t-i CM -* CO CM 00 2 O CM . 03 . . CO tH N M > CN CO MCOnH OS • o OS • CO ^1 o • T3 -rl OOCO fflrffOOffi ®M ■ CDCOCO dO'CHW 2<» ■ . . . g . . . 03 . . O O CO^ CO to CM £ O • (Nr-HCO|2i T-H i-l H M5 E bl © d d H-9 o3 CL P 4 a pi a a "53 © d bJO o3 /- a o "8 03 © d • 'a : © ' 03 OQ * a a ao o +3 © • d ^ a «h ^73 ^ : fl^-s • S © O d> ^"^ 03 so © o ^ "03 ° ta S ® ? d "§ d SR ^ T3 a o; t R X c 1 OE 4 C ji p ! E cc !l 3 -^ c.2 ISC C s I *81 CD ■<-* S <»30 §.2 ^8^ cd E si I s o® ■^ fi ©.2 o ^ W ' w 3 3 ' o bio a ft 03 5 S a o£ h bXJCD^-? c _- S d o > to a ■ii « .§■§ lis _,.t3 o c 3^©SftS ft 2ii§§ §g§8 H! o CO CD 3 5 S.a;d g %$*& ! ! § !1 g-2^g 8 M ^=! ! g f cd S 3^5 « *3 ■0 c3"q X3 -5 5 © 0) *^ CO O g ° <»_ d So O CD — ? "S d be O g ca S Id § 3 1° 8 d • _ & o &© H Eh bug xj ©x: "H-, M «>iS ^"d "5 474 STEAM POWER PLANT ENGINEERING TABLE 61. SUMMARY OF INSPECTORS' REPORTS FOR THE YEAR 1907. (Hartford Steam Boiler Inspection and Insurance Company.) Nature of Defects. Whole Number. Dangerous. 18,917 1,315 38,427 1,333 3,010 258 12,802 528 10,230 768 2,219 578 6,363 699 7,564 396 3,551 568 4,878 499 898 92 3,582 823 1,764 238 11,357 1,599 8,266 3,054 1.947 563 5,557 430 3,008 707 4,216 1,250 413 156 1,231 415 1,211 407 7,651 465 194 194 27 10 Cases of deposit of sediment. . . Cases of incrustation and scale Cases of internal grooving Cases of internal corrosion. . . . Cases of external corrosion. . . . Defective braces and stays Settings defective Furnaces out of shape Fractured plates Burned plates Laminated plates Cases of defective riveting. . . . Defective heads Leakage around tubes Cases of defective tubes Tubes too light Leakage at joints Water gauges defective Blow-offs defective Cases of deficiency of water . . . Safety valves overloaded Safety valves defective Pressure gauges defective Without pressure gauges Unclassified defects Totals 159,283 17,345 The neutralization or elimination of the impurities may be effected by one of the following methods: 1. Chemically. Boiler compounds. Purifying plants. 2. Mechanically. Filters. Blow-off. Tube cleaners. 3. Thermally. Feed-water heater. Distillation. FEED-WATER PURIFIERS AND HEATERS 475 The following chart (" Boiler Waters," W. W. Christie) outlines some of the troubles arising from feed water, their cause and means for preventing them. Trouble. Cause. Remedy or Palliation. Incrustation. Corrosion. Priming . Sediment, mud, clay, etc.. Readily soluble salts Bicarbonate of magnesia, lime, iron Organic matter. . Sulphate of lime < Organic matter Grease Chloride or sulphate of ) magnesium ( Sugar } Acid f Dissolved carbonic acid and oxygen Electrolytic action Sewage ■] Alkalies Carbonate of soda in large ) quantities ) Filtration. Blowing off. Blowing off. Heating feed and precipitate. Caustic soda. Lime. Magnesia. See below. Sodium carbonate. Barium chloride. Precipitate with alum ) Precipitate with ferric > and filter chloride ) Slaked lime Carbonate of soda Carbonate of soda. and filter Alkali. Slaked lime. Caustic soda. Heating. Zinc plates. Precipitation with alum or ferric chloride and filter. Heating feed and precipitate. Barium chloride. Analysis of Water for Softening by Chemical Processes: Jour. Soc. Chem. Ind., May, 1899, p. 520. Volumetric Determination of Calcium and Magnesium in Water: Jour. Soc. Chem. Ind., May 31, 1901, p. 507. Simple Tests for Boiler Water: Soc. Engng., May, 1904, p. 238. Qualitative Tests of Feed Water: Power, Dec, 1904, p. 756. New Testing Apparatus for Boiler Feed Water: West. Elecn., Aug. 1, 1903, p. 85. A Simple Method of Calculating Water Analyses and Amounts of Substances to be Added for Preventing Scale and Corrosion: Jour. Frank. Inst., Vol. CLIX, p. 217. Description of Dearborn Drug and Chemical Co.'s Laboratories for Analyses of Boiler Feed Water: St. Ry. Review, Sept. 15, 1901. Boiler Corrosion: Power, Jan., 1906, Oct., 1905, p. 591, Dec, 1900; Eng. Rev., Oct., 1904, p. 12; Eng. Mag., Dec, 1905, p. 425, Oct., 1900, p. 118; Engng., Oct. 10, 1902, p. 482; Engr. U.S., Jan. 1, 1907, p. 103, May 1, 1902, p. 388; Engr., Lond., Aug. 5, 1904, p. 131, July 29, 1904, p. 115, Dec. 4, 1896, p. 574; Elec Age, Dec, 1905, p. 456; Am. Elecn., Aug., 1905, p. 436, April, 1902, p. 184; Jour. Am. Soc Nav. Engrs., May, 1902; Mines and Min., Sept., 1903; Stahl u. Eisen, Jan. 15, 1904. Boiler Incrustation: Am. Elecn., April, 1904, p. 206, Dec, 1901, p. 576, May, 1901, p. 220, Oct., 1898, p. 473; Cassier's Mag., July, 1903, p. 273; Chem. News, 476 STEAM POWER PLANT ENGINEERING Oct. 18, 1901, p. 191; Engr., Lond., Jan. 21, 1898, p. 52; Engr. U.S., April 16, 1906, March 15, 1904, p. 202, May 15, 1904, p. 354, Sept. 1, 1904, p. 608; Ice and Refrig., Nov., 1905, p. 173. Chemistry of Scale: Jour. Frank. Inst., Aug., 1901, p. 113, Aug. 1891, p. 145. Boiler Scale: Power, Dec, 1905, p. 779 ; St. Ry. Review, April 2, 1904, p. 545. Scale Prevention: Am. Elecn., Dec, 1901, p. 578; Am. Engr., May, 1900, p. 138; Eng. Mag., 1897, 12-959, 13^74, 232, 419; Elec Engr., Lond., July 20, 1900, p. 91; Engrs. Gaz., July, 1902. Foaming. — Foaming Water and Scaling Water for Locomotive Boilers: Eng. News, July 21, 1904, p. 71, Sept. 1, 1904, p. 198; Foaming and Priming: Soc Nav. Arch, and Marine Engrs., 1902; R.R. Gazette, Oct. 12, 1900; Christie, Boiler Waters, Chap. V; Stromeyer, Steam Boilers, pp. 67-83; Rowan, Modern Steam Boilers, pp. 321-354. 239. Chemical Purification. — Chemical treatment of boiler feed water has been remarkably developed during recent years and a number of manufacturing concerns make this their sole business. The two most common systems of chemical treatment involve (1) boiler compounds and (2) purifying plants. In the former the necessary chemical action takes place inside the boiler and in the latter the water is purified before it enters the boiler. In either case the usual procedure is to submit for analysis a sample of the feed water and the resulting scale to a competent chemist who will specify the character and quantity of chemicals necessary to bring about the desired result. 240. Boiler Compound. — The object of treatment with boiler com- pounds is to neutralize the evil effects of the impurities in the feed water or to change them into others which are less objectionable and which are easily removed. When properly compounded and intro- duced into the boiler such preparations are of great benefit and prac- tically overcome the deleterious effects, but when improperly used they may produce even greater troubles than the impurities which they are expected to eliminate. Boiler compounds may be divided into three classes : 1. Those converting the scale-forming elements into new sub- stances which will not form a hard, resisting scale and which are readily removed by skimming, blowing off, or by tube cleaners. For example, feed water containing sulphates of lime and magnesia will form a dense, tenacious scale. If carbonate of soda be added in correct amount, the sulphates are converted into insoluble carbonates which are precipitated and form scale varying from a more or less porous, friable crust to a soft " mush " or mud. The resulting sulphate of soda remains in solution and does not form scale unless allowed to concentrate, and this is prevented by blowing off. An excess of soda FEED-WATER PURIFIERS AND HEATERS 477 is apt to cause foaming and at high temperatures is liable to attack the inside of gauge glasses. Bisodium and trisodium phosphate, sodium t annate, fluoride of sodium, sugar, etc., have all proved satis- factory, but as each case requires special treatment no detailed dis- cussion is possible within the scope of this work and the reader is referred to the accompanying bibliography. 2. Those enveloping the newly precipitated scale-forming crystals with a surface which prevents them from cementing together. The ingredients used to bring about this result are starches, woody fibers, dextrine, slippery elm, and the like. 3. Those preventing the formation of hard scale by a solvent or " rotting " action, as kerosene and petroleum oils. Boiler Compounds. — Use of Compounds: Eng. News, July 27, 1905, p. 112; Am. Mach., Dec. 7, 1899, p. 115, Oct. 26, 1899, p. 1014; Power, Aug., 1903; Eng. and Min. Jour., Aug. 12, 1905, p. 253. 241. Use of Kerosene and Petroleum Oils in Boiler Feed Water. — Kerosene oil and other refined petroleum oils are sometimes used with good effect in boilers to prevent scale from adhering. These oils are said to change the deposit of lime from a hard scale to a friable material which may be easily removed. They are ordinarily fed to the boiler with the feed water, drop by drop, through a sight feed apparatus similar to a cylinder oil lubricator. From extended experiments made on a 100-horse-power tubular boiler fed with water containing 6.5 grains of solid matter per gallon it was found that one quart of kero- sene per day was sufficient to keep the boiler entirely free from scale. Prior to the introduction of the oil the water had a corrosive action upon some of the fittings attached to the boiler, but after the oil had been used for a few months it was found that the corrosive action had ceased. In another case 40 gallons of kerosene were used in 24 hours in a steamer of about 3000 horse power. These boilers showed no incrustation but considerable corrosion. Evidently oil does not have the same effect or give the desired results in all cases. Kerosene used in moderate quantities will not cause foaming. Crude oil should never be used, as the heavy residue causes the formation of a tough, imper- vious scale productive of bagged sheets and collapsed flues. Use of Kerosene in Boilers: Engr. U.S., Sept. 15, 1905, p. 634; Eng. News, May 24, 1890, p. 497; Power, Aug., 1895, p. 13, May, 1896, p. 16; Trans. A.S.M.E., 9-247, 11-937; Locomotive, July, 1890, p. 97. 242. Use of Zinc in Boilers. — Zinc is often introduced into boilers to prevent corrosion. The theory is that a feeble but continuous cur- 478 STEAM POWER PLANT ENGINEERING rent of hydrogen is generated over the whole extent of the iron by electrolytic action. The bubbles of hydrogen formed isolate the metallic surface from scale-forming substances. If there is but a little of the scale-forming element it is precipitated and reduced to mud; if there is considerable, coherent scale is produced which takes the form of the iron surface but does not adhere to it, being prevented from doing so by the intervening bubbles of hydrogen. Zinc is ordinarily sus- pended in the water space of the boiler in the shape of blocks, slabs, or as shavings in a perforated vessel. Electrical connection between the metallic surfaces is essential. Rolled zinc slabs 12x6xJ inches have found much favor in marine practice. Generally speaking one square inch of zinc surface is sufficient for every 50 pounds of water in the boiler, though the quantity placed in the boiler should vary with the hardness. The British Admiralty recommends the renewing of the zinc slabs whenever the decay has penetrated to a depth of J inch below the surface. Zinc does not prevent corrosion or scale formation in all cases and may even aggravate the trouble. Use of Zinc in Boilers: Am. Elecn., Dec, 1901, p. 572; Kent, Steam Boiler Econ- omy, p. 318; Christie, Boiler Waters, p. 137; Stromeyer, Marine Boiler Management and Construction, p. 81. 243. Methods of Introducing Compounds. — Boiler compounds may be introduced into the boiler continuously or intermittently. Small quantities introduced continuously or at short intervals are more effective than large quantities at long intervals. Continuous feeding is ordinarily brought about by connecting the suction side of the feed pump with a reservoir containing the compound in solution, arranged similarly to an ordinary cylinder oil lubricator. In large plants an independent pump is often used to force the solution into the feed line. Intermittent feeding is brought about by temporarily connecting the suction of the feed pump with the reservoir containing the compound. The use of boiler compounds does not necessarily prevent scale from forming in time, though it will reduce the evil to a minimum. In some instances where compounds are used it is found necessary to run a tube cleaner through the tubes at certain intervals, in others such a course has not been found necessary. 244. Weight of Compound Required. — The weight of compound introduced depends upon the nature of the reagents used and the character of the feed water, and ranges from a few ounces to several pounds per 100 gallons of feed water. For example, water containing 4 grains of calcium sulphate and 6 grains of magnesium sulphate per gallon will require 3.57 pounds of carbonate of soda per 1000 gallons FEED-WATER PURIFIERS AND HEATERS 479 of water for the reduction of the sulphates. The chemical reaction and analysis is as follows: CaS0 4 + Na 2 C0 3 = CaC0 3 + Na 2 S0 4 MgS0 4 + Na 2 C0 3 = MgC0 3 + Na 2 S0 4 If x = grains of Na 2 C0 3 necessary for the calcium, CaS0 4 : Na 2 C0 3 + 10H 2 O = 4 : x. 40 + 32 + 4 X 16 : 2 X 23 + 12 + 3 X 16 + 10 (2 + 16) = 4 : x. x = 8.41 grains. If y = grains of Na 2 C0 3 necessary for the magnesium, MgS0 4 : Na 2 C0 3 + 10H 2 O = 6 : y. 24 + 32 + 4 X 16 : 2 X 23 + 12 + 3 X 16 + 10 (2 + 16) = 6 : y. y = 14.3. The total weight of carbonate of soda per 1000 gallons is therefore 1000 (14.3 + 8.41) = 22,710 grains. = 3.24 pounds. This amount would effect the desired result if the chemical reaction is permitted to take place for some time, otherwise an excess of reagent is necessary. 245. Mechanical Purification. — Waters containing sand, mud, organic matter, and in fact all matter which is not in solution or in chemical combination with the water may be purified by mechanical filtration. Mud and sand may be eliminated by simply permitting the water to stand for some time in settling tanks. Suspended matter which will not gravitate to the bottom may be removed by filtering the water through coke, cloth, excelsior, or the like. Filters should be in duplicate for continuity of operation. Vegetable and other organic impurities commonly float on the sur- face of the water when the boiler is making steam, and may be blown out through a " surface blow-out." (See paragraph 82.) Precipitated matter may be ejected from the boiler by fre- quent blowing off before it has time to adhere and bake to a crust. This procedure is particularly essential when boiler compounds are used. For description and use of mechanically operated tube cleaner see paragraph 86. 246. Thermal Purification. — (See also Live Steam Purifiers, para- graph 263.) The carbonates of lime and magnesia are held in solution 480 STEAM POWER PLANT ENGINEERING in fresh water by an excess of carbon dioxide and are completely pre- cipitated by boiling. At ordinary temperatures carbonate of lime is soluble in approximately 20,000 times its volume of water, at 212 degrees F. it is slightly soluble, and at 290 degrees it is insoluble. Sul- phate of lime is much more soluble in cold than in hot water, and is com- pletely precipitated at 290 degrees. (Revue de Mecanique, November, 1901, pp. 508, 743.) Thus it will be seen that a feed heater may be relied upon to remove part or all of the lime, depending upon the tem- perature to which the water is raised and the time in which the pre- cipitation is permitted to take place. Influence of Temperature and Concentration on the Saline Constituents of Boiler Water : Jour. Soc. Chem. Ind., Oct. 31, 1900, p. 885. Solubility of Sulphate of Lime: Rev. de Mecanique, Jan., 1901, p. 5, Nov., 1901, p. 508. 247. Purifying Plants. — The function of a purifying plant is the elimination of all impurities from the feed water before it enters the boiler. In the Scaife system for water purification feed water first enters the heater, where it attains a temperature of from 200 to 210 degrees F. As a portion of the free C0 2 is driven off by the heat the carbonates of lime and magnesia are precipitated and are deposited in removable pans inside the heater. On its way the heated water is forced by the boiler feed pump into a large precipitating tank, where the necessary chemicals are introduced by means of two small pumps. These pumps take the solution of chemicals from the solution tanks which hold a sufficient quantity to operate the plant from eight to twelve hours. The precipitating tank is so constructed as to cause intimate and thorough mixing of the chemicals with the water. Thus the acids are neutralized, and the scale-forming substances are precipitated by being changed to insoluble substances which sink to the bottom of the precipitating tank whence they are readily removed. Some of the lighter substances remaining in suspension are carried along with the water as it passes into the filters, which effectively remove all suspended matter. This system is continuous in operation, and purification is accomplished without appreciably retarding the onward flow of feed water. Fig. 237 shows a modification of the system. The chemicals are pumped from the " chemical tank "into the " solution tanks," where the feed water and chemical solution are thoroughly mixed. The treated water is taken from these tanks and pumped into the " precipitating tanks " where a large portion of the scale-forming element is precipitated. From the precipitating tanks the water is forced through a series of filters to the boiler. Fig. 238 illustrates the We-Fu-Go system of water purification. In FEED-WATER PURIFIERS AND HEATERS 481 PRECIPITATING TANKS L_JL_Ji J i i i i i ! L_j L_J L.i Fig. 237. General Arrangement of Scaife System of Feed-Water Purification. Fig. 238. General Arrangement of We-Fu-Go System of Feed- Water Purification. 482 STEAM POWER PLANT ENGINEERING this installation the water supply first enters the settling or treating tanks into which the chemicals are fed. A thorough mixture is effected by the use of the two armed paddles located near the bottom of the tanks. From the treating tanks the water flows by gravity into the filters, which remove all remaining impure solid matter which does not settle to the bottom of the treating tank. The pipes conducting the water from the settling tanks to the filter are fitted with a flexible joint and float so that the outlets are near the surface at all times, rising and falling with the water level. From the filters the purified water gravi- tates into the clear water storage reservoir, from which it is pumped into an open heater and thence to the boiler. This system is intermit- tent in operation, and in order to provide sufficient time for thorough chemical treatment of large quantities, two or more settling tanks are employed. Both the We-Fu-Go and Scaife systems are modified in a number of ways to meet different conditions. Fig. 238a. Anderson System for Preventing Corrosion in Condensers. Fig. 238a shows the general arrangement of the Anderson system for preventing corrosion in condensers and removing oil from condensed steam. The method consists in injecting into the exhaust steam as it passes from the preheater to the condenser a solution containing a coagulant which changes the emulsion of the cylinder oil to a flaky condition so that it may be separated by settling, flotation, or filtering. The air pump delivers the water to the settling tank F, whence it is taken to the open gravity filters G, G, of a superficial area proportional to the amount of water to be passed and containing a filter bed of four feet of crushed quartz. This will run about four days without any marked difference in efficiency, after which time the bed is stirred to a depth of two feet by mechanical agitators and flushed with clean water, by which all impurities are carried to the sewer. The solution is pre- FEED-WATER PURIFIERS AND HEATERS 483 pared in tank A, in which the water level is preserved by a ball float and into which filtered water is admitted through pipe B, while the substance with which the water is treated is pumped in through the pipe D by a small pump operated from the main engine. The flow to the " rose head " above the condenser is controlled by the valve E, and a meter in this pipe records the amount being fed. The water ordinarily required for " make up " is sufficient to carry in the solution. There is very little loss of water, and the rapid corrosion of the con- denser tubes, which has been so great an obstacle to the successful use of surface condensers, is much reduced. The chemicals used perform a twofold duty, viz., to neutralize the water and make it chemically inactive and to coagulate the oily matter contained in the steam so that mechanical filtration is possible. (Power, June 1903, p. 304.). Water-softening plants cost from $4 to $5 per horse power for plants of 1000 horse power and less, from $3 to $4 for plants of 1000 to 2000 horse power, and as low as $1.50 for plants of 5000 horse power or more. The depreciation of wooden tanks is as high as 15 per cent a year, while that of steel tanks should not be greater than 5 per cent. Unless wooden tanks are considerably cheaper than steel tanks they are not a good investment. The cost of water purification varies from a fraction of a cent to 2 cents per 1000 gallons, depending upon the size of the plant and the quantity and character of the impuri- ties. (American Electrician, March, 1905, p. 125.) Feed Water Purification: Am. Elecn., March, 1900, p. 145, April, 1900, p. 190, Dec, 1904, p. 618; Cassier's Mag., April, 1904, p. 506; Engng., Oct. 25, 1901, p. 595; Eng. News, May 22, 1902, p. 408; Eng. Rec, April 5, 1902, p. 322, June 10, 1905; Jour. Amer. Chem. Soc, Nov., 1893, p. 610; Jour. Soc. Chem. Ind., Aug., 1901, p. 828; Power, Nov., 1900, p. 7, Sept., 1902, p. 33, Nov., 1904, p. 693; R.R. Gaz., Aug. 24, 1900, p. 568; Elec. Rev., Nov. 12, 1904; Engr. U.S., Oct. 15, 1903, Jan. 1, 1906; Elec. World, Sept. 1, 1906. Water-Softening Plants. — Water-Softening Processes: Prac. Engr. U.S., Mar , 1910. Four Systems of Softening Water for Industrial Purposes: Eng. News, July 2, 1903, p. 4. An Inquiry into the Working of Various Water Softeners: Inst, of Mech. Engrs., Dec, 1903. Report on Soft Water for Locomotive Plants: Eng. News, March 17, 1904. The Development of Water Purification in the U.S.: R.R. Gaz., 38-19. Gen- eral Information on Water Softening: Eng. News, May 26, 1904, p. 500, 508, June 2, 1904, p. 530; Eng. Rec, Oct. 24, 1903, p. 483; Loco. Engng., Nov., 1903, p. 501; Elec. Engr., Lond., April 21, 1905; Ir. and Coal Tr. Rev., Sept. 1, 1905; Jour. W. Soc. Engrs., Dec, 1905; Eng. and Min. Jour., Dec. 2, 1905; Am. Engr. and R.R. Jour., Jan., 1905. Harris- Anderson Water Softener: Engng., Aug. 15, 1902, July 10, 1903. Kennicott Water Softeners: Am. Elecn., Nov., 1902, p. 545; Eng. News, May 15, 1902, p. 386; Eng. Rec, May 3, 1902, p. 419; St. Ry. Jour., April 2, 1904, p. 545. 484 STEAM POWER PLANT ENGINEERING Burt Continuous Water-Softening Process: Eng. News, Sept., 15, 1904, p. 238; Engr. U.S., July 15, 1905, p. 426. Bachman Method of Water Purification: St. Ry. Review, May 15, 1900, p. 282. We-Fu-Go and Scaife Systems: St. Ry. Rev., Oct., 1901, p. 771; Engr. U.S., Jan. 1, 1903, p. 90. Holmes System of Water Purification: Power, April, 1905, p. 248. The American Water Purifier and Softener: Eng. U.S., Aug. 1, 1904, p. 551. 248. Economy of Preheating Feed Water. — Although a feed water heater acts to some extent as a purifier its primary function is that of heating the feed water. Generally speaking, for every 10 degrees that the feed water is heated there is a gain in heat of 1 per cent and a corresponding saving of coal, if the heat which warms the feed water would otherwise be wasted. Again, the smaller the difference in temperature between the steam and the feed water the less will be the strain on the boiler shell due to unequal expansion and contraction, an item of no small consequence. If X represents the total heat of one pound of steam above 32 degrees F., t the temperature of the cold water, and t the temperature of the water leaving the heater, then S, the per cent gain in heat due to heating the feed water, may be expressed S = 100 x ( *~* o) ^ • ( 104 ) X~(t -32) The expression is not theoretically correct, since it assumes a con- stant value of unity for the specific heat, whereas the specific heat varies with the temperature. The variation is so slight, however, that it may be neglected for all practical purposes. Example: Steam pressure 100 pounds gauge; temperature of water entering heater 80 degrees F.; temperature of water leaving heater 210 degrees F. Required, saving due to heating the feed water. Here \ (from steam tables) is 1185, t = 80, t = 210. S =10 o (210-80) 1185- (80-32) = 11.42 per cent. This formula gives the thermal saving only, and the first cost of the heater, interest, depreciation, attendance, and repairs must be taken into consideration before the net saving measured in dollars and cents is ascertained. In the average installation the net saving is a sub- stantial one. FEED-WATER PURIFIERS AND HEATERS 485 Table 62 based upon formula (104) may be used in determining the percentages of saving due to the increase in feed-water temperature. TABLE 62. PERCENTAGE OF SAVING FOR EACH DEGREE OF INCREASE IN TEMPERATURE OF FEED WATER. Initial Boiler Pressure above Atmosphere. Temp. of Feed. 20 40 60 80 100 120 140 160 180 200 32 .0872 .0861 .0855 .0851 .0847 .0844 .0841 .0839 .0837 .0835 .0833 40 .0878 .0867 .0861 .0856 .0853 .0850 .0847 .0845 .0843 .0841 .0839 50 .0886 .0875 .0868 .0864 .0860 .0857 .0854 .0852 .0850 .0848 .0846 60 .0894 .0883 .0876 .0872 .0867 .0864 .0862 .0859 .0856 .0855 .0853 70 .0902 .0890 .0884 .0879 .0875 .0872 .0869 .0867 .0864 .0862 .0860 80 .0910 .0898 .0891 .0887 .0883 .0879 .0877 .0874 .0872 .0870 .0868 90 .0919 .0907 .0900 .0895 .0888 .0887 .0884 .0883 .0879 .0877 .0875 100 .0927 .0915 .0908 .0903 .0899 .0895 .0892 .0890 .0887 .0885 .0883 110 .0936 .0923 .0916 .0911 .0907 .0903 .0900 .0898 .0895 .0893 .0891 120 .0945 .0932 .0925 .0919 .0915 .0911 .0908 .0906 .0903 .0901 .0899 130 .0954 .0941 .0934 .0928 .0924 .0920 .0917 .0914 .0912 .0909 .0907 140 .0963 .0950 .0943 .0937 .0932 .0929 .0925 .0923 .0920 .0918 .0916 150 .0973 .0959 .0951 .0946 .0941 .0937 .0934 .0931 .0929 .0926 .0924 160 .0982 .0968 .0961 .0955 .0950 .0946 .0943 .0940 .0937 .0935 .0933 170 .0992 .0978 .0970 .0964 .0959 .0955 .0952 .0949 .0946 .0944 .0941 180 .1002 .0988 .0981 .0973 .0969 .0965 .0961 .0958 .0955 .0953 .0951 190 .1012 .0998 .0989 .0983 .0978 .0974 .0971 .0968 .0964 .0962 .0960 200 .1022 .1008 .0999 .0993 .0988 .0984 .0980 .0977 .0974 .0972 .0969 210 .1033 .1018 .1009 .1003 .0998 .0994 .0990 .0987 .0984 .0981 .0979 220 .1029 .1019 .1013 .1008 .1004 .1000 .0997 .0994 .0991 .0989 230 .1039 .1031 .1024 .1018 .1012 .1010 .1007 .1003 .1001 .0999 240 .1050 .1041 .1034 .1029 .1024 .1020 .1017 .1014 .1011 .1009 250 .1062 .1052 .1045 .1040 .1035 .1031 .1027 .1025 .1022 .1019 Multiply the factor in the table corresponding to any given initial temperature of feed water and boiler pressure by the total rise in feed-water temperature; the product will be the percent- age of saving. Feed Water Heating. — How Should Feed Water be Heated ? — Power, July, 1907, p. 456; Feed Water Heating: Engr. U.S., Jan. 1, 1906, p. 8, Aug. 15, 1904, p. 15; St. Ry. Jour., July 22, 1905, p. 145; Am. Elecn., Dec, 1904, p. 570; Am. Elecn., Nov., 1904; Engr., Lond., July 28, 1905. 249. Classification of Feed- Water Heaters. — Feed-water heaters may be classified according to the source of heat, as 1. Exhaust steam, in which the heat is received from the exhaust of engines, pumps, etc. 2. Flue gas, in which the waste chimney gases are the source of the heat. 3. Live steam purifiers, or those using steam at boiler pressures; or according to the method of heat transmission, as 486 STEAM POWER PLANT ENGINEERING 1. Open heaters, in which the steam and feed water mingle and the steam in condensing gives up its heat directly to the water. 2. Closed heaters, in which the steam and water are in separate chambers and the steam gives up its heat to the water by conduction. Heaters may also be classified according to the pressure of the heat- ing steam, as 1. Vacuum or primary, in which the pressure is less than atmos- pheric and applies particularly to heaters utilizing the exhaust of con- densing engines. These are always of the closed type. Open heaters in which the pressure is less than atmospheric are not usually classed as vacuum heaters. 2. Atmospheric or secondary, in which the pressure is atmospheric or, literally, that corresponding to the back pressure on the engines and pumps. 3. Pressure, in which the pressure corresponds to that in the boiler and in which the heat is used primarily for purifying purposes. Heaters may be still further classified as 1. Induced, in which only such steam is admitted as is induced by its condensation. That is, the feed water condenses the steam. This creates a partial vacuum which draws in more steam. 2. Through, in which all the steam is forced through the heater irrespective of condensation. CLASSIFICATION OF A FEW TYPICAL HEATERS. Exhaust steam Open Atmospheric ~, . ( Atmospheric Closed.. ]„ ( Vacuum or pressure Cochrane Hoppes Stillwell Webster Wainwright ) Water Wheeler . . . ) Tube Otis ) Steam Berryman . j Tube {Green American Sturtevant Live Steam Open Pressure ] ~ ,, ^ ( Baragwanath 250. Open Heaters. — Fig. 239 gives a sectional view of a Cochrane special feed heater and receiver and is a typical example of an open heater. Exhaust steam enters the heater through a fluted oil separa- tor as indicated, and passes out at the top, while the oily drips are automatically drained to waste by a suitable ventilated float. The feed water enters through an automatic valve and is distributed over FEED-WATER PURIFIERS AND HEATERS 487 a series of copper trays so arranged and constructed that the water is forced to fall in a finely divided stream before reaching the reservoir in the bottom. The steam coming in contact with the water particles gives up latent heat and condenses. Much of the scale-forming ele- ment is deposited on the surface of the trays, from which it is readily removed. The suspended matter is eliminated by a coke filter in the Fig. 239. Cochrane Special Heater and Receiver. bottom of the chamber, and the floating impurities are decanted by a skimmer or overflow weir. The particular heater shown in the illustra- tion is especially designed for use in a steam-heating plant; i.e., besides performing all the functions of an open heater, it provides for the reception and heating of the condensation returned to it from the heating system. 488 STEAM POWER PLANT ENGINEERING Fig. 240 gives a sectional view of a Webster " star vacuum " heater. W^ater enters the heater through balanced valve F, which is controlled by float E, and is deflected over a series of perforated copper trays T, T. Exhaust steam enters at A, passes through oil filter S, and, mingling with the finely divided streams of water, gives up its latent heat and is condensed. Only so much steam enters the heater as is condensed by the feed water. The condensed steam and feed water fall to the bottom Fig. 240. Section Through Webster Heater. of the upper chamber, maintaining a practically constant level WW. From this upper or heater chamber the water gravitates to the settling chamber at the bottom, through down-cast pipe CB. From the set- tling chamber the water rises through perforated screen M and filtering material P to the outlet 0. A large portion of the scale-forming ele- ment is precipitated on the trays or collects in the settling chamber at the bottom. FEED-WATER PURIFIERS AND HEATERS 489 Fig. 241 shows a section through a Hoppes open heater, illustrating the " pan " type. Exhaust steam enters at H, passes through oil filter 0, and completely surrounds pans T, T. The feed water enters at B, and the rate of flow is regulated by valve F, which is controlled by a Fig. 241. Hoppes Horizontal Feed-Water Heater. suitable float in the lower part of the chamber. The water in flowing over the sides and bottoms of the pans comes in direct contact with the steam. 251. Combined Open Heater and Chemical Purifier. — Combined feed- water heaters and chemical purifiers are finding increased favor with engineers in many districts where the feed water is particularly bad. A description of the Webster combination will be found in Part II of the general catalogue issued by the Warren Webster Company, Camden, N.J. A description of the Cochrane-Sorge combined heater and chem- ical purifier will be found in the heater catalogue issued by the Harrison Safety Boiler Works, Philadelphia, Pa. 252, Temperatures in Open Heaters. — The temperature to which feed water is raised in an open heater may be determined as follows : Let A represent the total heat of steam corresponding to the pressure in the heater, t the temperature of the water entering heater, t the temperature of the water leaving heater, and S the ratio of exhaust steam to the feed water, by weight. 490 STEAM POWER PLANT ENGINEERING Then, allowing a loss of 10 per cent due to radiation, etc., 0.9 S (a — t + 32) will be the B.T.U. given up by the exhaust steam to each pound of feed water, and (t — 1 ) will be the B.T.U. absorbed by each pound of water. Therefore 0.9 S (X - t + 32) = t t , from which = t + 0.9 S (X + 32) 1 + 0.9 S (105) If more steam passes through the heater than can be condensed by the feed water, then this equation gives t a fictitious value; in other words, t can never be greater than the temperature of the exhaust steam. Substituting t = 212, the maximum obtainable temperature with exhaust steam at atmospheric pressure, and solving for S, we find that only 17 per cent of the main engine exhaust is necessary to heat the feed water to a maximum. t is assumed to be 60 degrees F. Table 63 has been determined from this equation and gives the final temperatures obtainable in open heaters for various conditions of operation. TABLE 63. FINAL FEED- WATER TEMPERATURES. OPEN HEATER. (Temperature of steam, 212 degrees F.) Initial Temperature of Feed Water , Degrees F. 40 50 60 70 80 90 100 110 120 130 2 60.1 69.9 79.7 89.5 94.4 109.2 119.0 128.8 138.7 148.5 p 3 69.9 79.6 89.3 90.1 108.8 118.6 128.3 138.0 147.8 157.5 a 4 79.5 89.1 98.8 108.5 118.1 127.8 137.4 147.1 156.7 166.4 I s 5 89.0 98.5 108.1 117.7 127.2 136.8 146.4 155.9 165.5 175.1 ;i 6 98.3 107.7 117.2 126.7 136.2 145.7 155.2 164.7 174.2 183.6 7 107.4 116.8 126.2 135.6 145.0 154.4 163.8 173.2 182.5 192.1 o 2 8 116.4 125.7 135.0 144.4 153.7 163.0 172.4 181.8 191.0 200.3 O >j 9 125.2 134.5 143.7 153.0 162.2 171.5 180.7 190.0 199.2 208.5 t* 10 133.3 143.1 152.3 161.4 170.6 179.8 189.0 198.1 207.3 212.0 1 11 142.5 151.6 160.7 169.7 178.9 188.2 197.0 206.2 212.0* 212.0* 1 di 12 150.9 159.9 168.9 177.9 187.0 196.0 205.0 212.0* 212.0* 212.0* * All of the steam not condensed. Example : A power plant has 1200 I.H.P. of engines using 20 pounds of steam per I.H.P. hour. Auxiliaries use equivalent of 10 per cent of main engine steam. Pressure in heater pounds gauge, temperature of hot- well supply 110 degrees F. Required temperature of feed water leaving heater. FEED-WATER PURIFIERS AND HEATERS 491 Here X = 1146 (from steam tables), t = 110, S = 0.10. Substituting these values in (105), 0.9 X 0.10 (1146 - 1 + 32) - t - 110. t = 198 degrees F. 253. Pan Surface Required in Open Feed- Water Heaters. — Pan or tray surface required varies according to the quality of the water with regard to both scale-making material and mud, and may be approxi- mated by the formula Pan surface, sq. ft. ^ Lb- of water heated per hr.X horsepower : (1Q5a) Vertical Type. Horizontal Type. For very muddy water, c 118 166 500 110 Slightly muddy water, c 155 For clean water, c 400 254:. Size of Shell, Open Heaters. — General proportions of open heaters vary considerably on account of the different arrangements of pans or trays, filter and oil-extracting devices. A fair idea of the size of shell required may be obtained by the formulas . - , „ Horse power Area of shell = — — ; Sr~. — 7— ■ a X length in feet (106) Length of shell - Horsepower . a X area in square feet. a = 2.15 for very muddy water. a = 6 for slightly muddy water. a = 8 for clean water. The horse power in this case is obtained by dividing the weight of water heated per hour by the steam consumption of the engine per horse power per hour. Pans containing 2.5 square feet and less are usually made round, and larger sizes rectangular in plan. When circumstances will permit it is better to have not more than six pans in any one tier, since it is advisable to proportion the pans so as to obtain as low a velocity over each as practicable. Distance between trays or pans is seldom less than one-tenth the width for rectangular and one-fourth the diameter for round pans. Volume of storage and settling chamber in horizontal heaters varies 492 STEAM POWER PLANT ENGINEERING from 0.25 for good quality of water to 0.4 of the volume of the shell for muddy water, 0.33 being about the average. In the vertical type the settling chamber represents respectively 0.4 and 0.6 the volume of the shell with clear and muddy water. Filters occupy from 10 to 15 per cent of the volume of the shell in the horizontal type and from 15 to 20 per cent in the vertical type, the smaller percentage corresponding to clear water and the larger to muddy water or water containing a con- siderable quantity of impurities. Open Heaters: Cassier's Mag., Aug., 1903, p. 33; Engr. U.S., Jan. 1, 1906, pp. 17, 78; St. Ry. Jour., Feb. 4, 1905, p. 227; Am. Elecn., Sept., 1905, p. 481. SURFACE BLOW EXHAUST FROM HEATER 255. Classification of Closed Heaters. — Closed heaters may be grouped into two classes: 1. Water tube, Fig. 242, and 2. Steam tube, Fig. 246. Closed heaters, both water tube and steam tube may oper- ate with: 1. Parallel currents, where the water and steam flow in the same direction, Fig. 242, or with 2. Counter currents, where the water and steam flow in opposite directions, Fig. 244. Fig. 242. Goubert Single-Flow Closed Heater. Fig. 243. Details of Expansion Joint, Goubert Heater. Water-tube heaters may be still further classified as 1. Single- flow, in which the water flows through the heaters in one direction only, Fig. 242. FEED-WATER PURIFIERS AND HEATERS 493 2. Multi-flow, in which the water flows back and forth a number of times, as in Fig. 244. 3. Coil heater, in which the water flows through one or more coils, as in Fig. 245. 256. Water-Tube, Closed Heaters. — Fig. 242 shows a section through a feed-water heater of the single-flow straight-tube type. The ot/rtEr Fig. 244. Wainwright Multi-Flow Closed Heater. Fig. 245. Typical Coil Heater. tubes are of plain brass and the shell of cast iron. The tubes are expanded into the tube sheets by a roller expander. To provide for expansion the upper tube sheet and water chamber are secured to the main shell by means of a special expansion joint the details of which are shown in Fig. 243. R is a ring or gasket of soft annealed copper 494 STEAM POWER PLANT ENGINEERING and G, G two gaskets of special packing with brass wire cloth insertion. These gaskets form a flexible expansion joint between C and tube sheet D, so that the whole upper chamber, which is carried solely by the tubes, is free to move up and down as the tubes expand or contract under varying temperatures. Fig. 244 shows a section through a Wainwright heater, illustrating the multi-flow water-tube type. The body of the heater is of cast iron, the tubes of corrugated copper. The water passes through the tubes and the steam surrounds them. The feed water and exhaust steam do not mingle, and hence the oil in the exhaust does not contaminate the water. The water chambers are divided into several compartments, as shown in the illustration, and the par- titions are so arranged that the flow of feed water is directed back and forth through the various groups of tubes in succession. This arrangement gives a higher velocity of flow than the non-return type of heater, and therefore increases the rate of heat absorp- tion. The mud and impurities settle at the bottom and are discharged through the mud blow-off. Such impurities as rise to the surface are removed by the surface blow-off. The tubes are cor- rugated to allow for expansion and at the same time to increase the transmission of heat. Referring to Fig. 244: Exhaust steam enters at A and leaves at E, and the portion which is condensed is drawn off at D. Feed water enters at / and is discharged at 0. P, P are mud blow-offs and S is an opening for a safety valve. Table 66 gives results of tests showing the relative efficiencies of plain and corrugated tubes for various velocities. Fig. 245 shows a partial section through a Harrisburg feed-water heater. This apparatus is a typical example of the coiled-tube heater. Three sets of concentric copper coils are brazed to gun-metal manifolds and supported by clamp stays as indicated in the illustration. Feed water enters the heater at the bottom manifold and passes through the coils to the feed outlet. The exhaust steam enters the heater at the bottom and surrounds the coils in its passage to the outlet at the top. The coils are designed to withstand a pressure of 600 pounds per square inch. 257. Steam-Tube, Closed Heaters. — Fig. 246 shows a section through an Otis heater, illustrating the steam-tube type. Here the exhaust steam passes through the tubes which are surrounded by the feed water. The exhaust steam enters at A, and passes down one section of tubes into the enlarged space of the water and oil separator 0, in which the condensation and oil are deposited. From this chamber the steam passes up through the other section of tubes to outlet C, thus FEED-WATER PURIFIERS AND HEATERS 495 passing twice through the entire length of the heater. The water enters at E and is discharged at G. R is the blow-off opening. The tubes are of seamless brass and are curved to allow for expansion. Condensed steam is withdrawn at P. Fig. 247 shows a partial section through a Baragwanath steam jacketed steam-tube heater. Exhaust steam enters at A, passes up #M Aft Fig. 246. Otis Steam-Tube Feed-Water Heater. Fig. 247. Baragwanath Steam-Jacketed Feed- Water Heater. through the tubes, returns down annular space E between the inner shell and jacket, and passes out at B. Feed water enters at C and leaves at D. E is the scum blow-off, G the heater drain, and H the jacket drain. 496 STEAM POWER PLANT ENGINEERING 258. Heating Surface, Closed Heaters. — It is generally assumed that the transfer of heat between two bodies is directly proportional to the difference in temperature between them. Let T = temperature of the water entering the heater. T 2 = temperature of the water leaving the heater. T s = temperature of the exhaust steam. A = square feet of transmitting surface. T = temperature of a unit of water t seconds after entering the heater. h = B.T.U. absorbed per square foot per second per degree difference in temperature between the steam tempera- ture T s and the water temperature T. t = time in seconds. w = number of pounds of feed water per second. Then — = square feet of surface brought in contact with one pound of w water per second, and dT, the rate at which the temperature of the water is increasing at this instant, will be dT= — (T 8 -T)dt. (108) w Integrating, pjL = Mf 4 (111) J T Ts-T w J loge TS ~ T ° = — • (112) Let W = number of pounds of feed water heated per hour. U = B.T.U. transmitted to the feed water per square foot of sur- face per hour per degree difference in temperature. Then (112) may be written from which x ^¥^rw' (113) A =v^T^k- (114) FEED-WATER PURIFIERS AND HEATERS 497 Knowing the weight of water to be heated, the temperature of the s^eam, the desired temperature of the feed water, and the coefficient of heat transmission, U, this equation enables one to determine the area of heating surface required for the given conditions. Since the extent of heating surface increases rapidly as T 2 approaches T s , and becomes infinity for T 2 = T s , it is desirable to limit T 2 to some practical figure. An average maximum for T 2 = T s — 4. Table 64 has been calculated from this formula and gives the square feet of heating surface necessary to heat 1000 pounds of water per hour for different ranges in temperature. Mean Temperature Difference. If we let d = average temperature difference between the steam and feed water, then AUd = heat given out by the steam per hour. W (T 2 — T ) = heat absorbed by the feed water per hour. AUd = W(T 2 -T ). (115) (116) From (113), ^- = log, ^~ ^ . (117) Therefore d = T 2~ T ° . (118) Table 65 has been calculated from formula (118) and gives the mean temperature difference for various conditions of operation. The arithmetic mean temperature difference d t may be taken with safety for the average heater problem and has the advantage of sim- plicity. d t = T s ~ T «\ T * - (119) Closed heaters are sometimes rated on the basis of \ square foot of heating surface per horse power, i.e., a heater with 500 square feet of heating surface would be rated at 1500 horse power. 259. Heat Transmission in Closed Heaters. — Table 66 gives the results of a series of tests on the absorption of heat by water passing through brass and copper tubes surrounded by steam. The curves in Fig. 248 were plotted from the data given in this table. An inspection of the table and the curves will show that the absorption of heat per square foot of surface per degree difference in temperature varies with d = W(T 9 - T ) AU AU W T — = iog. r ;_ T„ T 2 d ■- r 2 -7 iog. r ;_ ■T. 498 STEAM POWER PLANT ENGINEERING < H W H II P a K n o i p o H o to n oo a> O O O o »o o »o o CM CM CO CO S3 el 03 w o "S .d a ai O a o CN CM p. S CO 1 a 8 a < 0) 1 O ts 1 a "3 d s © i-H CM •«* ^H O lO ITS OS t>. lO CO f— I 00 t- CN CN CN CM i-H i-4 o CO 1—1 *•- o •^ O 00 OS CO ifl o o o 00 o CN t-4 O0 lO -Ct< O t^ »0 CO i-H OS OO O O O O OS OS CM OS CO OS CO CM OS 00 o OS lO O to OS 00 CO oo oo oo o CM CO 00 CO CO o o t^ »o co cn o oo 00 00 00 OO GO t*» CO CM CM O0 o co o> r~ O) N lO CO CO CO O o CM CO ^ lb H N N »0 "^ i-H o oo «o r-- t~ t» r>- co co CO CO co CM CM CO CO CO OS lO OS "* ^ N CO ■* to lO lO CO OS »0 O0 ""* CN i-l OS CO »0 ■<*< CM O t^ co co co co co >o lO co ->* 00 CM U3 lO CO OS lO OS t^ CO CN OS i-4 t}4 OS 00 t^ lO O OS t~- »0 CO i-H co »o *a io *o »o CO OS OS CO o CO lO CO CO .-I oo 1 d o O "So CM i— i »o OS CO CO O CO CM 00 CO NO«OiO»0 O O o 1 00 CM S3 1 03 1 O I "cl K s "3 d £ o OS O0 1>- CO TJ4 CO o ia os i— i o: ifl^COMii O CN i-4 O0COMNH lO CM OS lO O o 00 T»4 t» CO TJ4 HiOOSH CO CM i-H t-i »o >O©H00N CD CO O ^O O <^4 ^4 r*4 CO CO o CO CO GO OS -^ t^ HH o CClOONCO OS CO CM O CO CO CO CO CO CM o ' «# i-H CJ c3 > CM o r-4 00 OO O OS OKJOiON CO lO lO TJ4 CO o "«* I-l 03 '3 H S3 s 11) e I > ft: ■# o CO OS i-H »C OS CO CM O CO CM 00 CO CO iO »o ^ o o CM CO CO CM i-H t~ CO 00 CM CM ^4 ^4 CO CO CM o OS CM OS OS i-H CO CM t^ CM t^ OO CO CM CM i-I CM CM CO O0 CM CM CM OS lO CM I>» iO "^ ^ ^ CO o O lO CO CO i-H iO H CO \H o "^Cf **f CO C ") CO O ' »o CN II Si > "cO CM o O0 CO i-H CO oo i-i t^ ia oo cm oo i>- r~ eo co U5 CO t~- CM CO CO OOiOCNOO-* CO CO CO CM CM LO t^ CM CO OS OO ONCONN OiOlO>*^ O CO OS CO o t^ CO CM OS lO O CO CO CM CM CM o MNlftiON OS lO i-i CO c Tj4 ^4 t}4 CO CO iO o CO -^4 OS CO o OS CO CM OS »o lO o co cm co co a-. "& OS CO 00 CM Tt< CO CO CM CM 1 "e8 "cl l| ►H g o 03 H O O O O O if o ooc TJ4 »0 CO l> c cc FEED-WATER PURIFIERS AND HEATERS 499 Initial Temperature of Water T . o o o o •* lo co t^ o O OS o o LO o o »o 1—1 1— ( o »o o CN CM CO i a o 1 p. CO O a < CN cn d E S g CO - •* CM O QO ^H "^H ^< CO CD CO ro CO CN co co CM CM O OS CO (M O OS 00 OO CO LO CM CM CM o (N lo CM © CO CO r-H 00 CO K) lO ^ ^ CM 00 CO OS co CM 00 CO OS t^ CO LO CO CO ■^ CO T-l CO CO CO o o i-i CO CM CO O t» to • CO CO CO lO OS <* LO co CM LO CO OS OS CO LO LO OS CO CO i—l o CN o> CO OS i— I CM o r- lo cm t— CO CO CO OS LO CO LO 1— 1 eo LO OS 1— ( LO CO 00 O oo LO "<* CM CO OS N K5 M 1 3 CO -a 3 6 s s 'I m 3 I § 1 s 1 eo ft S a 3 3 § > "So CN (J co "el 3 el co ft a 3 s lo CN CO y-t "tf CO O0 ■«4< CM OS CO CO CO CO CM CM CM o ' o o ft a CO H d > ' 00 CM 13 el o ft a "3 3 o OS OS 00 CO i-H ■* CM CM CM i-H r-i o IN CO t- t^ CM "* ON^HOO ■* CO CO CO CM o 00 CO l~~ oo t— CO CM OO "^ CO CO CM CM >o LO CM OS lO CM ^ ^ CO CO CO o 1 NihN • COOS'* • "* CO CO • o HNNNOS O CO CO OS lO LO ^ ^ CO CO O " ■<* ft a ci CM cu ft a £ "3 3 LO o 00 O T-H i-t t^ O 00 LO CM 00 CO CM CM CM i-H fa o CO S- 3 "c! © ft s a 3 3 O > CN CO el o CO | ft a 3 E o eo ©OOHHN ONiONO) •* CO CO CO CM o o O 00 CO CM LO CO CM OS CO CM CO CO CM CM CM o OS *00©ON -* O CD CO 00 -* -<* CO CO CM to CM i-H i-i OS »0 CO CO O CO CO Tt< ■* i*l CO CO o CN LO ^ ^ ^t* CO o LO CN ft a 1 CN a; "cl o ft a "3 s o CN CO OS Tf t>- i— 1 00 LO CO © OO CM CM CM CM i-H LO CO OS -* 00 o lO-HCO-^rH iO »0 ""* "^ "^ LO © CO ^ © CO LO CM © CO CO CO CO CM CM CM o CO CO i-H (M CM OS LO CM 00 -* lO LO to •* ■* o rt< eo i—i oo co © t^ ■* © i-~ Tt< CO CO CO CM lO o OS CM lO CO CO N OJ lO rH N LO o © OS CM CO 00 (M © O0 ■* © ^ CO CO CO CO CO 3 "eS "cl 'If ca o o o © o •<* »o co r- oo ' 3 3 tl IP •-T a; © © © © © ■«*< LO CO t~- QO 500 STEAM POWER PLANT ENGINEERING the velocity of the water and the material and character of the tubes. Increasing the velocity of the water passing through the heater in- creases the rate of heat transmission and thereby renders the heating surface more effective. In order to employ moderately high velocities and at the same time allow suffi- cient time in which to raise the temperature to a maximum, the tubes should be as long as prac- ticable and of small diameters. Other things being equal, a heater containing a large number of tubes of small diameter is more economical than one containing a small number of large tubes. It is important to proportion the heater according to the amount of water to be heated and the maximum temperature to which the water must be raised. In designing a heater, then, the maximum amount of heat to be transmitted per degree difference in temperature per hour per square foot should be assumed, and the velocity of the water made such that it is capable of absorbing this amount. A good average figure for multi- "*5 50 75 ioo 125 150 175 fl ow heaters is U= 250 B.T.U. for plain brass or copper tubes and [/= 300 B.T.U. for corrugated tubes with a water velocity of 50 feet per minute; for single-flow heaters, U= 175 (for plain brass) with a water velocity of 12.5 feet per minute and for coil heaters U= 300 (copper) with a water velocity of 150 feet per minute. These figures are for water-tube heaters only. For steam-tube heaters (iron tubes) a good average figure is U= 120. Experiments show that heaters and condensers operating with counter-currents are more efficient and are capable of obtaining a higher final temperature than those operating with parallel currents. Example: Determine the size of vacuum and atmospheric heaters for a condensing plant of 1200 I.H.P. Engines use 20 pounds of steam per I.H.P. hour; auxiliaries use the equivalent of 10 per cent of the main engine steam; vacuum 25 inches referred to 30-inch 3 INDICATE PLAIN TUBES CORRUGATED TUBES FEED-WATER PURIFIERS AND HEATERS 501 barometer; feed water, T = 50 degrees; temperature of hot well, T 2 = 110 degrees; coefficient of heat transmission, U= 300 B.T.U. Vacuum or Primary Heater. Feed water for main engines, 20 X 1200 = 24,000 pounds per hour. Feed water used by auxiliaries, 10 per cent of 24,000 = 2400 pounds per hour. Total feed, W= 24,000 + 2400 = 26,400 pounds per hour. From formula (114), W T S -T Q A = — - log e - U -=• T s -T 2 26,400 lo 134 - 50 300 ° 134-110 = 110 square feet. On the basis of £ square foot of surface per horse power the rating of this heater will be 110 X 3 = 330 horse power. Atmospheric or Secondary Heater. The temperature of the feed water leaving the atmospheric heater, formula (105), will be , _ t + 0-9 S (A + 32) % ~ 1+0.9S where S = .10, t = 110 degrees, X = 1146 B.T.U. whence t = 110 + 0.9x0-10(1146 + 32) 1 + 0.9 X 0.10 = 198 degrees. The required surface is U °°T S -T 2 where T 8 = 212, T = 110, T 2 = 198, whence A = &™ log e 212 ~ 110 300 6 212 - 198 = 175 square feet. The horse-power rating will be 175 X 3 = 525 horse power. 502 STEAM POWER PLANT ENGINEERING •J9^naraiJ8dxa flfififlflflCCann E iB«io!i2£SSS£S2S 00000000~~^~'rr'-r;3:33:3:3:3:3:33 c c a c a a fl •anon jad "duiax JO 'Jia 89.129(1 Jsa aoi3jjngjo-ijj-bgi9d paqaosqv s.nxg •jnoH J9d paqjosqy s.-fl-j/a jo jaquiriM i^px •ajmiijn jad ^9aj ut aaj-BAi jo A'^ioopA •aqnx 9]Suig xdd anon J9d J8j^^ J° spunoj OS ri^cuCN 'HNrtMMCCOONlOlONO'HOQO'taNrHtN.QOtON QO 1-H o oo o «o «o t~- oo o 10WNNO>0 t— !>• OO i— i o t~- CO CO i-H OS »o oo t~ uo »o 05CON OSi— I N o m o ^ ^ »- imooo^no«ooooooooo MiOffiiOOOffiCa^OiHCOOflNH^rtONO NOlOSiONiOiOOO^COOOOONWlOaiWi-i^tio (NlOON'*(ONiOi-i(N'*rH^MiOiHrHlOCONO rtHHHHHlOOOONHCO^^NOOOOtniH HHHHMNMCONCOit OO OONNOOOONNOOmHOOOCONION^NMOM — I CO i-H !0»ON CO' 0010NOOOU5N50 CO lOCOCOlOlOCOCOiOCMi-HCOlOOcMOOi-iOOCMOOcNcMOO"^ NiO«OONlO(OOON05NNHONNOOrtiM'*lNNN(N'HN«e«C'Jt 1-Hr^i— It-It-Hi— It— It— llOlOlOlOlOUOi-Hi-Hi-Hi— li— It— It— ll— ll—t 1139H jo 199J 9.renbg COcO-t-i+j+J+J-U>+J+i+i-t^+i+s. ( _} J- (-4 t-l -j_D_Q_Q f - |t - |t - |t - |t - p/H/cl 0,"3 T3-OT3T3T3-OT3l373T3T3'r5T3T3T3'C3- £: l' £ 3^ r-H l>- (N 00 (M r- <-i 0(X)000 , *»0(»rHNOJ'*'t'0H 00-^00 OO""^ cm oo ONniN^coioaooHHntqooaoioo(0 rttiOCNOiOOOOON^iOiO^OOGOiO^OOONOOOH fflCiO^^iOfOtflM^OOiCCOH^lCKOOOOSOJOtOaiN 0>00CDHMC0©©O^tOtO(MCO'*tO'<^ (MfMCO^TtHTtitOrJHeOC^ flNlO©N«OCOOCOrHlOTl>'-i'-t- ON^»0(ONNffl<0»HiO©'* 0(OOOM>ONh(X)OONO>'5'*>ONN(»NM05hh<0 COO^ , M05©MtO>OiO©'OOH050©NOO(Nffl'HncO ©O^iONaN-HlOOOaoONINOO^OOmHQOrtlN© OO :M ^'JllOONHMTtlOOCHlONHHOQOlOnOOOHiON OO05OOO0505OOO050ii-(i-iT-HOOOO05'-i<-iCS| ,-,,-HtOi— itOiOtOtOlOtOtOtOlOtOtOlOtOlOtOtOi00505CO rH^i-Hr-Hi— It-Hi-Hi-Ht-It-Hi—Ii— li— li— (i— It-It-Hi-H,— lr-li— li— (CX1CO (N(NN(NH|« HN iH|C< H*» H« H« HN i-(|N HN H(e« H^ i^N (M 2000 Fuel and Fixed Charges. Fuel, 2988 tons at $2.50 $7,470.00 Fixed charges, 16 per cent of $12,250 1,960.00 $9,430.00 In like manner Cases V, VI, VII and VIII have been treated and are tabulated in the summaries. SUMMARY (1). I.... II... III.. IV.. V... VI.. VII. VIII Temperature of Feed Water. Degrees F. 152 198 179 217 208 294 290 270 Power Consumed by- Efficiency. First Cost. Fuel Cost per Year. Auxiliaries. Per Cent. Per Cent. 10 10.8 $5,080 $8,308 10 11.3 5,080 8,120 3 11.8 11,100 7,595 5 12.05 12,250 7,470 10 11.4 5,280 7,900 14 12 9,000 7,750 10 12.2 9,300 7,380 8 12.3 8,250 7,075 Cost of Operation per Year. $9,120 8,932 9,371 9,430 8,744 9,190 9,570 8,395 FEED-WATER PURIFIERS AND HEATERS SUMMARY (2). 521 Case. Efficiency. First Cost. Fuel. Cost per Year. I 8 1 8 4 II 7 1 7 2 Ill 6 6 4 6 IV 3 5 7 2 3 6 7 V 3 VI 4 4 5 5 VII 2 5 2 8 VIII 1 3 1 1 Summary (2) gives the ranking; thus: Case I is eighth in point of efficiency; first in cheapness of installation; eighth in yearly cost of fuel; and fourth in yearly cost of operation. Case VIII is apparently the best arrangement for the given conditions. CHAPTER XIII. PUMPS. 269. Classification. — Pumps used in connection with steam power plants may be conveniently classified under five groups according to the principles of action. 1. Piston pumps, in which motion and pressure are imparted to the fluid by a reciprocating piston, plunger, or bucket. The action is positive and a certain definite amount of fluid is handled per stroke under predetermined conditions of pressure and velocity. 2. Centrifugal pumps, in which the fluid is given initial velocity and pressure by a rotating impeller. The action is not positive, as the amount of fluid discharged is not necessarily proportional to the impeller displacement. 3. Rotary pumps, in which motion and pressure are imparted to the fluid by a rotating impeller. The volume discharged is practically equal to the impeller displacement regardless of pressure. 4. Jet pumps, in which velocity and pressure are imparted to the fluid by the momentum of a jet of similar or other fluid. The ordinary steam injector is the best known of this group. 5. Direct-pressure pumps, in which the pressure of one fluid acts directly on the surface of another fluid, thereby imparting all or part of its energy to the latter. The pulsometer is an example of this type. These groups may be variously subdivided as follows: Piston. Direct-acting.. \%^ Forcing. Lifting. ^-heel j^ptf Power dnven •• I Triplex r< , ., , f Volute Single stage Centrifugal . . . . ( Turbine Multi-stage Rotary j Power driven . . j L ^ n g g T , ( T . , j Positive . . . Jet | Injector j Automatic Direct nressure \ P uls O m eter Lifting. . . . Uirect pressure -j Air _ Hft Lifting Piston or plunger pumps are the most common in use. Boiler- feed pumps, city waterworks pumps, and force pumps are ordi- narily of this type. In the direct-acting type, Fig. 260, the water 522 Air. Vacuum. Forcing. Lifting. PUMPS 523 plunger and steam piston are secured to a single piston rod and the steam pressure is transmitted directly to the water. There is no fly- wheel, connecting rod, or crank. The velocity of the delivery is pro- portional to the resistance offered by the water; when the resistance equals the forward effort of the steam pressure the pump stops. This class of pump is well adapted for boiler-feeding purposes, since it may be operated as slowly as suits the requirements of feeding by simply throttling the discharge. The steam consumption is very large in proportion to the work performed, since the steam is not used expansively. Fly-wheel pumps, Figs. 273, 308, are ordinarily classified as pumping engines. In this class steam may be used expansively, as sufficient energy is stored in a fly wheel to permit the drop in steam pressure during expansion. These pumps find wide application in city water- works, elevator plants, and the like, where high duty is required. They are little used as stationary boiler feeders, but are used to some extent in river boat practice and in plants operating continuously for long periods at comparatively steady loads. Practically all sizes of dry-air pumps and a number of large jet condenser pumps are of this type. Piston pumps, Fig. 279, driven by gearing or belting are ordinarily classified as power-driven pumps. The driving power may be steam engine, electric motor, or gas engine. The single-cylinder machine is often designated as a " simplex " power-driven pump, the two-cylinder as a " duplex," the three-cylinder as a " triplex," and so on. Centrifugal pumps, Fig. 292, are supplanting to a considerable extent the present type of piston pump for many uses. Though particularly adapted for low heads and large volumes they are used in many situations requiring extremely high heads. They are not as efficient as high-grade pumping engines, but the extremely low first cost fre- quently offsets this disadvantage, and they are much used in connection with dry docks, irrigating plants, sewage systems, and as circulating pumps in condensing plants. Rotary pumps, Fig. 305, are employed to a limited extent in the same field as the centrifugal pump. Being positive in action, they permit of a much lower rotative speed for the same delivery pressure. Jet pumps, Fig. 282, are seldom used as pumps in the ordinary sense of the word, on account of their extremely low efficiency, but are frequently employed for discharging water from sumps. Their greatest field of application lies in boiler feeding and in this respect their efficiency is comparable with that of the average piston pump. 524 STEAM POWER PLANT ENGINEERING Direct-pressure pumps operated by steam, such as the " pulsometer," Fig. 309, are used principally for pumping out sumps, surface drains, and the like, where the operation is intermittent. Direct-pressure pumps of the air-lift type, Fig. 310, are quite common and are used a great deal in situations where water is to be pumped from a number of scattered wells. 270. Boiier-Feed Pumps, Direct-Acting Duplex. — Figs. 259 and 260 illustrate a typical duplex boiler-feed pump, which consists virtually of AIR CHAMBER DISCHARGE STEAM SUPPLY Fig. 259. Typical Duplex Pump. two direct-acting pumps mounted side by side, the water ends and the steam ends working in parallel between inlet and exhaust pipe. The piston rod of one pump operates the steam valve of the other through the medium of bell cranks and rocker arms. The pistons move alter- nately, and one or the other is always in motion, the flow of water being practically continuous. In general construction the steam pistons and valves are similar to those of steam engines. The valves in duplex pumps, however, have no lap. In order to reduce the valve travel to a minimum, and^till have sufficient bearing surface between the steam ports and the main exhaust ports to prevent the leakage of steam from one to the other, separate exhaust ports are provided which enter the cylinder at nearly the same point as the steam ports. This arrangement offers PUMPS 525 a simple means of cushioning the piston by exhaust steam, thus pre- venting it from striking the cylinder heads at the ends of the stroke. The valves of the duplex pump having no lap would, if connected rigidly to the valve stem, open one port as soon as the other had been closed, at about mid-stroke of the piston, thus cutting down the stroke 01SCHAR4C L per Fig. 260. Section Through a Typical Duplex Boiler-Feed Pump. to about one-fourth the usual length. To obviate this difficulty the valves are given considerable lost motion by allowing sufficient clear- ance between the lock nuts on the valve stem; the latter, therefore, imparts no motion to the valve until the piston operating it has nearly completed the stroke. The lost motion between valves and lock nuts renders it impossible to stop the pump in any position from which it can- valve stem not be started by simply admitting steam, and therefore the pump has no dead centers. When one piston moves to the end of the stroke it pulls or pushes the opposite valve to the end of its travel; then when the piston starts back to the other end of its stroke the valve remains stationary, owing to the lost motion, until the piston has completed about one-half the stroke. J PISTON ROD Fig. 261. 526 STEAM POWER PLANT ENGINEERING VALVC STEM PISTON ROD During this time the opposite piston has completed a full stroke and the valve operated by it will have opened the steam port wide, so that while one valve covers M — jrV, _ - both steam ports the other Qf|:| M ; | I I fi'ltD is at the end of its travel. In some makes of pumps the stem is rigidly attached to the valves, the lost mo- tion being adjusted outside the steam chest as shown in Figs. 261 and 262, which represent two common constructions of duplex valve gear. Fig. 263 shows the valve and piston in the position occupied at the commencement of the stroke. At one end of the valve the steam port P is open wide and at the opposite end the exhaust port E is open wide. When the piston nears the opposite end of the stroke and reaches the posi- tion shown in Fig. 264 the steam escape through the exhaust port E is cut off by the piston, and since the steam port is closed, the remaining steam is compressed between the piston and cylinder head, thus arresting the motion of the piston gradually without shock or jar. The construction of the water end of single-cylinder and du- plex pumps is practically the same; any slight differences which may be found are con- fined to minor details which in no way affect the general design or operation of the pump. The piston is double acting, the single-acting cylinder being confined to power pumps or to steam pumps intended for very high pressures. In the old-style pumps it was the custom to use one large valve with a lift sufficient to give the required passage, but in modern practice the required area is divided among several small Fig. 263. Fig. 264. PUMPS 527 valves, so that each one is easily and cheaply removed in case of accident or wear, and slip is lessened. * The valves are carried by two plates or decks, the suction valves being attached to the lower plate and the delivery valves to the upper one, as shown in Fig. 260. The valves in practically all boiler- feed pumps are of the flat disk type, Fig. 265, held firmly to the seat by conical springs and guided by a bolt through the center. All pumps are provided with an air chamber on the discharge side, which acts as a cushion for the water, pre- vents excessive pounding, and insures Fig. 265. A Typical Pump Disk-Valve. a uniform flow. Fig. 266 shows a section through the steam end of a compound duplex pump. w H £ \ a §1 \ & ^ \ 1/ X X ^ y a 2000 § t ^ fe ^ / <% e y st* 1000 f / /, // k> 20 30 40 50 60 TO Single Strokes per Min. Fig. 275. 5 S 90 100 The determination of the power consumption of a boiler-feed pump is best illustrated by the following example. Example: A small direct-acting duplex pump uses 150 pounds of steam per I.H.P. hour. Gauge pressure 150 pounds per square inch; feed-water temperature 64 degrees F. Required the per cent of rated boiler capacity necessary to operate the pump. The head pumped against, 150 pounds per square inch, is equivalent to 150 X 2.3 = 345 feet of water. PUMPS 535 The friction through the valves, fittings, and pipe, and the vertical distance between suction and feed-water inlet, are assumed to be equiva- lent to 20 per cent of the boiler pressure, giving a total head of 150 + 30 = 180 pounds per square inch, or 414 feet of water. A boiler horse power, taking into consideration leakage losses and the steam used by the feed pump, will be equivalent to the evapora- tion of approximately 32 pounds of water per hour from a feed tem- perature of 64 degrees F. to steam at 150 pounds gauge. The actual work done in pumping 32 pounds of water against a head of 414 feet is 414 X 32 = 13,248 foot-pounds. This corresponds to 13,248 = 0.0067 horse power. 60 X 33,000 The total heat of one pound of steam above 64 degrees F. is 1161.2 B.T.U. The heat delivered to the pump per I.H.P. hour is 1161.2 X 150 = 174,180 B.T.U. The amount used by the pump for each boiler horse power, disregard- ing efficiency, is 174,180 X 0.0067 = 1167 B.T.U. per hour. The mechanical efficiency of the average feed pump ranges from 50 to 85 per cent, depending upon its condition and the number of strokes per minute. Assuming it to be 65 per cent, the heat used by the pump per hour to deliver 32 pounds of water into the boiler is 1167 4- 0.65 = 1795 B.T.U. A boiler horse power is equivalent to 33,320 B.T.U. per hour. There- fore the per cent of boiler output necessary to operate the pump is 100X 3lS= 5 - 4perCent - If the exhaust steam is used for heating the feed water, the steam con- sumption will be 1.37 per cent of the boiler capacity, thus: The weight of steam consumed per boiler horse-power hour 1795 1161.2 = 1.54 pounds. Allowing 10 per cent for condensation, the heat in the exhaust avail- able for heating the feed water is 966 X 0.90 X 1.54 = 1340 B.T.U.* * Surface Condenser Plant. 536 STEAM POWER PLANT ENGINEERING 1795 — 1340 = 455 B.T.U., or the net heat required by the pump per hour to deliver 32 pounds of water to the boiler. The per cent of boiler output necessary to operate the pump is 100 455 = 1.37. 33,320 Pump performances are generally given in terms of the foot-pounds of work done by the water piston per thousand pounds of dry steam or per million B.T.U. consumed by the engine, thus: 1. Duty - Foot-pounds of work done x 1(m Weight of dry steam used 2. Duty= , Foot-pounds of work jong x 1;00 0,000. Total number of heat units consumed (See A.S.M.E. code for conducting duty trials of pumping engines, Trans. A.S.M.E., 12-530, 563.) Example: A compound feed pump uses 100 pounds of steam per I.H.P. hour; indicated horse power, 48; capacity, 400 gallons per minute; temperature of water, 200 degrees F.; total head pumped against, 175 pounds per square inch; steam pressure, 100 pounds gauge; moisture in the steam, 3 per cent. Required the duty on the dry steam and on the heat-unit basis. 175 pounds per square inch is equivalent to 175 X 2.4 = 420 feet of water at 200 degrees F. Weight of 400 gallons of water at 200 degrees F. = 400 X 8.03 = 3212 pounds. Work done per minute = 3212 X 420 = 1,349,040 foot-pounds. Weight of dry steam supplied per minute = X 0.97 = 77.6 pounds. B.T.U. supplied per minute = 10 ° * 48 (0.97 X 876.2 + 308.8-200 + 32) = 79,256. Duty per thousand pounds of dry steam = 1 > 349 » 040 x 1000 = 17,384,150 foot-pounds. 77.6 Duty per million B.T.U. = i; 349 ? 040 x 1000,000 = 16,893,863 foot-pounds. 79,256 ' ' F Table 68 may be used in approximating the duty, thus : The mechanical efficiency of the pump in the preceding problem is PUMPS 537 o o iONOnOiON0c©cot^oooi p H pq d iO^W5rHH©cOOOOOOOi©HOOiOOO(NO ^NO^CCNNMOOOO^COlOrHOJCNCC ©OOt^O—iiOOO© (OiOOOOONniOOO lOlOCDCOCONNQOOJOJO'HfOOOO'HNCO HHrtHHCqiNCO oio'O'HcoocONOca CO^iOCOCONNOOOlOl o 1 i X d s I -g HI co S o o o CO CO X a !*> o s a; o fa o CO o TfNOCSNNOliOOOOO^iOCOOcOOO ffiMcoos^os^rHCiceQoiocjooiceNNco ooooo**t^©co© (MOOtOOi-iOOi— IO H fin oq Q P ^H^H^I^^H^CSJC^CO r)*lOCOCONNCX)0>0)0 CD o 'tOJUlN^ffiOOHOOOOOOOOMCONtO ^NuiOO^CqOlNNCCHOO^^Ni-iOO OOOfNOt^OOOO OO^OOi— IQIOOIOCOO C05ONN0000O05OHCN«H000pH10(NCN HHHHHHCN(NC<3'* COOOIOMOOMO'NCON ">ct<»ocot~-t^c»05a50'-H O o o MHOiOtOO)00)>00.ooo cOOOOOCOI^OOiO p n N00 00 0505 0'H(NCO'*iOCOa)(NCO'-iC7lCN NNOOCNOOO^ 00 o (NCOOOO'-ilNMtOMT-iCNCOeO'tiO^NO <#OOCOO)iONOOiOnMNOOO(DOH OOONOOlr- lOOO CO (ONOOCBOOwlNMiO o OS o t-h0500COO»0'-i>0-^hC<)OI>-00'<^>00 05COOHOiH05NNOO(NOOOO(NlONCOiO^ 000»00'-H'-iOOO OOOOCMOt^cOOCOO 3 O'H'-H(NC0-*t0NO5 = °- 77 VlI^- < 133) ^D ^ + fi + Q ll, (134) / = w (p + R + °- 433 H ) 2 - 3 nw 33,000 X 60 X E K In average practice the piston or plunger displacement is made about twice the capacity found by calculation from the amount of water required for the engine, to allow for leakage, steam consumption of the auxiliaries, blowing off, and pump slip. For pumps with strokes of 12 inches or over, the speed of the plunger or piston is usually limited to 100 feet per minute as a maximum to insure smooth running. "For shorter strokes a lower limit should be used. The maximum number of strokes ranges from 100 for strokes over 12 inches in length, to 200 for strokes under 5 inches. Boiler-feed pumps should be designed to give the desired capacity at about one-half the maximum number of strokes or less. Pump slip varies from 5 to 40 per cent, depending upon the condition of the piston and valves and the number of strokes. An average value for piston and plunger pumps in first-class condition is 8 per cent when 540 STEAM POWER PLANT ENGINEERING operating at rated capacity, but it is wise to allow a much larger figure, say 20 per cent, for leakage caused by wear. The area of the steam cylinder is made from 2 to 2.5 times that of the water end to allow for the various friction losses and the drop in pressure between the pump throttle and the boiler. The total head pumped against includes the suction lift, the friction of valves and fittings, the distance between the suction inlet and the boiler level, and the boiler pressure. The excess head varies in practice from 15 to 40 per cent of the boiler pressure; an average figure is 25 per cent. In allowing for the drop in steam pressure between boiler and pump a liberal figure is 25 per cent. The application of formulas (132) to (135), including the practical considerations stated above, is best illustrated by a specific example. Example: Determine the size of direct-acting single-cylinder feed pump necessary to supply water to 1000 horse power of boilers. Gauge pressure 100 pounds per square inch; feed-water temperature 150 degrees F. One horse power is equivalent to the evaporation of 34.5 pounds of water from and at 212 degrees F.; but the pump is usually designed to supply about twice the capacity. Thus W = 62,400 (under the given conditions). S = 0.8 (by assumption). LN = 1200 (on the basis of 100 feet per minute). Substitute these values in (133) : D = 0.77 J 62,4Q( lo = 6 - 2 inches, — call it 6 inches, V 1200 X 0.8 since the assumptions have been very liberal. Assume (.433 H + R) = 0.25 p and E = 0.65. Substitute these values in (134) : d = *\ll 100 + 25 0.65 X 100 = 8.35, — call it 8.5 inches. Allowing 100 strokes per minute the length of the stroke must be L = 1200 -^ 100 = 12 inches. The dimensions of the pump are 8J x 6 x 12. The indicated horse power at maximum load may be obtained by substituting the proper values in (135), thus: j m 62,400 (100 + 25) 2.3 33,000 X 60 X 0.65 = 13.9 I.H.P. PUMPS 541 276. Steam-Pump Governors. — Fig. 276 shows a section through a Fisher pump governor, illustrating a device for maintaining a practically constant pressure in the discharge pipe irrespective of the quantity of water flowing. It embodies a pressure-reducing valve in the steam e supply pipe of the pump, actuated by the slight variations in water pressure. When the demand for water increases, the pressure in the discharge pipe tends to decrease, and this drop in pressure (transmitted to the pump governor by suitable piping) causes more steam to be admitted, which increases the speed of the pump. The governor is connected to the steam inlet of the pump at B and the steam enters at A. Double-balanced valve C regulates the supply of steam to the cylinder by the amount it is raised from the seat. The valve is held open by spring G, the com- pression of which may be regulated by hand wheel K. The water pressure from the discharge pipe acts on piston F and tends to overcome the resist- ance of the spring. The difference in pressure between the water and the spring determines the position of valve C. Piston rod H is pinned to sleeve I and valve stem L screwed into this sleeve by means of hand wheel K. Hence during ordinary operation the piston, piston rod sleeve, valve stem, and valve act as a single unit. By turning the hand wheel K, valve stem L will screw into sleeve / and the tension on the spring will be increased. Hand wheel J serves as a lock nut and prevents K from turning during normal operation. 277. Feed- Water Regulators. — The water level in the boiler should be kept as nearly constant as possible, and this necessitates considerable attention on the part of the fireman, especially with fluctuating loads. There are a number of devices on the market which are designed to automatically maintain a constant level, and in many small plants where the duties of the fireman are numerous such devices in connec- tion with high and low water alarms are of considerable assistance. Their action, however, is not always positive on account of wear or sticking of parts, and engineers as a rule prefer to rely upon hand regula- tion. In large stations regulators are seldom used. Fig. 277 shows a section through a Kitts feed-water regulator, con- sisting of two parts, the chamber F and the regulating valve V. The Fig. 276. Fisher Pump Governor. 542 STEAM POWER PLANT ENGINEERING float chamber is connected to the boiler or water column at and E, and the regulating valve to the feed main at R and to the boiler feed pipe at W. When the water in the boiler falls below the mean level, the weight B overcomes the counterweight G and closes needle valve L by means of compound levers. At the same time an extension on valve L lifts spring A and opens exhaust valve D. This removes the steam Fig. 277. Kitts Feed-Water Regulator. Fig. 278. Rowe Feed-Water Regulator. pressure from the top of diaphragm C, in the regulating valve, through the agency of pipe K. The pressure from the pump raises the disk T and water flows into the boiler until the water rises to the mean level. When weight B becomes submerged its weight is overcome by counter- weight G, valve L is opened and exhaust valve D is closed. This admits steam pressure to the diaphragm C and forces disk T to its seat, cutting off the supply of water to the boiler. The Rowe feed-water regulator, Fig. 278, depends for its operation on a familiar float-controlled valve mechanism. The vessel A is con- nected to the boiler above and below the water line, and the float C, following the water level up and down, actuates a balanced valve in accordance with the boiler-feed requirements. When this apparatus is used to regulate the feed of a single boiler the opening G in the valve chamber is connected to the steam space of the boiler and the outlet H PUMPS 543 is carried to the steam inlet of the feed- water pump. When the water level is normal the float closes the valve L and thereby cuts off the supply of steam to the pump cylinders. Communication between chambers A and R is prevented by means of a diaphragm M. When the water level falls below normal the float pulls the valve down, open- ing the way for steam to pass from the inlet G to the outlet H and thence to the pump. When the regulator is used to control a battery of boilers the pump discharge delivers into the inlet G and the water Fig. 279. A Typical Geared Triplex Pump. passes through H to the boiler-feed main. Should the water level fall beyond a predetermined limit by reason of any accidental discontinuance of the water supply which the apparatus cannot correct, the float would open the valve F of the alarm whistle mounted on the top of the main vessel. 278. Power Pumps. — Piston pumps, geared, belted, or direct con- nected to electric motors, gas engines, and water motors, are used chiefly where steam power is not available. Their general utility is evidenced by the rapidly increasing number installed in situations formerly occupied by the direct-acting steam pump. The efficiency of 544 STEAM POWER PLANT ENGINEERING this type of pump depends in a large measure upon the character of the driving motor and the efficiency of the transmitting mechanism. High-speed power pumps direct connected to electric motors give Knowles High Speed Electric Pump Direct connected to M P 6-100 H.P.-280-220 V.Form L Load and Efficiency 1 1 I Noz de Hprizoi ttal 1 / / ' / / 1 / 4& G all on sPer Minu te -2C o- / / / 1 -260- 1 & v/ A J 10- -ic -1 )0- 3 -200- Pl / -9 /— 1 / >0- 00 -180- MOtC r / Q) -8 o s / S ^ / a 60- 0> / V / -u I 3 -7 c // / / 3 a '0- 40- w / '/ / / s 3 -6 // / K / ^' 20- / / / N V bj -5 / / / y- » 00- // / [3 L > y o -4 u / / V % / / / / y s 80 80 -3 / / / s\ 60- '/ / 1 -2 / 10 u / y / / y 20- J } 100 200 300 400 500 600 i 1 I 1 | 1 Gauge Pressure at Valve (Lb.) Fig. 279a. efficiencies from line to water horse power as high as 83 per cent, while the low-speed geared type seldom exceed 70 per cent. The curves in Fig. 279a give the performance of a direct-connected triplex pump, PUMPS 545 and those in Fig. 280 the performance of a triplex pump geared to an electric motor. Both of these performances are exceptionally good and are considerably above the average. Power Pumps: Eng. Mag., Jan., 1905, p. 616; Col. Guard, Nov. 17, 1905; Elec. Rev., Jan. 10, 1902; Elec, N.Y., Oct. 12, 1904; Elec. World, Oct. 14, 1905, p. 667; Engr., Lond., Oct. 17, 1902, p. 377; Engineering, Sept. 1, 1905, p. 275; Engr. U.S., Jan. 1, 1904, p. 47. 1UU Motor _ Pump md Gearin 0> Set Mechanical Efficienc Per Cent 8 8 Head Cons ant, 350 Ft. 200 400 600 800 1000 Discharge, Gallons Per Minute 1200 1400 Head Constant. Speed Variable. 100 I P O 50 Motor puroP_3 id Gearing Set y 10 x 2 Triplex P 5 H.P. Mote imp r Mc C Capacit tor B.P.M. (earing 10 to y 1250 Gal. p 150 jr Min. 25 50 75 100 125 Total Head, Lb. Per Sq. In. Gauge Speed Constant. Head Variable. 150 .175 Fig. 280. Performance of a 65-Horse-Power, Motor-Driven Triplex Pump. Geared Type. 279. Injectors. — As a boiler feeder the injector is an efficient and convenient device, cheap and compact, with no moving parts, delivers hot water to the boiler without preheating, and has no exhaust 546 STEAM POWER PLANT ENGINEERING steam to be disposed of. Its adoption in locomotives is practically universal, but in stationary practice it is limited to small boilers or single boilers or as a reserve feeder in connection with pumps. The objections to an injector are its inability to handle hot water, the difficulty of maintaining a continuous flow under extreme variation of load, and the uncertainty of operation under certain conditions. Fig. 281 illustrates the simplest form of single-tube injector. Boiler steam STEAM SUPPLY Fig. 281. Elementary Steam Injector. is admitted at A and, flowing through nozzle and combining tube to the atmosphere through G, partially exhausts the air from pipe B, thereby causing the water to rise until it comes in contact with the steam. Overflow Fig. 282. Hancock Double-Tube Injector. Fig. 283. Penberthy Automatic Injector. The steam emerging from nozzle C at high velocity condenses on meeting the water and imparts considerable momentum to it. The energy in PUMPS 547 the rapidly moving mass is sufficient to carry it across opening 0, lift check H from its seat and force it into the boiler. The steam then ceases to escape at G. 280. Positive Injectors. — Fig. 282 shows a section through a Han- cock injector, illustrating the principles of the double-tube positive type. Its operation is as follows: Overflow valves D and F are opened and steam is admitted, which at first passes freely through the overflow to the atmosphere and in so doing exhausts the air from the suction pipe. This causes the feed water to rise until it meets the jet of steam and the two are forced through the overflow. As soon as water appears at the overflow, valve D is closed, valve C partially opened, and valve F closed. This admits steam through the forcing jet W and, the overflow valves being closed, the water is fed into the boiler. In case the action is interrupted for any reason it is necessary to restart it by hand. The chief advantage of the double-tube positive type lies in its ability to lift water to a greater height and to handle hotter water than the single-tube. Its range in pressure is also greater, that is, it will start with a lower steam pressure and discharge against a higher back pressure. Double-tube injectors are used almost exclusively in locomotive work. 281. Automatic Injectors. — Fig. 283 shows a section through the Penberthy injector. Its operation is as follows: Steam enters at the top connection and blows through suction tube c into the combining tube d and into chamber g, from which it passes through overflow valve n to the overflow m. When water is drawn in from the suction intake and begins to discharge at the overflow, the resulting condensation of the steam creates a partial vacuum above the movable ring h and the latter is forced against the end of tube c, cutting off the direct flow of water to the overflow. The water then passes into the boiler. Spill holes i, i, i are for the purpose of relieving the excess of water until communication with the boiler has been established. The action of opening and closing the overflow is entirely automatic. Where the conditions are not too extreme the automatic injector is to be preferred for stationary work because of its restarting features. It is also used on traction, logging, and road engines, where its certainty of action and special adaptability render it invaluable for the rough work to which such machines are subjected. Injectors, Theory of: Trans. A.S.M.E., 10-339; Sibley Jour., Dec, 1897, p. 101; Power, May, 1901, p. 23; Thermodynamics of the Steam Engine, Peabody, Chap. IX; Theory of the Steam Injector, Kneass. Injectors, General Description: Engr. U.S., Oct. 1, 1907, Nov. 15, 1907, July 15, 1904, p. 501, Feb. 2, 1903, p. 151; Power, Aug., 1906, p. 478; Engr., Lond., March 10, 1905, p. 244; Engineering, Aug. 30, 1895, p. 281. 548 STEAM POWER PLANT ENGINEERING w = (136) 383. Performance of Injectors. — The performance of an injector may be very closely determined from the equation xr + q - t + 32 (Kneass, " Theory of the t-t Injector/' p. 83). in which w = pounds of water delivered per pound of steam supplied. x = quality of the steam supplied. r = heat of vaporization. q = heat of the liquid. t — temperature of the discharge water. t = temperature of the suction water. Figs. 284a, 284b, and 284c give the performance of a Desmond auto- matic injector as tested at the Armour Institute of Technology. The results check very closely with those calculated from above equation. Referring to Fig. 284a it will be seen that the weight of water delivered per pound of steam decreases as the initial pressure is increased, all other factors remaining the same. From Fig. 284b it will be noted that the weight of water delivered per pound of steam decreases as the temperature of suction supply is increased up to a point where the injector " breaks " or becomes inoperative. This critical temperature varies with the different types of injectors, being highest for the double-tube type, but seldom exceeds 160 degrees F. Fig. 284c shows that the weight of water delivered per pound of steam is practically constant for all discharge pressures within the limits of the apparatus. Table 71 gives the range of working steam pressures for standard " Metropolitan " injectors with varying suction heads and temper- atures, and, though strictly applicable to this particular type only, is characteristic of all makes. In selecting an injector the following information is desirable for best results: 1. The lowest and highest steam pressure carried. 2. The temperature of the water supply. 3. The source of water supply, whether the injector is used as a lifter or non-lifter. 4. The general service, such as character of the water used, whether the injector is subject to severe jars, etc. Injectors, Tests of: Eng. News, March 17, 1898, July 16, 1896, p. 39; Locomotive Engineering, May, 1900, p. 204; Power, Oct., 1904, p. 602; Railroad Gazette, Dec. 11, 1896; Thermodynamics of the Steam Engine, Peabody, Chapter IX; Theory of the Injector, Kneass. PUMPS 549 SA 19 18 IS p m. r 15 13 Constant Discharge Pressure 20 Lb. Per Sq.In. Constant Suction Temp. 55 Deg.Fah. CN v () 65 70 75 90 95 Initial Gauge Pressure,Lb.Per Sq.In. Fig. 284a. Performance of an Automatic Injector with Varying Initial Pressure. 20 u 19 la 18 > « ^£ 17 P<3 17 |^« 16 3 15 U Constant Initial Pressure 70 Lb. Per Sq. In. Constant Discharge Pressure 70 Lb. Per Sq. In. ( )^ ( 5*^* C 55 65 75 85 95 105 Temperature of Suction, Deg. Pah. 115 Fig. 284b. Performance of an Automatic Injector with Varying Suction Temperature ■a £ a p<« u ° 2 * 20 19 18 a* 3 17 16 Constant Initial Pressure,70 Lb. Per Sq. In. Constant Suction Temperature, 56 Deg. Pah. c ) - ( ) 1 y 1 i ( J 30 AO 50 60 70 Discharge: Pressure, Lb. PecSq. In. Gauge Fig. 284c. Performance of an Automatic Injector with Varying Discharge Pressure. 550 STEAM POWER PLANT ENGINEERING TABLE 71. RANGE IN WORKING PRESSURES. Standard " Metropolitan " Steam Injectors. Automatic. Suction Temperature, Suction Head, Feet. Degrees F. 2 8 14 20 Under Pressure. Under 60 25 to 150 30 to 130 42 to 110 55 to 85 20 to 160 100 26 to 120 33 to 100 55 to 80 25 to 125 120 26 to 85 140 Double Tube. Suction Temperature, Suction Head, Feet. Degrees F. 2 8 14 20 Under Pressure. Under 60 14 to 250 23 to 220 27 to 175 42 to 135 14 to 250 100 15 to 210 26 to 160 37 to 120 46 to 70 15 to 210 120 20 to 185 30 to 120 42 to 75 20 to 185 140 20 to 120 35 to 70 20 to 120 283. Injector vs. Steam Pump as a Boiler Feeder. — From a purely thermodynamic standpoint the efficiency of an injector is nearly perfect, since the heat drawn from the boiler is returned to the boiler again, less a slight radiation loss. As a pump, however, the injector is very inefficient and requires more fuel for its operation than very wasteful feed pumps. This is best illustrated by an example: An injector of modern construction will deliver say 15 pounds of water to the boiler per pound of steam supplied, with delivery temperature of 150 degrees F. This corresponds to a heat consumption of 71 B.T.U. per pound of water delivered, thus : With initial pressure of 115 pounds absolute, X = 1185. PUMPS 551 Heat in the water delivered to the boiler, 150 - 32 = 118 B.T.U. above 32 degrees F. Heat of 1 pound of steam above a feed temperature of 150 degrees F. 1185 -118 = 1067 B.T.U. Heat required to deliver 1 pound of water to the boiler, 1067 15 71 B.T.U. A simple direct-acting duplex pump consumes say 200 pounds steam per I.H.P. hour. Assume the extreme case where the exhaust steam will not be used for heating the feed water and the latter is fed into the boiler at 60 degrees F. The heat supplied to the pump per I.H.P. hour, 200 J1185 - (60 - 32)} = 231,400 B.T.U. Assuming the low mechanical efficiency of 50 per cent, the heat required to develop one horse power at the water end will be 231,400 + 0.50 = 462,800 B.T.U. per hour. Since the steam pressure is 100 pounds gauge, the equivalent head of water at 60 degrees F. is 2.3 X 100 = 230 feet. Assume the friction in the feed pipe, the resistance of valves, etc., to be 30 per cent of the boiler pressure; the total head pumped against will be 230 + 69 = 299, say 300 feet. 1 horse-power hour = 1,980,000 foot-pounds per hour. 1,980,000 300 = 6600 pounds, that is, 1 horse power at the pump will deliver 6600 pounds of water per hour to the boiler against a head of 300 feet. The heat consumption per pound of water delivered, / ^ X \ r J == ^ — - -»" ■ 200 100 200 300 400 500 Revolutions Per Min. 600 700 Fig. 297. In water-supply systems in which the friction of the piping is a large part of the total head at full delivery, the characteristic shown in Fig. 300 is especially useful. Thus, when the system reduces its demand for water and the frictional head is consequently considerably reduced, 160 L \ m — ■ g J^- 120 GU 80 ^> A\C V \ *fc •6, p* 20 40 60 80 100 120 140 Capacity Fig. 299. Centrifugal Pump Characteristic for Dry-Dock Service. 160 ^100 - Ch im ?.t.e us \o P3 «n S \ — jig i- -N £0 ^ \ 1 / / L / 40 100 120 140 60 80 Capacity Fig. 300. Centrifugal Pump Characteristic for Water Works with Large Friction Head. 566 STEAM POWER PLANT ENGINEERING 6 "3 c i Sm d ^cceoeo T— 1 1— 1 1— 1 o *ftiflOOOiflOiOOOOC»00 OF-lOO^^CfitO^OOOCOrH wcoocctoic^Ttf^corocoeo tH i-H i— 1 fa G 1> ©©©©0©U5i«»0©»Ot^© ©eo^iooi— itOMr- (cocoi— i© T— 1 1—4 o CO 000000>0i0 O «5 W lO o N^NO©NOTOffiMrHCJ5N i— 1 i— 1 fa O OOiOiOifliOiOifliOOOiOO fa o OOOOiOOOOOOiOOOO OiOHOOOOO^MOOfOiON fa CO OOOOiOOOOO>ON>OiO N05NU5MN^(NOtO'*Ci5iH rH0O«O»OT)(C0«lP0WNNNN S- o cc ©©©©iO©©©©t^©©© ■*CO(NthOiOMOOO^COQO PUMPS 567 the pump would automatically adjust itself to the reduced head without change of speed. Figs. 297 to 304 are based upon experiment and show the relationship between speed, head, capacity, efficiency, and power consumption of various types of pumps. Tables 73 and 74 give the capacity, speed, head, and power require- ments for commercial sizes of centrifugal pumps, and may be used as a guide in selecting the size of pump for general service. TABLE 74. DATA PERTAINING TO WORTHINGTON MULTI-STAGE TURBINE PUMPS. c o 1 oi •3 § is 5 — — P-l o 11 * Total Head in Feet, R.P.M., Number of Stages. Diameter of Di charge Pipe, Inches. 100 Ft. 200 Ft. 300 Ft. 400 Ft. S fc ti o . If 2 2 2 2 2 2 1 1 1 1 o . & 1 § P^ tf if §00 3 t-r O) 1 1.5 2 2.5 3 4 5 6 7 10 12 30 45-60 75-100 125-150 200-250 350-450 600-700 800-1000 1500-1800 2500-2800 0.02 0.0395 0.0625 0.095 0.134 0.222 0.297 0.396 0.643 1.00 2000 1500 1300 1200 1100 950 800 750 600 500 1.5 2000 1800 1600 1400 1200 1300 1200 1000 800 3 3 3 3 3 2 2 2 2 2 1500 1300 1200 1000 1150 1000 800 700 4 4 4 4 3 3 3 3 2.5 3 4 5 6 8 10 1050 950 780 670 5 5 5 4 4 4 4 * Horse power based on maximum capacity. Tests of Centrifugal Pumps: Engr. U.S., Oct. 15, 1906, p. 685 ; Eng. and Min. Jour., April 14, 1906, p. 698; Eng. News, June 2, 1904, p. 512; Eng. Rec, July 1, 1905, p. 25, Sept. 29, 1906, p. 352; Iron Age, Sept. 1, 1904, p. 25; Machinery, Nov., 1906, p. 144; Power, Nov., 1906, p. 688; Trans. A.S.M.E., 22-262, 831; Jour. Am. Soc. Naval Engrs., 17-85. 296. Rotary Pumps. — Rotary pumps are often used for circulating cooling water in condenser installations, and give about the same efficiency as centrifugal pumps under similar conditions of operation. For moderate pressure and large volumes they offer the advantage of low rotative speed, thus permitting direct connection to slow-speed steam engines. At high speeds they are noisy, due chiefly to the gearing. They occupy considerably less space than piston pumps of the same capacity, but require more room than the centrifugal type. Fig. 305 shows a section through a two-lobe cycloidal pump. The shafts are connected by wheel gearing, the power being applied to one of the shafts. The water is drawn in at / and forced out at 0, the 568 STEAM POWER PLANT ENGINEERING |50 45 40 35 30 25 20 15 10 5 1 1 ., 7" ; ' (i .--• 3*"* "s. j^. **'&— \ A$ H *S / V / / <' / / / 7 / 1 / / / «o 800 900 1000 100 200 300 400 500 000 700 Capacity, Gallons per Minute. Fig. 301. Performance of a Six-Inch Worthington Conoidal Pump TOO 200 300 400 600 Fig. 302. Performance of a Single-Stage De Laval Centrifugal Pump. PUMPS 569 oo to no * ■OO i no Z TOTAL LirT. rEET 8 1 u «rp ^ejjS 1— 9- 70 °-~^. 60 < K *2 \ *£j Z*~ •" ~~~ ^V N b 05 W / ?0 R / 10 z s' r CALL DNS f •ER N 1INUT C J zoo 400 eoo eoo iooo isoo 1400 1600 leoo zooo 2200 2400 2soo zeoo 3000 3200 3400 3600 3aoo 4Q00 Fia. 303. Performance of a Two-Stage Lea-Degan Turbine Pump. 50 100 ISO 200 J SO 300 )S0 400 Fig. 304. Performance of a Two-Stage De Laval Centrifugal Pump. 570 STEAM POWER PLANT ENGINEERING Fig. 305. Two-Lobe Cycloidal Pump. Fig. 306. Rotary Pump with Movable Butment. 3 65 I 60£ 3 * "3 55 50 0.2, Efiicien cy ^y^ ^ ^^ ^" ,/ s t^ - ^^ / «p^>*^ — ° / ^2 ^e^^-- O^ Average I lead, 75 Ft. 300 400 500 600 700 Revolutions Per Minute 900 50 45 40 30- o 25=3 20 15 1000 £70 2.4 2.2 o 2.0 C4 1 1.8 I W 1.0 70 Head Constant, Speed Variable. 1 ^' :&£L~ ****' . *# r^>- jcjj2 io^ p$> Average Speed 812 R.P.M. Average Capacity 45.5 Gal.Per Minute •^ 100 110 120 130 140 150 Total Head, Eeet, Speed Constant, Head Variable. 160 170 180 Fig. 307. Performance of a Small Rotary Pump. PUMPS 571 displacement per revolution being equal to four times the volume of chamber A. There is no rubbing between impellers and casing. In this type of pump the pressure is independent of the speed of rotation, and the capacity varies almost directly with the speed. The slip varies from 5 to 20 per cent according to the dis- charge pressure. Fig. 306 shows a sec- tion through a rotary pump with movable but- ment. Fig. 307 illus- trates the performance of a 45-mm. Siemens- Schuckert rotary pump at different speeds and discharge pressures. (Zeit. d. Ver. Deut. Ing., June 24, 1905, p. 1040.) Large rotary pumps give much higher efficiencies, but the general charac- teristics are about the same. A combined effi- ciency of pump and engine as high as 84 per cent has been recorded. (Trans. A.S.M.E., Vol. 24, p. 385.) Rotary Pumps: Engr. U.S., Jan. 1, 1904, p. 51 Am. Mach., March 12, 1896 p. 238, Jan. 25, 1906, p. 103 Trans. A.S.M.E., 24-385 Eng. News, -March 1, 1900 p. 152; Power, Dec, 1903 p. 59, Aug., 1905, p. 477; Fig. 308. 10,000,000 Gallon Circulating Pump. The Constructor, Releaux, p. 226; Eng. Rec, Jan. 10, 1903, p. 59. Tests of Rotary Pumps: Zeit. d. Ver. Deut. Ing., June 24, 1905, p. 1040; Trans. A.S.M.E., 28-503. 572 STEAM POWER PLANT ENGINEERING 297. Circulating Pumps. — This term is ordinarily applied to the pumps which supply injection water to the condenser. These types are found in practice: the piston, the centrifugal, and the rotary pump. Figs. 224 and 236 show the application of recipro- cating pumps to condenser installations and Fig. 211a and Fig. 443 a similar application of centrifugal pumps. For large volumes of water and low heads the centrifugal or rotary pump is generally adopted on ac- count of minimum space require- ments and low first cost. In very large central stations where the demand for circulating water is enormous and the lift is moderately high, the high-duty pumping engine is often installed. Fig. 308 shows a section through one of the nine high-duty circulat- ing pumps at the New York Rapid Transit Company's power house. The steam end is operated by Corliss cylinders and is of the cross compound type. The maxi- mum capacity is 10,000,000 gallons per day (24 hours) against a head of 50 feet at mean low water. The actual lift is much less than this, as the discharge is aided by the vacuum in the condenser. Pumping Engines: Engr. U.S., Dec. 1, 1905, p. 786; Engng., Jan. 27, 1905, p. 132; Eng. Mag., Jan., 1905, p- 77 ; Eng. Rec, July 8, 1905, p. 58; Eng. News, Dec. 6, 1905, May 26, 1904; Trans. A.S.M.E., 3-141, 9-476, 12, 534, 975, 13, 83, 176, 14-1340, Fig. 309. The Pulsometer. 15-1103, 16, 49, 169, 21, 327, 788, 1018. 298. Air Lift. — The air lift is a simple arrangement of piping whereby water may be raised by means of compressed air. There are no working parts, and no valves are employed except to regulate the supply of air. Its particular field of application lies in pumping water from a number of scattered wells, and on account of the total absence of working parts it is peculiarly adapted to handling water containing sand, grit, and the like The device consists of a partially submerged PUMPS 573 water pipe and an air-supply pipe variously arranged as in Fig. 310(A) to 310(D). Compressed air forced into the water pipe at or near the bottom forms a series of bubbles or " pistons," as shown in D, which displace an equal volume of water. (For the theory of the air lift see Compressed Air, October, 1905, p. 3696.) The pressure required to operate the lift after it is once started is considerably less than the WATER LEVEL_ WELL Fig. 310. Different Arrangements of the " Air Lift. pressure of water due to the head, while that required to start it is slightly greater, consequently the pressure may be reduced after the pump begins to work properly. The successful operation of this device depends upon the ratio of the depth of submersion A, Fig. 310(D), to the total lift B, and the ratio of the area of the air pipe to that of the water pipe. The best results are obtained, in a general A sense, when the ratio — lies between 0.55 and 0.85, an average figure being 0.65, and when the area of the air pipe is 0.16 that of the water pipe. The quantity of air needed may be closely approximated by the following equation {Engineer, London, Aug. 14, 1903, p. 173): V = QL 20 in which V = cubic feet of free air per minute. Q = cubic feet of water per minute. L = lift in feet above the surface of the water. The velocity of the air should not exceed 4000 feet per minute. 574 STEAM POWER PLANT ENGINEERING The efficiency ("water" horse power divided by "air" horse power) varies from 30 to 50 per cent, increasing as the ratio — B increases from 0.55 to 0.85. {Engineer, U.S., Aug. 15, 1904, p. 564.) A number of tests give efficiencies (" water " horse power divided by I.H.P. of steam cylinder) varying from 20 to 40 per cent. The horse power required to compress one cubic foot of free air to different pressures per square inch, as determined from actual practice, is approximately as follows: Pressure in Pounds. Horse Power Required to Compress 1 Cubic Foot. Pressure in Pounds. Horse Power Required to Compress 1 Cubic Foot. 176 140 100 80 0.434 0.376 0.201 0.189 60 45 30 0.159 0.145 0.121 (Engr., Lond., Aug. 14, 1903, p. 174, Dec. 11, 1903, p. 568, Feb. 12, 1904, p. 172.) When it becomes necessary to raise water to a height exceeding say 175 feet above the level in the well, it is customary to use two or more pumps, the total lift being divided between them. Air Lift: Engr., Lond., Aug. 14, 1903, p. 173, Dec. 11, 1903, p. 568, Feb. 12, 1904, p. 172, Jan. 10, 1908; Eng. Rec, Jan. 7, 1905, p. 8; Engineering, Jan. 22, 1904, p. 135, Jan. 29, 1904, p. 166; Compressed Air, Aug., 1903, Oct., 1905; Engr. U.S., July 15, 1903, p. 547, Sept. 21, 1906, Aug. 15, 1904, p. 564; Mech. Engr., Aug. 20, 1904; Engng. Rev., July, 1904; Power, March, 1905, p. 173; Eng. News, Jan. 16, 1908. Pulsometer: Tech. Quar., Sept., 1901; Public Works, Aug. 15, 1904; Engr. U.S., July 15, 1904; Experimental Eng., Carpenter, p. 621; Thermodynamics, Wood, p. 293; Trans. A.S.M.E., 13-211. Cost of Operating American Pumping Stations: Eng. Rec, Aug. 6, 1904; Proc. Engrs. Club of Phil., Oct., 1906. Complete Description of Various American Types of Steam, Rotary, and Centrifugal Pumps: Engr. U.S., Jan. 1, 1904. The Growth of the Pumping Station: Eng. Rec, July 14, 1906, p. 50. The Selection of Waterworks Pumping Machinery: Eng. News, Vol. 52, p. 39. Centrifugal Pump for Boiler Feeding: Power & Engr., Mar. 15, 1910. Recent Records of High Duty Pumping Engine: Eng. News, Feb. 3, 1910. CHAPTER XIV. SEPARATORS, TRAPS, DRAINS. 299. Live-Steam Separators. General. — The function of a steam separator is the removal of entrained water from steam. Unless a boiler is liberally provided with superheating surface, the steam may contain an amount of moisture varying from 0.3 to. 5 per cent. If the boiler is poorly proportioned or forced far above its rating, this percentage may be greatly increased. The quality of the steam is still further reduced by condensation in the steam pipe, which may vary from one to ten per cent, depending upon the length of pipe and efficiency of covering. One of the effects of moisture in steam is to increase its density and reduce its elastic- force. It also increases its conductivity, so that during the work of expansion more heat is absorbed from the walls of the cylinder and discharged into the atmosphere or into the condenser without doing useful work. (Ewing, " The Steam Engine," p. 151.) Although the heat loss from this cause is small, the danger arising from the introduction of a considerable amount of water in the cylinder renders the removal of the moisture necessary. See page 248 for influence of moisture on steam consumption. The essentials of a good separator are high efficiency as a water eliminator, ample storage capacity for any sudden influx of water, simplicity and durability in construction, and small resistance to the current of steam passing through. A good separator may be relied upon to remove practically all of the moisture from steam containing under ten per cent entrainment and all but two per cent from steam containing as much as twenty per cent. (Engineer, U.S., Jan. 15, 1904.) Table 75 gives the results of a series of tests made by Professor R. C. Carpenter in 1891 of six steam separators. (Power, July, 1891, p. 9.) Conclusions from these tests were: 1. That no relation existed between the volume of the several separators and their efficiency. 2. No marked decrease in pressure was shown by any of the separa- tors, the most being 1.7 pounds by separator E. 575 576 STEAM POWER PLANT ENGINEERING 3. Although changed direction, reduced velocity, and perhaps centri- fugal force are necessary for good separation, still some means must be provided to lead the water out of the current of the steam. A series of tests made at Armour Institute of Technology in 1905 on a number of separators showed that the efficiency of separation decreased as the velocity of the steam increased.* At the low velocity of 500 feet per minute all separators were equally efficient, at a velocity of 5000 feet per minute several had little effect on eliminating the moisture present, and at a velocity of 8000 feet per minute only one gave efficient results. TABLE 75. TESTS OF STEAM SEPARATORS. (R. C. Carpenter.) Test with Steam of about 10 per Cent of Moisture. Make of Separator. Quality of Steam Before. Quality of Steam After. Efficiency. B Per Cent. 87.0 90.1 89.6 90.6 88.4 88.9 Per Cent. 98.8 98.0 95.8 93.7 90.2 92.1 Per Cent. 90.8 80.0 59.6 33.0 15.5 28.8 A D C E F Tests with Varying Moisture. Quality of Steam Before. Per Cent. 66.1-97.5 51.9-98 72.2-96.1 67.1-96.8 68.6-98.1 70.4-97.7 Quality of Steam After. Per Cent. 97.8-99 97.9-99.1 95.5-98.2 93.7-98.4 79 . 3-98 . 5 84.1-97.9 Average Efficiency. Per Cent, 87. 76. 71. 63. 36. 28.4 300. Classification of Separators. — Separators are based on one or more of the following principles of action: 1. Reverse current. The direction of the flow is abruptly changed, usually through 180 degrees. This causes the water in the steam, on account of its greater specific gravity, to be thrown into a receiving vessel, while the steam passes on in a reverse direction. 2. Centrifugal force. A rotary motion is imparted to the steam whereby entrained water particles are eliminated by centrifugal force. 3. Baffle plates. The flow is interrupted by corrugated or fluted plates to the surfaces of which the water particles adhere and from which they fall by gravity to the well below. 4. Mesh. The separation is brought about by mechanical filtration through screens or meshes. The following outline shows the classification of typical, separators, in accordance with the above principles : * See Power, May 11, 1909, p. 834. SEPARATORS, TRAPS, DRAINS 577 Live-steam separators Exhaust-steam separators Reverse current Centrifugal Baffle plate Mesh Hoppes, Fig. 311. |Stratton, Fig. 312. Keystone, Fig. 313. Mosher. Robertson. Bundy, Fig. 314. Austin, Fig. 315. Detroit. Direct, Fig. 316. Potter mesh. ( Jacketed baffle Baum, Fig. 317. (Absorption Loew, Fig. 318. 301. Reverse-Current Steam Separators. — Fig. 311 shows a section through a Hoppes steam separator and illustrates the principle of reverse-current separation. Steam may flow through in either direction. Both the inlet and outlet ports are surrounded by gutters C, C, partly filled with water, which intercept the moisture following the surface of the pipe, while the down- ward plunge of the steam throws the entrained water to the bottom of the separator. The condensation is carried from the troughs by pipe P to the well below, from which it is trapped at D in the usual way. The velocity of the steam in passing through this separator is greatly reduced to prevent the steam from taking up the water in the bottom of the well. This is brought about by increasing the area of the passage through the separator. Fig. 312 gives a sectional view of a Stratton separator, which, though primarily of the reverse-current type, embodies also the principle of centrifugal force. The separator consists of a vertical cast-iron cylinder with an internal central pipe C extending from the top downward for about half the height of the apparatus, leaving an annular space between the two. The current of steam on entering is deflected by a curved partition and thrown tangentially to the annular space at the side, near to top of the apparatus. It is thus whirled around with all the velocity of influx, producing the centrifugal action which throws the particles of water against the outer cylinder. These adhere to the surface, so that the water runs down continuously in a thin sheet around Fig. 311. Hoppes Steam Sepa- rator. 578 STEAM POWER PLANT ENGINEERING the outer shell into the receptacle below. The steam, following in a spiral course to the bottom of the internal pipe, abruptly enters it, and passes upward and out of the separator without having once crossed the stream of separated water. The rapid rotation of the current of steam imparts a whirling motion to the separated water which tends to interfere with its proper discharge from the apparatus. The separator has therefore been provided with wings or ribs E projecting at an acute angle to the course of the current, which have the effect of breaking up this whirling motion and allowing the water to settle quietly at the bottom, whence it passes off through the drain pipe D. Fig. 312. Stratton Steam Separator. Fig. 313. Keystone Steam Separator. 303. Centrifugal Steam Separators. — Fig. 313 shows a section through a Keystone or Simpson's centrifugal separator. The separator consists of a cast-iron cylinder with vertical pipe C extending down- ward about two-thirds of the whole length; this pipe has a thread or screw wound spirally around it, the space between the threads being somewhat greater than the area of the steam pipe. The steam passing around the spiral course causes the water to be thrown against the outer walls by centrifugal force, while the dry steam passes through SEPARATORS, TRAPS, DRAINS 579 Fig. 314. Bundy Steam Separator. the small holes in the central pipe. The water passes down the outer walls, where its motion is arrested by obstructing ribs E, and is thence carried away by a drip pipe D to a suitable drain. 303. Baffle-Plate Steam Separators. — Fig. 314 gives an interior view of a Bundy separator and illustrates the application of baffle plates for live steam separation. This separator consists of a rectangular cast-iron casing with a cylindrical receiver beneath it. Directly across the steam passage are baffle plates corrugated for the reception of entrained water. The plates consist of vertical castings, each containing a main artery or channel which leads directly to the receiver. The fronts of the plates are flat, with a series of recesses sloping inwards and downwards, terminat- ing in an opening of capillary size leading to the main artery. The plates are staggered, so that the steam must impinge against all of them in its passage. The particles of water adhere to the plates, collect, and fall by gravity into the receiver. The flanges at the bottom constrict the opening of the reservoir so as to prevent the steam from picking up any portion of the water. Fig. 315 shows a section through an Austin separator and illustrates another class embodying the fluted baffle plate principle. The steam in passing through the chamber impinges against the fluted baffle plate B. The moisture adheres to the surfaces, collects, and trickles along the corrugations to the bottom of the well. These corrugations are formed in such a manner that the steam cannot come in contact with the water particles after they have been once eliminated. A per- forated diaphragm D prevents the water in the well from coming in contact with the steam. The current of steam is also reversed, thus giving additional separating properties to the apparatus. 304. Mesh Separators. — Fig. 316 shows a section through a " direct " separator, illustrating the principle of mesh separation. These separa- Fig. 315. Austin Steam Sepa- rator. 580 STEAM POWER PLANT ENGINEERING tors are made with steel bodies and cast-iron heads and bases, in all sizes up to six inches inclusive, the larger sizes being constructed of cast iron or boiler plate. The cone C, perforated lining E, and dia- phragm S are made of cold-rolled copper; the cone is a substantial gray-iron casting, resting on three cast- iron supports hooked over the top of inner pipe as indicated. The method of operation is as follows: The accumu- lated moisture around the walls of the steam pipe is caught by the upper edge of cone C and carried down back of lining E to the water chamber. The current of steam entering the separator impinges upon the conical surface, which is composed of solid plate covered with sieve S, through which water may freely pass but from which it cannot readily escape. Passing through the sieve and depositing on the solid surface of the cone 0, this water is carried by conductors P to the water chamber. Perforated lining E permits the moisture content of the steam to pass through the opening to the water below and prevents it from coming in contact again with the current of steam. A trough is provided at the lower edge of the inverted cup which leads all the water that may ad- here to it to the water chamber. The steam flows through the passages indicated by arrows and is sub- jected to a whipsnapping action which tends to throw off any re- maining moisture. The perforated plate D prevents the steam from picking water out of the water chamber. 305. Location. — Live-steam separators may be located 1. Inside the boiler. 2. Between boiler and engine. 3. At the steam chest. Where the steam pipe is very short, and particularly in marine and locomotive work where the tossing of the boiler induces excessive priming, the separator may be placed inside the boiler and its function becomes that of a dry pipe. In this location it prevents the water due to foaming and priming from passing to the engine, and Fig. 316. " Direct " Steam Sepa, rator. SEPARATORS, TRAPS, DRAINS 581 reduces condensation in the pipe by supplying dry steam. The " Potter mesh " and the " De Rycke centrifugal " are types of sepa- rators designed for this service. The arrangement of separator between engine and boiler, other than at the throttle, or inside the boiler is sometimes necessary for economy of space. Where possible, however, the separator should be placed close to the steam chest. Current practice recommends that a receiver separator, which is an ordinary separator with a volume of two to four times that of the high-pressure cylinder, be placed close to the engine if the load is intermittent or sharply fluctuating. This forms a cushion for absorb- ing the force of the blows caused by cut-off, delivers steam at a prac- tically uniform pressure, and reduces the vibration of the piping to a minimum. It also provides a reservoir for sudden demands made by the engine. Smaller pipes and higher velocities may be used with this arrangement. 306. Exhaust-Steam Separators and Oil Eliminators. — The function of an exhaust-steam separator is the removal of cylinder oil from the steam exhausted by engines and pumps. In plants where exhaust steam is used for heating it is quite essential to remove the oil from the steam before it enters the heating system, for the oil not only reduces the efficiency of the radiators by coating them with an excel- lent non-conducting film but is an element of danger to the boiler itself. In condensing plants the separator will prevent the oil from fouling the condenser tubes and those of the vacuum heater if one is installed; this is an important factor, since the oil or grease lowers the efficiency of the heat transmission. In a general sense a live-steam separator is also an oil eliminator, and all the separators previously described perform this function to a cer- tain extent, since the underlying principles governing the elimination of oil from exhaust steam are similar to those employed in removing water from steam. Most of the separators described above are also designed, in lighter form, as oil eliminators, but by far the greater number are based on the fluted baffie plate principle, of which the Hine, Bundy, Cochrane, Utility, Peerless, and Keiley are well-known examples. This type of oil separator will eliminate a considerable portion of the oil in the steam provided the baffle plates or corrugated surfaces are frequently cleaned. The following is taken from the report of Professor R. Burnham of the Armour Institute of Technology on the test of a six-inch horizontal oil separator of the baffle-plate type: " For purposes of test the separator was placed in the exhaust line 582 STEAM POWER PLANT ENGINEERING of a 9 x 18 x 24 cross compound Corliss engine running under its maximum load at 80 pounds pressure and exhausting into a Wheeler surface condenser against 26 inches vacuum. " Cylinder oil was fed through the lubricators of the high and low pressure cylinders at the rate of from 5 to 20 drops per minute, a record being made of the exact quantity of oil fed per hour. The separator was so arranged, by means of a receiver connected to the air pump, that the accumulation of oil and water could be readily trapped from it at any time. In order to determine the quantity of oil given up by the condenser, and not properly charged against the separator, each series of efficiency tests was preceded by a run of three hours during which time no oil whatever was fed to the cylinders. During the last hour a record was made of the weight of steam used and a sample of the con- denser discharge retained for analysis. " The efficiency tests were made by feeding at an excessive rate through the lubricators as described above, and when conditions became practically constant, records were made for one hour, of the weight of oil used, weight of condensed steam, and drain from separator. Samples of the two latter were retained for analysis and the percentage of oil in them accurately determined, correction being made for the oil given up by the condenser. A second series of tests was made exhausting at atmospheric pressure. The results obtained are tabulated below. Oil in condensed steam with no oil feeding. (Charged to condenser.) Pounds per hour Oil fed to cylinder, pounds per hour Steam condensed per hour, pounds Oil caught by separator, per hour, pounds A Oil in condensed steam (corrected), pounds per hour .B Percentage of oil in condensed steam by weight, per cent Efficiency of separator, percent . „ ... Exhausting into 26-inch Vacuum. .051 .057 .0559 .401 .562 .934 1000 1120 1096 .341 .450 .743 .009 .010 .0096 .0009 .001 .00088 97.4 97.8 98.8 Exhausting at Atmospheric Pressure. .0353 .0340 .621 .710 905 872 .552 .583 .0071 .0050 .00078 .00057 98.7 99.1 " There was practically no free oil on the surface of the condenser discharge in any case, the small quantity of oil which passed the separa- tor (from 5 to 10 parts in a million of water by weight) existing as an emulsion, imparting a slight milky color to the water." It is a well-established fact that oil can be more effectually removed from wet than from dry steam, and some makers, notably the Austin SEPARATORS, TRAPS, DRAINS 583 FROM ENGINE Separator Company, inject a cold-water spray into the separator cham- ber. A similar result is brought about in the Baum separator, Fig. 317, in which the corrugated baffle plate is hollow and cold water is forced through the chamber thus formed. Referring to Fig. 317 : The diverged baffle plate forms the wall of a chamber in which cold water is con- tinually circulated. This circulation causes moisture to appear on the baffle plate surface. The particles of oil, coming in contact with this moist sur- face as the steam current is diverged, adhere to it and fall by gravity into the well below, where they are completely isolated from the purified steam. A large portion of the oil and water, however, does not enter the separator at all but is caught by the inside ledge near the junction of the exhaust pipe and the separator. The oil and condensation which are carried along the bottom of the pipe come in contact with this ledge and are carried directly to the outlet pipe. A very successful method of removing oil from steam is to project the steam on to the surface of a body of water. The water may be hot or cold and will hold the oil if it once reaches the surface. It is essential, however, to reduce the velo- city of the steam as it passes on its way to the outlet. Baldwin's grease separator is based upon this principle. (Baldwin on Heating, p. 234.) The most efficient method of removing oil is by combined filtration and absorption. {Engineering News, May 22, 1902, p. 406.) A large chamber filled with coke, brick, broken tile, or other absorption material is placed in series with the exhaust pipe. The steam passing through this chamber is entirely freed from oil and moisture, provided the absorb- ing material is sufficient in quantity and is replenished as soon as it becomes saturated with oil. The annoyance attending the removal and replenishing of the absorbing material at frequent intervals and the great size of the apparatus are serious drawbacks. An example of this system of purification in which many of the objectionable features are reduced to a minimum is the Loew grease and oil separator, Fig. 318. Fig. 317. DRAIN Baum Oil Separator. 584 STEAM POWER PLANT ENGINEERING The exhaust steam enters the chamber at the top, strikes a large deflecting plate shaped like an inverted V, and permits part of the condensation and oil to be drawn off by the drain pipe. The steam then rises and is deflected, as indicated, against a series of shelves filled with fibrous material covered with coarse wire screens. The grease is removed from each shelf by suitable drains. This apparatus is sectional and any number of sections may be added without affecting the rest. In a non-condensing plant where the exhaust steam is used for heating pur- poses the oil separator is ordinarily placed in the main exhaust pipe just before it enters the heating system. Where several branches enter one main it is not customary to place a sepa- rator in each branch, one large separator located as above being sufficient. In condensing plants oil separators are seldom installed except where sur- face condensers are used, in which case the separator may be placed anywhere between the engine and condenser. In case a vacuum heater is used the separator may be placed on either side of the heater, depend- ing upon the type of separator. If the separator is of the " jacket cooling " or " spray " type, it may be placed between the engine and the vacuum heater; if. however, it is of the " baffle-plate " type, the oil will be more efficiently removed if the separator is placed between the heater and condenser so that it will get the benefit of the moisture formed in the heater. In the latter location, however, the separator will not prevent the oil from fouling the heater tubes. Where a jet condenser is used and water is taken from the hot well, the hot well itself acts as an oil separator. (Trans. A.S.M.E., 24- 1144.) All separators, steam and oil, should be provided with gauge glasses and should be thoroughly drained and the drainage should be automatic. 307. Exhaust Heads. — The function of the exhaust head is the elimination of oil and water from steam exhaust before permitting it Fig. 318. Loew Grease Extractor. SEPARATORS, TRAPS, DRAINS 585 to be discharged into the atmosphere. Unless removed, the water and oil rot the roofs and walls in summer and pollute the atmosphere surrounding the plant. The exhaust head also acts as a muffler reducing the noise of the escaping steam. Exhaust heads are built on the same principle as steam and oil separators and most separator builders manufacture them. Fig. 319 shows a section through a typ- ical exhaust head. The condensa- tion is ordinarily drained to waste, though with proper purification it may be returned to the boiler. With an efficient oil separator in the exhaust line the condensation in the exhaust head may be returned directly to the boiler without further purification. Live-steam separators are propor- tioned so that it is only necessary, in the average installation, to specify the size of pipe, the type of engine, the steam pressure, and the style, whether horizontal or vertical. Gauge glasses, gauge cocks, and companion flanges are usually provided by the maker. In some cases the capacity of the reservoir is also specified. In specifying oil extractors the following additional data are necessary for an intelligent choice : the number of engines and pumps exhausting into the line, the location of the separator, the steam pressure, velocity, and the quality and quantity of cylinder oil used. A guarantee of efficiency and of material and workmanship is often demanded. i REFERENCES. Water and Oil Separators: Am. Elecn., Jan., 1900. Steam Separators: Am. Elecn., June, 1905. " Dry Steam ": Goubert Mfg. Co., 85 Liberty St., N.Y. (Catalogue). Cochrane Separator: Harrison Safety Boiler Works, Philadelphia, Pa. (Catalogue). Bundy Separator: A. A. Griffin Iron Co., N.Y. (Catalogue). A Bad Case of Discharge Water with Steam from Water-Tube Boilers: Trans. A.S.M.E., Vol. 26. Location of Steam Separators: Power, Oct., 1904. Fig. 319. A Typical Exhaust-Head. 586 STEAM POWER PLANT ENGINEERING Condensing Exhaust Head: Eng. News, Vol. 49, p. 419; Eng. Rec, Vol. 40, p. 177. Experiments in Separating Oil from Condensed Steam: Eng. News, May 22, 1903; Eng. Rec, April 27, 1901; Engr., Lond., March 12, 1897, Oct. 20, 1905; Power, Aug., Sept., 1896, May, 1903; Heating and Ventilation, Feb., 1897; Trans. A.S.M.E. Vol, 17, p. 295. Oil Separators: Baldwin on Heating, pp. 233-237. Oil Separation in a Combination Engine and Turbine Plant: Power, Oct., 1906. Test of a Cochrane Steam Separator: Power, April, 1898. Test of Lippincott Separator: Engr. U.S., Aug., 1902. Test of a Linstrum Steam Separator: Engr. U.S., June 15, 1904. Test of an Austin Steam and Oil Separator: Trans. A.S.M.E., Vol. 20, p. 489. Test of a Detroit Live Steam Separator: Engr. U.S., April 15, 1904; Power, Jan., 1902. Tests of Direct Separators, Oil and Steam: Direct Separator Co., Syracuse, N.Y. (Catalogue). Tests of Six Steam Separators: Power, July, 1891, p. 9; Engr. News, Vol. 26, p. 233. Test of a " Utility " Oil Eliminator: Engr. Rec, May 2, 1903; Engr. Rev., May, 1903. The Hot Well as an Oil Extractor: Trans. A.S.M.E., 24-1144. 308. Drips. — No matter how thoroughly a steam pipe or reservoir may be covered with insulating material considerable condensation takes place. With the best covering this loss approximates one-sixth of a pound of steam per square foot of pipe surface per hour for steam pressures of one hundred pounds, and runs as high as one pound of steam for bare pipes. See Table 80 for results of experiments on the loss of heat from bare pipes, and Table 81 for data on the efficiency of pipe coverings. In addition to this water of condensation, from \ to 2 per cent of moisture is carried over by the steam from the boiler. This water, unless thoroughly removed, is a constant source of danger to the engines and causes water hammer and leaky joints in the piping. A joint on a steam pipe may safely withstand a steam pressure of 100 pounds without leaking and still leak badly under a water pressure of half that amount. This is due to the fact that the steam with its high temperature causes the pipe to expand, thus insuring a tight joint, while the entrained water (which cools as it collects) causes the pipe to contract and allows a leak. The entrained water and water of condensation are usually spoken of as " drips." Drips may be divided into two classes, low pressure and high pressure. 309. Low-Pressure Drips. — Low-pressure drips include the steam condensed in exhaust steam feed heaters of the closed type, exhaust steam piping, receiver barrels, steam chests, and exhaust heads. As SEPARATORS, TRAPS, DRAINS 587 these drips are impregnated with oil and are useless for boiler feed without purification, they are usually discharged to waste. Most city Fig. 320. Closed Heater Installation for Abstracting Heat from Oily Drips. ordinances require the drips to be cooled to 100 degrees F. before being discharged into the sewer. In this case they must be first dis- charged into a tank and permitted to cool. This tank must be vented to the atmosphere to prevent back pressure. Fig. 320 shows an installation in which the heat asbtracted from the drips, etc., is used to heat the feed water. The drips from the throttle valve and steam chest in a non-condensing plant are ordinarily discharged into the exhaust pipe as shown in Fig. 321. In a con- densing plant the throttle drips are piped to a trap or to the free exhaust pipe. The returns from a steam- heating system are sometimes classi- Simpie Method of Draining fied as low-pressure drips. They are Dn P s - invariably returned to the boiler. In small plants all the low-pressure drips may be connected to one large pipe and this pipe in turn to a single trap, provided there is but Fig. 321 588 STEAM POWER PLANT ENGINEERING little difference in pressure in the various drip pipes. In case of different pressures separate leads should be run to waste or traps. The drips from the receiver and vacuum heater barrels in a con- densing plant are oftentimes under less than atmospheric pressure, and sometimes the pressure varies from a slight vacuum to 10 or 20 pounds gauge, and consequently cannot be disposed of as described above. If possible, the heaters and receivers should be placed so as to drain into the condenser (see Fig. 334). Should this arrangement prove impracti- cable, the barrels may be drained by a trap especially arranged as shown in Fig. 335. 310. Size of Pipe for Low-Pressure Drips. — In the average exhaust- steam feed-water heater one pound of steam in condensing gives up approximately 1000 heat units. This will heat about 6 pounds of water from 60 to 200 degrees F. Hence the area of the drip which carries the water of condensation from the closed heater need be but J- that of the feed pipe. In no case, however, should a pipe smaller than J inch in diam- eter be used. Should the same pipe be used for both exhaust head and heater drips, an area of \ area of feed pipe would prove of ample capacity. In practice it is customary to use the size of pipe conforming with the outlet furnished by the manufacturer of the apparatus, and only when several pieces of apparatus are connected to one main are calculations made for the size of this main. The drip pipe from the throttle valve is ordinarily J inch diameter irrespective of the size of steam pipe; this is also true of the steam-chest drip. 311. High-Pressure Drips. — High-pressure drips consist of those which are condensed under practically boiler pressure and include the steam condensed in steam pipes, cylinder jackets of engines, reheating coils of receivers, and separators. Being free from oil and containing considerable heat, they are usually returned to the boiler. Drips may be returned to the boiler automatically by means of 1. Steam traps. 2. Holly steam loop. 3. Pumps. 312. Classification of Steam Traps. — Steam traps may be divided into two classes, depending on their use, — return and non-return. Both of these two classes may be subdivided into five types according to the principle of operation, viz. : I. Float. III. Bowl. II. Bucket. IV. Expansion. V. Differential. SEPARATORS, TRAPS, DRAINS 589 Return Traps. Traps which receive the condensed steam and return it to a boiler having considerably higher pressure than that acting on the returns are known as return trays. They are made in a great variety of styles. The general principle of operation is shown in Fig. 331 and described in paragraph 318. Non-Return Traps. Non-return traps, as the name implies, are used where the water of condensation is not returned to the boiler but is discharged into any receptacle having less than boiler pressure. CLASSIFICATION OF A FEW WELL-KNOWN STEAM TRAPS. Float Bucket Dump Steam Traps Expansion Differential (McDaniel. ( Cookson. !Acme. Albany. !Bundy. Morehead. Metal Volatile-Fluid (Flinn. (Siphon. {Columbia. Geipel. ! Dunham. Heintz. Fig. 322. McDaniel Float Trap. 313. Float Traps. — Fig. 322 shows a section through a McDaniel improved trap, illustrating the principles of the float type. A hollow sphere C of seamless copper pivoted at E rises and falls with the change of water level in the vessel. The discharge valve M is operated by the 590 STEAM POWER PLANT ENGINEERING float. When the trap is empty the float is in its lowest position and the discharge valve is closed. Water of condensation flows into the trap by gravity through opening D to a certain depth, when the float opens the discharge valve and the steam pressure acting on the surface of the water forces it through outlet S to tank or atmos- phere. After the water is discharged the float closes the valve and permits the condensation to collect again. A gauge glass indicates the height of water in the chamber. Unless float traps are well made and proportioned there is a danger of considerable steam leakage through the discharge valve, due to unequal expansion of valve and seat and the sticking of moving parts. The discharge from a float trap is usually continuous, since the height of the float, and consequently the area of the outlet, is proportional to the amount of water present. When the trap is working lightly, this adjustment is apt to throttle the area and create such a high velocity of discharge as to cause a rapid wear of valve and seat. This defect is more or less evident in all steam traps discharging continuously. For this reason all wearing parts should be accessible and readily replaced. Fig. 323. Acme Bucket Trap. 314. Bucket Traps. — Fig. 323 shows a section through an " Improved Acme " steam trap. The water of condensation enters the cast-iron vessel at A, filling the space D between the bucket E and the walls of SEPARATORS, TRAPS, DRAINS 591 the trap. This causes the bucket to float and forces valve V against its seat (valve V and its stem being fastened to the bucket as indicated) . When the water rises above the edges of the bucket it flows into it and causes it to sink, thereby withdrawing valve V from its seat. This permits the steam pressure acting on the surface of the water in the bucket to force the water through the annular space H to discharge opening G. When the bucket is emptied it rises and closes valve V and another cycle begins. By closing valve R the trap is by-passed and the condensation blows directly through passage C to discharge G. The discharge from this type of trap is intermittent. 315. Dump or Bowl Traps. — Fig. 324 shows sections through a Bundy bowl trap of the " return " type. The water enters the bowl Fig. 324. A Typical Tilting Trap. through trunnion D and rises until its weight overbalances counter- weight E and the bowl sinks to the bottom. As the bowl sinks, arm G, which is a part of the bowl, rises and engages the nuts N on valve stem H and opens valve /, thus admitting live steam pressure on to the surface of the water. The trap then discharges like all others. After the water is discharged weight E sinks and raises bowl A, which in turn closes valve /, and the cycle begins again. Bowl traps are necessarily intermittent in their discharge. Fig. 335 shows the application of a bowl trap to a receiver where the drips are under a vacuum, and Fig. 336 a similar application to an engine receiver where the pressure varies from less than atmospheric pressure to a pressure of 40 or 50 pounds. 592 STEAM POWER PLANT ENGINEERING 316. Expansion Traps. — Expansion traps may be divided into two groups : (1) Those in which the discharge valve is operated by the relative expansion of metals and (2) Those in which the action of a volatile fluid is utilized. Expansion traps will never freeze, as they are open when cold and all the water drains out before the freezing temperature is reached. Since traps of this type have little capacity for holding water, 5 to 10 feet of pipe should be provided between the trap and the pipe to be drained in order that the condensation may collect and cool. Fig. 325 shows the general appearance of a Columbia expansion trap in which the valve is operated by the expansion of metallic tubes. INLET D OUTLET Fig. 325. A Typical Expansion Trap. Water gravitates to the trap through opening marked " inlet/' passes through brass pipe 0, then downward to the main body of the valves and back to outlet valve C. Below pipe and parallel to it is an iron rod S, at the end of which is the support or fulcrum of lever R. The lower end of this lever is connected to the stem of the valve C, so that any movement of the lever is communicated to it. When the trap is cold, valve C is open and all water of condensation passes out. The moment steam enters the pipe it expands. The amount of expansion is multiplied several times by the action of the lever R, so that the movement of the valve is much greater than the expansion of the pipe 0. The compensating spring D prevents the brass tube from damaging itself by excessive expansion. Lever A permits the trap to be blown through by hand. Fig. 326 shows a section through a Geipel trap in which the valve is operated directly by the expansion of two metallic tubes and the movement is not multiplied by levers as with the Columbia. The lower or brass pipe constitutes the inlet, and is connected to the vessel to be drained; the upper or iron pipe is the outlet for discharge. The two pipes form the sides of an isosceles triangle, the base F of which is SEPARATORS, TRAPS, DRAINS 593 rigid, while the apex A is free to move in a direction at right angles to the linear expansion of the tubes. When cold, the brass pipe is con- tracted and the apex, in which the valve seat is placed, is moved down so that the valve is open and the water is discharged. As soon as steam Fig. 326. Geipel Expansion Trap. enters the brass pipe the latter expands and forces the valve seat against the valve. The trap may be adjusted for any pressure by means of the lock nuts E. When it is desired to blow through, the valve may be operated by hand by pressing the lever. Fig. 327 shows a section through a Dunham trap. It operates upon the expansion principle, utilizing a fluid of a volatile character as its v~> Fig. 327. Dunham Expansion Trap. motive force. The corrugated bronze disk B is filled with a volatile fluid, and expands and contracts according to the pressure exerted by the fluid. The water enters at the top, surrounds disk B and passes through valve opening D to discharge outlet at E. As soon as steam strikes the disk B the volatile fluid flashes into a vapor and causes the disk to expand. This expansion forces valve D against its seat and the discharge ceases. The valve will remain closed until the con- densation collects and cools the disk B, which then contracts, opens 594 STEAM POWER PLANT ENGINEERING the valve, and condensation enters as before. The adjustment, how- ever, is such that the discharge may be made continuous instead of intermittent. The Dunham trap is claimed to be the smallest trap of its capacity on the market. The 1-inch size, having a capacity for draining 10,000 lineal feet of 1-inch pipe under 60 pounds pressure, weighs but 5 pounds and may be connected to the pipe line as if it were a globe valve. Fig. 328 shows an internal view of a Heintz steam trap. This works on the principle of the volatile fluid expansion trap but in Fig. 328. Heintz Expansion Trap. a different manner from any of those described above. The requisite movement is obtained by the elongation and contraction of the extremities of a bent metallic tube T filled with a highly volatile fluid. This tube is inclosed in a cast-iron box and presses against the point of regulating screw P. The other extremity of the tube carries the valve and is free to move under the action of the variations of temperature. Spring S has no connection with the action of the trap. It is used as a simple means of holding one end of the expansion tube on its pivot. The trap operates as follows: Water enters at /, surrounds the tube T and passes through the valve to the discharge outlet O. As soon as steam enters the chamber the volatile fluid in the tube flashes into a vapor and the pressure thus created tends to straighten out the tube; this forces the valve against its seat and the discharge ceases. As the trap cools the tube returns to its normal position and the discharge valve is opened, thus permitting the condensation to drain out. The adjustment permits of continuous or intermittent discharge and of variable pressures. 317. Differential Traps. — Fig. 329 shows a cross section through a Flinn differential trap. The column of water X acting on dia- phragm D closes valve V. The water entering pipe E and the action of the spring equalize column X and open the valve. SEPARATORS, TRAPS, DRAINS 595 Describing the action in further detail, the water of condensation enters at A, fills lower chamber Y, pipe X, and receiving chamber C up to the level of the top of pipe E. This column of water acting on the under side of the diaphragm D forces the valve to its seat against the counter pressure of the spring S. Any additional water that enters the trap overflows through pipe E, filling chamber F and pipe E to a point about midway of its height, where the effect of the column of water in pipe X is balanced. The pressure on each side of the diaphragm is then equal, the short column in pipe E, aided by the spring, balancing the pressure of the longer column in pipe X. Any further increase in the height of the water in pipe E causes a depression of the valve V, which allows water to escape until the column has fallen to a level a little below the middle of pipe E, when this valve closes again. This action is repeated at intervals according to the quantity of water entering the trap. So long as the water keeps coming in sufficiently large quantities the valve remains wide open. Fig. 330 gives a general view of a siphon trap which is much used in draining low-pres- sure systems, as, for example, the separator in an ex- haust steam heat- ing system. It con- sists essentially of two legs A and B, which may be close together or any distance apart but the lengths of which must be sufficiently great to prevent pressure acting through pipe / from forcing the water out of B. C is a vent pipe extending to the air to prevent siphoning; is the discharge for the condensed steam. In ordinary operation the leg B is filled with water which is constantly overflowing, and A with steam and water, the FP° Y Fig. 329. V Y Flinn Differential Trap. G Fig. 330. Simple Siphon Trap. 596 STEAM POWER PLANT ENGINEERING EQUALIZING VALVE total pressure in both legs being equal. The siphon trap is applicable for low pressure only, as it requires approximately 2.3 feet of vertical space E for each pound per square inch pressure in the pipe. The maximum allowable head is represented by vertical distance N. 318. Location of Traps. — Wherever possible a trap should be located so that the condensation will flow into it by gravity. This will insure positive drainage. Sometimes, however, the coils, cylin- ders, or pipes to be drained are located in a pit or trench or lie on a basement floor where it is impossible to set the trap so as to receive the drains by gravity without placing it in an inaccessible position. With very low pressures this is often unavoidable, but with pressures of five pounds or more the trap may be placed above the point to be drained. If a trap is set in an exposed place a drain should be pro- vided at the lowest point to free the pipe of water when steam is shut off. A dirt catcher or strainer should be placed in the pipe leading to the trap to prevent scale, etc., from reaching the valve. All pockets and dead ends should be drained, and no condensation should be allowed to accumulate. High and low-pressure drips should be kept separate. All tanks should have gauge glasses. Fig. 331 shows the ap- f \ STEA " SUPPLY — r ^ plication of a float trap for automatically returning water to the boiler. For this purpose the trap must be placed three feet or more above the water line in the boiler, so that the water may gravitate to it. Water is forced into the trap from the returns through pipe A until it reaches a level where the float opens the equalizing valve V and permits steam from the boiler to enter the trap, thus equalizing the pressures. The water then flows into the boiler by gravity through check valve D. At the end of discharge the float closes the equalizing valve and another cycle begins. Check valve C prevents the water from being forced back to the return pipe. If the pressure in the return pipe A is not sufficient to force the water into F£> RETURN TRAP ■=-=--WATER LEVEL Fig .331. Return Trap . SEPARATORS, TRAPS, DRAINS 597 the trap, a pump or another trap may be used to effect this result. Practically any high-pressure trap may be converted into a return trap by proper installation and an " equalizing " valve. TRAP TRAP Fig. 332. Drainage System for Jackets and Receivers of Triple Expansion Pumping Engines. Figs. 332 and 333 show different applications of steam traps to the receiver coils and jackets of triple expansion pumping engines. The drawings are self-explicit. SEPARATOR tn FEED TANK Fig. 333. Drainage System for Jackets and Receivers of Triple Expansion Pumping Engines. 319. Drips under Vacuum. — Conditions frequently make it neces- sary to remove condensation from apparatus working under a vacuum, as, for example, a primary heater. The simplest method is to pipe the drips to the condenser and per- mit the condensation to gravitate to it as in Fig. 334. Where this is impracticable, as in an installation with the condenser above the 598 STEAM POWER PLANT ENGINEERING heater, a steam trap is usually employed. Fig. 335 shows the applica- tion of a Bundy trap to a vacuum or primary heater. A close-fitting weighted check valve W, set to open outwards prevents intake of air through the discharge pipe while the trap is filling. Connection E is made from the vent underneath the valve stem V back to the heater so as to equalize the pressures. The operation is as follows: Condensation gravitates from the heater through check C to the body of the trap, the check Fig. 334. Gravity Drainage ; Vacuum Heater. W being closed. When the bowl is full enough to overcome the weight of the counterbalance, it sinks and opens up the live-steam valve V. This admits steam to the trap through pipe D, which in Fig. 335. Method of Draining Heater under Vacuum. turn closes check C and forces the water past the weighted . check W to the discharge tank. After the water is discharged the bowl returns SEPARATORS, TRAPS, DRAINS 599 to its original position and closes valve V, the weight closes check W, the vent check equalizes the pressure in the bowl and heater, and condensation gravitates to the trap again. 320. Drips under Alternate Pressure and Vacuum. — Occasionally the load on an engine is of such a character that the pressure in the receiver alternates from a pressure of 30 or 40 pounds absolute to a vacuum of varying degree. Where the periods of vacuum operation are very few and of short duration, as in the average installation, no attention is paid to the vacuum and the condensation is removed by a trap in the ordinary way. If, however, the periods are of sufficient duration and frequency, the ordinary method is not applicable and the arrangement shown in Fig. 336 may be used. The trap is placed Fig. 336. Method of Draining Receivers under Alternate Vacuum and Pressure. below the receiver as indicated. The delivery pipe is provided with a weighted check or resistance valve W set so as to open outwards from the trap, also a spring water relief valve R. Another weighted check P is placed in the line leading from the vent to the atmosphere, and a plain check C in the line leading back into the receiver. This arrange- ment of valves permits the venting of the trap after discharge and effectually excludes air from the trap when there is less than atmos- pheric pressure on the receiver. With the relief valve set to open at a 600 STEAM POWER PLANT ENGINEERING pressure in excess of the maximum receiver pressure it acts as a " stop " in the pipe and the water must enter the trap. When the trap discharges, the live steam supplied through the pipe attached to the steam valve forces the water through the weighted check and relief valves into the sewer or receiving tank. When working with a vacuum, the pressures in receiver and trap are equalized through the vent con- nection and the condensation flows into the trap by gravity. The operation of discharge is the same as in the case of pressure. 321. The Steam Loop. — Fig. 337 illustrates the principles of the " steam loop " for automatically returning high-pressure drips to the \ HORIZONTAL "7- cS kZ o5 ^. ===L—. ; WATER LEVEL: Fig. 337. General Arrangement of the Simple " Steam Loop.*' boiler. In the figure the loop is returning the condensation from a steam separator to a boiler above the level of the separator. The apparatus is very simple, consisting of a horizontal and two vertical lengths of plain pipe placed as indicated. Pipes R and B may be cov- ered but " horizontal " A is left uncovered, as its function is that of a condenser. The operation is as follows: Circulation is first started by opening stop valve at the bottom of the drop leg until steam escapes. The valve is then closed and the steam in the horizontal A condenses and gravitates to the drop leg B. On account of the slight reduction in pressure in the horizontal a mixture of spray and steam flows from the separator chamber to the horizontal, and, condensing, gravitates to the drop leg. The column of water in the drop leg rises until its static head balances the difference in pressure in the riser R and the horizontal. In other words, a decrease in pressure in the horizontal produces similar effects on the contents of the riser and drop leg but in a degree inversely proportional to their densities. SEPARATORS, TRAPS, DRAINS 601 602 STEAM POWER PLANT ENGINEERING Any further accumulation causes an equal amount to pass from the bottom of the column to the boiler, since the pressure in the boiler is then less than that at the bottom of the column, that is, the steam pressure on the top of the water column plus the hydrostatic head H is greater than the pressure in the boiler. Once started the process is continuous and requires no further attention. 322. The Holly Loop. — In the application of the steam loop where many points requiring drainage are connected to many boilers and conditions are more complex, some method other than the simple one of radiation may be advisable to secure the necessary lower pressure at the top of the loop. Such a method is illustrated in Fig. 338. This arrangement differs from the simple loop in that all condensation first gravitates to a " Holly " receiver (shown in detail in Fig. 339) before passing into the " riser." The receiver is placed be- low the lowest point to be drained and serves as a storage for large or unusual quantities of water and en- ables the riser to act at a constant rate independent Fig. 339. Holly Receiver. Q f var i a ble discharge into the receiver. Furthermore, the lower pressure in the discharge cham- ber necessary to secure the lifting of the mingled steam and water through the riser, instead of being created by condensation as in the simple loop, is produced by a reducing valve B discharging into the feed-water heater. The operation of the Holly loop is as follows: Circulation is started by opening valve D until steam appears. Valve D is then closed and the reducing valve is put into commission. Condensa- tion from separators, traps, and pipes gravitates to the " receiver," from which it is forced into the " riser " in the form of a spray. The spray- ing effect is produced by a series of holes drilled in pipe A, Fig. 339. From this receiver the spray and moisture rise to the " discharge chamber/' on account of the lower pressure at that point, where the steam and entrained water are separated, the water gravitat- ing to the bottom of the chamber and thence to the drop leg, and the steam discharging through the reducing valve into the heater. The principles of operation are exactly the same as in the simple steam loop. 323. Returns Tank and Pump. — Low-pressure drips in connection with heating systems may be returned to the boiler along with the condensation from the heating system by a combined pump and receiver SEPARATORS, TRAPS, DRAINS 603 as shown in Fig. 340. The height of water in the tank controls the operation of the pump through the medium of a float and throttle valve. This combination of float and balanced throttle valve is sometimes called a " pump governor." In the illustration the pump forces the returns through a closed heater before delivering them to the boiler, Drip Trapped •to Sewer ^^rr Fig. 340. Returns Tank and Pump. though they are oftentimes returned directly. The tank is vented to the atmosphere to prevent it from becoming " air bound." The cold- water supply or make-up water is sometimes discharged into the receiv- ing tank as indicated. With open heaters the cold supply is ordinarily controlled by a float within the heater itself. 324. Office Building Drains. — In the power plants of tall office build- ings the public sewers are often above the basement level, and it is necessary to remove all liquid wastes mechanically. The Shone pneumatic ejector has been found to serve this purpose effectually. This apparatus is placed in a pit in the basement floor into which all sewage, drips from engines, washings from boilers, and ground water gravitate, and are automatically discharged into the street sewer by means of compressed air. 604 STEAM POWER PLANT ENGINEERING Fig. 341 gives a sectional view of a Shone ejector of ordinary con- struction. It consists essentially of a closed vessel furnished with inlet and discharge connections fitted with check valves, A and B, opening in opposite directions with regard to the ejector. Two cast- iron bells, C and D, are linked to each other, in reverse positions, the *•« EXHAUST PI fl A,R SUPPLV Fig. 341. Shone Ejector. rising and falling of which control the supply of compressed air through the agency of automatic valve E. The bells are shown in their lowest position, the supply of compressed air is cut off from the ejector, and the inside of the vessel is open to the atmosphere. The sewage gravitating into the ejector raises the bell C, which in turn actuates the automatic valve E, thereby closing the con- nection between the inside of the ejector and the atmosphere and open- ing the connection with the compressed air. The air pressure expels the contents through the bell-mouthed opening at the bottom and the discharge valve B, into the main sewer. Discharge continues until the level falls to such a point that the weight of the sewage retained in the bell Dis sufficient to pull it down, thereby reversing the automatic valve. This cuts off the supply of compressed air and reduces the pressure to that of the atmosphere. SEPARATORS, TRAPS, DRAINS 605 The positions of the bells are so adjusted that compressed air is not admitted until the ejector is full, and is not allowed to exhaust until emptied down to the discharge level; thus the ejector discharges a fixed quantity each time it operates. Two ejectors, each of a capacity suitable for handling the average flow of tributary sewage and so arranged that they can work either independently or together, are usually installed at each ejector station. The main sanitary sewer of the building usually discharges directly into the ejectors, the surface water, drips, etc., being collected in a neighboring sump. The latter is connected to the sanitary sewer through a trap or back-water valve. CHAPTER XV. PIPING AND PIPE FITTINGS. 325. General. — The advent of high pressures and superheat is responsible for the elimination of many of the older systems of piping, the tendency being towards greater uniformity in design, particularly in electric central station work. In isolated stations the conditions of operation and installation are so variable that each case presents an entirely different problem. In any system of piping the fundamental object is to conduct the fluid in the safest and most economical manner. The material should be the best obtainable and the system so flexi- ble that a break-down in one element will not necessitate the closing down of the entire plant. On the other hand, flexibility increases the number of parts and, unless first cost is of little importance, tends to weaken the system as a whole. It is a safe general proposition to say that the best pipe and fittings, irrespective of first cost, will prove the most economical in the end, but few owners of power plants are will- ing to take this view. 326. Drawings. — An assembly drawing of the entire installation giving the location of all valves and fittings is necessary in order to avoid interference, and particularly where a number of fittings are to be close together. Detailed drawings should also be provided of each division of the piping to facilitate installation, as, for example, the high-pressure steam, the exhaust steam, the feed water, the condensing water, the oil, the heating and the sanitary piping. As a rule, lower and more uniform bids will be obtained from an isometric or perspec- tive sketch, as in Fig. 254, than from conventional plan and elevation drawings, due, no doubt, to the greater ease with which the drawing is interpreted. A complete set of specifications for a piping system is given in paragraph (415), and illustrates the usual practice along this line. 327. Materials for Pipes and Fittings. — The following materials are used in the construction of pipes for steam, water, and gases. Average Bursting Tension. Low-carbon or mild steel 65,000 lbs. per sq. in. Wrought iron 50,000 lbs. per sq. in. Cast iron, high grade 20,000 lbs. per sq. in. Cast steel 50,000 lbs. per sq. in. Wrought copper 33,000 lbs. per sq. in. Brass 18,000 lbs. per sq. in. Special alloys and compounds 15,000-60.000 lbs. per sq. in. 606 PIPING AND PIPE FITTINGS 607 Mild Steel. — The greater portion of the piping in the average steam power plant is of mild steel, lap or butt welded for high pressures and riveted for very low pressures and large diameters. Steel pipe is con- siderably cheaper than that manufactured from other material, and fulfills practically all requirements for general service. Wrought Iron. — Pipes manufactured from puddled iron, though not as strong in ultimate bursting strength as low-carbon steel pipes, are superior in many ways. Threads are cut more readily and with less power. It is more easily bent without injury and offers more resistance to corrosion. The life of wrought-iron pipe is greater than that of steel pipe under conditions of extreme exposure, or when buried under- ground, and it is recommended in all cases where corrosion is apt to be severe, as in blow-off pipes, drips, and drains or in pipes not in continuous use. Steam pipes well covered and protected from external moisture are ordinarily of mild steel, as the conditions do not warrant the use of wrought iron, which costs approximately 10 per cent more. Since the term " wrought-iron pipe " is used rather loosely in practice, the manu- facturer will generally furnish mild steel unless special stress is laid upon the term " wrought iron." Wrought Iron vs. Steel Pipe: Ir. Age, March 2, 1905, Jan. 18, 1906; Pro. Heat and Vent. Engr., Jan., 1900; Mach., Dec, 1903, p. 191; Am. Mfg. and Ir. World, April 29, 1898; Eng. News, 50-292, 296, 487, 51-62; Eng. Rec, 44-54; Locomotive, Jan. 1, 1906. Corrosion of Pipes: Trans. A.S.M.E., 18-282; Eng. Rec, 42-194, 43-354, 45-584. Cast-iron Pipes. — Cast iron is little used for high-pressure steam, except occasionally in the construction of headers where a number of branches lead into a single pipe, in which case the number of joints is greatly reduced and the cost considerably less than for wrought-iron or steel pipe with numerous fittings and joints. The chief objections to cast iron for high-pressure steam are its weight and lack of homogeneity. It is most used in connection with water service and sanitation. Cast-iron Pipe: Jour. New Eng. Waterworks Assn., March, 1907; Engr., Lond., Nov. 7, 1902, p. 454; Power, Jan., 1904, p. 334; Eng. News, 46-216, 48-193, 51-544. Cast-Steel Pipe. — Cast-steel headers are sometimes used in power plants for highly superheated steam, since the material is not affected by high temperatures to the same extent as cast iron. High first cost and the difficulty of securing castings free from blow-holes have pre- vented its more general use. Copper Pipes. — Copper steam pipes were in common use for many years in marine service on account of their flexibility. To increase the bursting strength, pipes above 6 inches in diameter were generally wound 608 STEAM POWER PLANT ENGINEERING with a close spiral of copper or composition wire. In recent years wrought-iron and steel pipe bends have practically superseded copper for flexible connections. As a rule the use of copper pipes should be avoided on account of the rapid deterioration of the metal under high temperatures and stress variations. The cost is prohibitive for most purposes and this alone prevents it from being seriously considered in the manufacture of pipe. Copper Pipes: Engr., Lond., April 15, 1898, p. 360, Aug. 11, 1893; Engng., April, 1898; Eng. Rec, July 30, 1898. Brass Pipes. — Brass is little used in the construction of pipes on account of its high cost. It withstands corrosive action much better than iron or steel and is often used in connecting the feed main with the boiler drum. Special alloys, nickel steel, " ferrosteel," malleable iron, and the like have been used in the manufacture of pipes, and possess points of superiority over wrought iron and steel for some purposes, as for highly superheated steam, but the cost is prohibitive for average steam power plant practice. Materials for Fittings. — Elbows, tees, flanges, and similar fittings are usually made of cast iron, malleable iron, or pressed steel, though cast steel, " ferrosteel," and other steel compounds are used to a limited extent. Standard cast-iron fittings are recommended for ordinary pressures of 100 pounds per square inch or less, and extra heavy cast- iron fittings for higher pressures. Malleable iron fittings are lighter and neater than cast iron and are extensively used for small sizes of steam and gas pipe. Manufacture of Pipe: Sci. Am., Dec. 12, 1903; Mach., Feb., 1904, Dec, 1903, p. 191; Eng. News, 50-232, 296. 328. Size and Strength of Commercial Pipe. — Wrought-iron and mild-steel pipe are marketed in standard sizes. Those most commonly used in steam power plants are designated as (1) Merchant or standard pipe. (2) Full weight pipe. (3) Large O. D. pipe. (4) Extra heavy. (5) Double extra heavy. Table 76 gives the dimensions of standard " full-weight " pipe, which is specified by the nominal inside diameter up to and including 12 inches and based on the Briggs standard. Pipes larger than 12 inches are designated by the actual outside diameter (O. D.), and are made in various weights as determined by the thickness of metal speci- PIPING AND PIPE FITTINGS 609 fied. Manufacturers specify that " full weight " pipe may have a variation of 5 per cent above or 5 per cent below the nominal or table weights, but merchant pipe, which is the standard pipe of commerce, such as manufacturers and jobbers usually carry in stock, is almost invariably under the nominal weight. It varies somewhat among the different mills, but usually lies between 5 and 10 per cent under the table weight. The smaller sizes of merchant pipe, J inch to 3 inches, are butt- welded and the larger sizes are lap-welded. Extra heavy and double extra heavy pipe have the same external diameter as the standard, but are of greater thickness and hence the internal diameter is smaller. Taking the thickness of the standard pipe as 1, that of the extra heavy is approximately 1.4 and of the double extra heavy 2.8. Wrought-iron and steel pipes are ordinarily designed with factors of safety of from 6 to 15, with an average not far from 10. The standard hydrostatic tests to which the various pipes are subjected at the mills are as follows: Hydrostatic Pressure, Lbs. per Sq. In. Standard, butt-welded, |-3 in 600 to 1,000 Standard, lap-welded, 3-12 in 500 to 1,000 Extra heavy, butt-welded, f-3 in 600 to 1,500 Extra heavy, lap-welded, 1^—12 in 600 to 1,500 Double extra heavy, butt-welded, J- 2 J in 600 to 1,500 Double extra heavy, lap-welded, 1^-8 in 1,200 to 1,500 The pressure necessary to burst piping is far above anything likely to occur in ordinary practice on account of the thickness of material necessary to permit of threading, thus: Actual Bursting Pressure, Lbs. per Sq. In. 2-inch " standard " mild-steel pipe* 5,800 2-inch " wrought-iron pipe 4,106 10-inch " mild-steel pipef 3,000 10-inch " wrought-iron pipe t 1,900 10-inch " extra heavy " wrought-iron pipe t 2,700 * Machinery, December, 1903, p. 192. f Crane Company, Published Tests. Riveted Pipes. — .For low pressures and large diameters, pipes are constructed of thin sheets of boiler steel with riveted joints, the seams being either longitudinal and circumferential, or spiral. Such pipes are not necessarily limited to large sizes and low pressures, though this is the usual practice. Pipe fittings are classed as screwed or flanged. 610 STEAM POWER PLANT ENGINEERING 329. Screwed Fittings, Pipe Threads. — For screw connections the ends of pipes and fittings are threaded to conform to the Briggs or United States standard system, as shown in Fig. 342. The end of the Fig. 342. Standard U. S. Pipe Thread. pipe is tapered 1 to 32 with the axis, the angle of the thread being 60 degrees and slightly rounded at top and bottom. The proper length of perfect threads is given by the formula T = (0.8 D + 4.8) (146) in which T = length in inches. D = actual external diameter of the tube, inches. n = number of threads per inch. The imperfect portion of the thread is simply incidental to the pro- cess of cutting. The object of the taper is to facilitate " taking hold " in making up the joint. Table 76 gives the number of threads per inch for various sizes of standard pipe. When properly constructed a screwed joint will hold against any pressure consistent with the strength of the pipe. For example, the ultimate bursting strength of a " standard" 2-inch pipe is about 6000 pounds per square inch, while the stripping strength of the joint (with perfect threads) is 225,000 pounds. The threads, however, are often poorly cut and the parts screwed together improperly cleaned and lubricated, thus causing leakage between the threads. 330. Flanged Fittings. — In cast-iron pipes, valves, tees, and other fittings the flange is always a part of the casting, but for joining the two ends of a steel or wrought-iron pipe the flanges may be fastened to the pipe in a number of ways. Fig. 343, A to H, illus- trates methods most commonly used. In A to C the pipes are screwed into cast-iron or forged-steel flanges and the two faces, with metallic or composition gasket between, are drawn together by bolts. A illustrates the most common and inexpensive of flanged joints, which requires no special tools and can be made up at the place of PIPING AND PIPE FITTINGS 611 •Maiog jo qoui jad sp-eaaqx P Jaqnin^ NO0O0'*'*'HiHrHrH»0000O0O000Q00000O000QOO0 C0J,-|T-l»-(,-ir-l1-ll-lrH Nominal Weight per Foot. w d P T((fq»OCOH©Tl*NOCOC<30«0010«DNNO©INOO N^iOOOHOlNttiONiOOCO^iOMNHNOOO) HHNMCOlONOlOINTtlOOMOOeoOiOOO o d §-& taining One Cubic Foot. »OO^H Nr-ilM OOMOOOKN- IN CO WN©lOMl'*mC0M rjlONfO^mMINrtrtrHrH T-l 1—1 Is H «1 1 lONNN^H OOQOrHiOOl'fNNrHCONiOlOO! T(lNiOniP300HOIM05iO'>t Tj<0«OiO«OOSCOOCOe000500t^- Oi CO - ■* iH-*tOO>(N'OOON.'^OOM05'<4lTtltD'*CDi© ^ H OS NiM©-*M05C0OsNO'*NNNiH00lM00M(NON OHHMCO^tONONINfOHtDCOlOOsnOOs^iO HrHM(N«IW*>0©(»OnMrti 1 ! % d h- ( m MiHNOOM© N^rH^c<3(N«O00tO^H»N i— i O0O0 Oi CO 00 l00050M!OOSM1000CCQOM<00500CO^«COe«:0) OHHCCOOO^OnNCOOONOSOSOONONOOOO H(N05^NO)nffi0000O(N00SOC<3 lHHlHMmiOQ0-*«O00^iO CS|r-icO'!jHiO<£>(MTtl CO rt< J>- (N(NiOiO«:iO!OC«3COOSINOOCOON©(N©«3CON HNCq^COSCHOOS^^iOKKNOOON HHHNMHHONOSON i I d 00TtKNN.OXNiO^rtlCOCOCO00(Nai^f0©CON00 •>*rJllOiO00O3C0©ailOC0^'*©'*iO©NNNir3 OOHiOOSWHl005'HlMCOCONOONCOlOtOOO(Ni*iO NOSINMOSMr-lffitOCOaitOMONr-ilOOSMNFHlO NOHtOINHNOSTliOanOHNTtlOOOsOINNOBO Hi-lfqNM'HiO>ONOJO(N'*iONOWNOM©0 •ssaroptqi, Ibuioio^i I d OOOOiHOSW* lfl-*-*NtON(0 0) .-< (N "* «D ©000SOH«'*^>0O>H(NC0^i000O(N'*© OOOHHrtHHHMINNNINMNCOMmP: o> "S s [ Approxi- mate Internal Diameter. ■8 d i— i T^^tleO^OO THNOONQOCO»lOlOM(NNai NCOfflNN^OOrHtOcOcOtiMO^CINOOCOrH NM^COOOMOO^OiOOiOOOOOSOSO HHHNeq««^^iiotoNNxoH C<1 Actual External Diameter . to > TtH lO i— i CO t^b- ONMNNiflifliC "*lO©QOOCO!OGMi)Om >0 lO©«©«ONNN HHrtrHCqiNCCOt^lflKJONOOOSOHCq 'imnaiui imnoio^ © d rHlOO H-* «*» Hn *« rtH»H|N W|(M ~<| P 02 » g t_ u 33 _• &.S £> ft . £n Q 3.01 3.25 3.18 3.10 3.13 2.78 2.71 * C. P. Paulding, Stevens Indicator, Vol. xix, p. 388. t Outside diameter. 616 STEAM POWER PLANT ENGINEERING 331. Coverings. — Steam pipes, feed-water pipes, boiler steam drums, receivers, separators, etc., should be covered with heat-insu- lating material to reduce radiation losses to a minimum. For most practical purposes the loss of heat from a bare steam pipe or drum may be taken as 3 B.T.U. per square foot per hour per degree differ- ence in temperature, Table 80. The actual loss depends upon the diameter of the pipe, on its position whether vertical or horizontal, the nature of the surface, and the velocity of the surrounding air currents. For a detailed analysis of these various influences, and interesting information on the transmission of heat, the reader is referred to Paulding's " Steam in Covered and Bare Pipes." By properly applying any good commercial covering, from 75 per cent to 90 per cent of the heat loss may be prevented. (See Fig. 344 and Table 81.) Diagram *. Heat Fig. 344. Efficiency of Pipe Coverings. Example : Required the saving per annum due to covering a pipe 10 inches in diameter and ltfO feet long; steam pressure 150 pounds; average temperature of the air 76 degrees F. ; cost of covering applied 65 cents per running foot; efficiency of covering 85 per cent; cost of coal $2.50 per ton; plant to operate 14 hours per day and 300 days per year. The temperature of steam at 150 pounds pressure = 366 degrees F. PIPING AND PIPE FITTINGS 617 Difference of temperature between the steam and air = 366 — 76 = 290 degrees F. Loss per square foot per hour, bare pipe = 3 X 290 = 870 B.T.U. Loss per square foot per day, bare pipe = 870 X 14 = 12,180 B.T.U. Loss per square foot per year, bare pipe = 12,180 X 300 = 3,654,000 B.T.U. 100 lineal feet of 10-inch pipe has an external surface of 282 square feet. Therefore the loss per year from the bare pipe is 282 X 3,654,000 = 1,030,000,000 B.T.U. (approx.). TABLE 81. EXPERIMENTS ON STEAM-PIPE COVERINGS. ("Condensation of Steam in Covered and Bare Pipes " [Paulding].) Kind of Covering. Hair felt Do Remanit for interme- diate pressure. Remanit for high pres- sure. Mineral wool Champion mineral wool Rock wool Asbestos sponge felted Do Do Magnesia Do Do Do Do Do Asbestos, Navy Brand Do Do Manville sectional. . . Do Do Asbestos air cell Do Asbestos fire felt Do Do Fossil meal Riley cement Diam. Thick- of Test ness of Pipe, Cover- Inches. ing, Inches. 2 0.96 8 0.82 2 0.88 2 1.30 8 1.30 8 1.44 8 1.60 2 1.125 10 1.375 2 1.14 4 1.12 2 1.09 8 1.25 2 1.08 2 1.00 10 1.19 2 1.20 2 1.125 10 1.375 8 1.70 2 1.31 4 1.25 4 1.12 2 0.96 8 1.30 2 1.00 2 0.99 8 0.75 8 0.75 Temperatures F. Steam 302.8 348.3 304.5 306.6 344.1 346.1 344.1 364.8 364.8 309.2 388.0 354.7 344.1 310.9 365.2 365.2 309.2 365.2 365.2 345.5 354.7 388.0 388.0 303.3 344.7 354.7 307.4 347.1 347.9 Air. 71.4 69.0 73.3 76.1 58. 74. 63.0 60.7 62.8 79.4 72.0 80.1 66.3 81.6 64.6 66.0 79.4 64.6 78 80 72 72 72 79 80 72.5 75.3 74.3 B.T.U. per Hour per Square Foot of Pipe Surface. Total. 89.6 117.9 100.3 83.7 81.3 86.1 72.0 145.0 85.0 59.7 147.0 155.8 106.6 69.8 155.0 103.0 69.9 176.0 112.0 93.4 157.0 143.0 166.0 165.5 133.5 198.0 180.0 238.0 260.0 Per Degree Differ- ence. 0.387 0.422 0.434 0.363 0.284 0.317 0.256 0.477 0.248 0.260 0.465 .567 ,384 304 ,515 ,347 ,304 .585 0.375 0.394 0.572 0.453 0.525 0.716 0.502 0.721 0.766 0.876 0.950 Date of Test 1901 1894 1901 1901 1894 1894 1894 1901 1901 1901 1896 1896 1895 1901 1901 1901 1901 1901 1901 1894 1896 1896 1896 1901 1894 1896 1901 1894 1894 Testing Ex- pert. Jacobus Brill Jacobus Jacobus Brill Brill Brill Barrus Barrus Jacobus Norton Paulding Brill Jacobus Barrus Barrus Jacobus Barrus Barrus Brill Paulding Norton Norton Jacobus Brill Paulding Jacobus Brill Brill 618 STEAM POWER PLANT ENGINEERING Assuming a net available heat value of 10,000 B.T.U. per pound for the coal, the equivalent coal consumption is 51.5 tons, valued at 51.5 X $2.50 = $128.75. The covering will save 85 per cent of this, or $109.50 per annum. The pipe covering applied will cost 100 X 0.65 = $65.00. In this case the covering will pay for itself in considerably less than a year. Pipe covering is applied in sections molded to the required form and held to the pipe by bands, or may be applied in a plastic form. The former is more readily applied and removed, and is usually adopted for pipes, while the valves and fittings are sometimes covered with plastic material. Piping should be tested under pressure before being covered, since leaks destroy the efficiency and life of the cover- ing. If the surrounding atmosphere is moist the covering should be given two or three coats of good paint. Coverings are sometimes applied to cold-water pipe to prevent sweating in a humid atmosphere. Pipe Coverings: Power, July, 1904, p. 407, Aug., 1904, p. 482, May, 1903, p. 239, Dec, 1901, p. 32; St. Ry. Jour., Nov. 29, 1902, p. 875; Engr., Lond., May 27, 1904, p. 547; Eng. Review, Nov., 1898, p. 15; Am. Elecn., May, 1903; Engng., Aug. 7, 1903; Mech. Engr., Nov. 25, 1905; Elec. World and Engr., April 6, 1901; Stevens Ind., Oct., 1902; Trans. A.S.M.E., 16-827, 23-791. Identification of Power House Piping by Colors: Power & Engr., Apr. 26, 1910, p. 752. 332. Expansion. — One of the most difficult problems in the design of a piping system is the proper provision for expansion and contrac- tion due to change in temperature. If a pipe is immovably fixed at both ends and under no strain when cold, and the temperature is increased, as by the admission of steam, it is subjected to a com- pression proportional to the rise in temperature (within the elastic limit). For example, a 6-inch standard extra heavy wrought-iron pipe 200 feet long at 66 degrees F., if heated to 366 degrees F. (the temperature corresponding to steam at 165 pounds per square inch absolute pressure), will exert an axial force of P = EA (t t — t)ft. (Mechanics of Engng., Church, p. 218.) (147) P = force in pounds. E = modulus of elasticity, 30,000,000. t t = final temperature, degrees F. t = initial temperature. jt* = coefficient of expansion, 0.0000075. A = sectional area of the pipe material, 8.5 square inches. Hence P = 30,000,000 X 8.5 (366 - 66) 0.0000075. = 573,750 pounds. PIPING AND PIPE FITTINGS 619 Unless well braced throughout its entire length the pipe will buckle and become distorted. If free to expand its length would increase. The temperature of the pipe is always less than that of the steam on account of radiation from the outer surface and varies with the effi- ciency of the covering. But ignoring radiation the increase in length is 1 = n(t x -t) L, (148) in which I = increase in length, inches. L = length of pipe, inches. Other notations as in (147). Substituting in (148), t t = 366. t = 66. Pl = 0.0000075. L = 2400. I = 0.0000075 (366 - 66) 2400 = 5.4 inches. Since the forces produced by expansion are practically irresistible, the pipe is invariably allowed to expand freely by suitable means so as not to strain the connections. The coefficients of expansion per degree difference in temperature for various pipe materials are given in Table 82. Headers less than 50 feet in length usually require no special pro- visions for expansion provided the ends are free and the leads to and a b «^^ c o Fig. 345. Types of Expansion Pipe Bends. from the header are not too short, the pipe usually being anchored at the middle and permitted to expand in either direction. Free expan- sion of the feeders may be provided for (1) By long radius bends, as in Fig. 345. (2) By double swing screwed fittings, as in Fig. 346, or (3) By packed expansion joints, Fig. 347. Where practicable the long radius bends will prove most satis- factory. The radius of the bend should not be less than 5 diameters of the pipe, and larger if possible. The length of straight pipe at the 620 STEAM POWER PLANT ENGINEERING end of each bend should not be less than twice the diameter of the pipe measured from the face of the flange. On account of the great strains to which the joints of pipe bends are subjected, the welded joint, G, Fig. 343, is recommended as giving the best results. The next best is the lap joint, F, Fig. 343. TABLE 82. COEFFICIENTS OF LINEAR EXPANSION PIPING MATERIALS. Material. Wrought iron and mild steel. . . Wrought iron Cast iron Cast steel Hardened steel . . . Nickel-steel, 36 per cent Nickel Copper, cast Copper, wrought Lead Cast brass Brass wire and sheets Tin cast Tin hammered Zinc cast Zinc hammered Temperature Range. Mean Coeffi- cient per De- gree F. 32-212 0.00000656 32-572 0.00000895 32-212 0.00000618 32-212 0.00000600 32-212 0.00000689 32-572 0.00000030 32-212 0.00000955 32-572 0.00001092 32-212 0.00001580 32-212 0.00001043 32-212 0. 00001075 32-212 0.00001207 32-212 0.00001500 32-212 0.00001633 32-212 0.00001722 LINEAR EXPANSION OR CONTRACTION OF CAST IRON IN INCHES PER 100 FEET, — DEGREES F. Temperature Difference. Expansion. Temperature Difference. Expansion. 100 150 200 250 0.72 1.1016 1.5024 1.9260 300 400 500 600 800 2.376 3.360 4.440 5.616 7.872 Multiply by 1 .1 for wrought mild steel. Multiply by 1 .5 for wrought copper. Multiply by 1.6 for wrought brass. Fig. 345, A, B, C, D shows applications of pipe bends to straight pipe runs. A is the cheapest and most common arrangement for all sizes of pipe. B is a modification for limited center to center spaces. C shows a common method of taking up expansion in straight runs of pipe of very large diameters where the space requirements prohibit the use of PIPING AND PIPE FITTINGS 621 a single U bend. Here the main runs are connected to manifolds which in turn are connected by a number of small U bends, the equivalent areas of which correspond to that of the large pipes. This makes a more flexible connection than if a single U bend were used. The arrange- ment D does away with the elbows required in A, but is not applicable to pipes over 8 inches in diameter. Figs. 357 and 358 show applications of pipe bends to boiler and header connections. Fig. 346 shows a double swing screwed joint in which expansion causes the fittings to turn slightly and thus relieve the strain. This method is usu- ally adopted where long ra- dius bends are not practicable on account of lack of space and where screwed fittings are used. Slip joints, Fig. 347, are now little used except with very large pipes and where space prohibits long radius bends. When slip joints are employed the pipe must be securely anchored to prevent the steam pressure from forcing the joint apart and at the same time permit the pipe in expanding to work freely in the stuffing box. Sagging of the pipe on either side, which might cause binding in the joint, is prevented by suitable supports. Expansion in Steam Pipes: Power, July, 1906, p. 426, Jan., 1904, p. 30, March, 1904, p. 160, Oct., 1904, p. 609, Dec, 1900; Am. Elecn., 10-432; Engr., U.S., Feb. 1, 1904, p. 125; Eng. News, 44-194, 47-468, 50-487; Power, June 2, 1908. FRONT ELEVATION Fig. 346. "Double-Swing Expansion Joint. Fig. 347. Slip Expansion Joint. 333. Pipe Supports and Anchors. — Pipe lines must be supported to guard against excessive deflection and vibration. Supports are con- veniently classified as (1) hangers, (2) wall brackets, and (3) floor stands. Fig. 348 illustrates a type of hanger for suspending pipes from I beams. The supports being free to swing, no provision for expansion is necessary. A properly designed hanger may be readily removed without disturbing the pipe line, and should be adjustable to facilitate " lining up." If of rigid construction the lower end should be provided with a roller. Fig. 349 gives the details of a wall bracket with rolls and roll binder. 622 STEAM POWER PLANT ENGINEERING Supports adjacent to long radius bends should be provided with roll binders as illustrated to prevent the pipe from springing laterally, but Fig. 348. A Typ- ical Pipe Hanger. Fig. 349. A Typical Wall Bracket with Binding Roll. Fig. 349a. A Typ- ical Floor Stand. they may otherwise be omitted. The rollers are often made adjustable to facilitate lining up. Fig. 349a illustrates a typical floor stand. Pipe lines are usually securely anchored at suitable points in a manner similar to that illustrated in Fig. 350, the pipe rest- ing on a saddle and being rigidly clamped to the bracket by a flat iron band with ends threaded and bolted. This limits expansion to one direction and prevents excessive strain on the fittings. 334. General Arrangement of High- Pressure Steam Piping. — The general ar- rangement of piping depends in a great measure upon the space available for en- gines and boilers. The engine and boiler room may be placed (1) Back to back, Fig. 361, (2) End to end, Fig. 351, (3) Double decked, Fig. 356. Fig 350. A Typical Pipe Anchor. PIPING AND PIPE FITTINGS 623 The back to back arrangement is the most common and, other things permitting, is to be preferred on account of the short and direct connection between engines and boilers and the ease of enlarge- ment. The engine and boiler rooms are separated by a wall, and as much of the piping as possible is located in the boiler room. The end to end arrangement is ordinarily limited to situa- tions where the distribution of space precludes the back to back system. The double decked arrangement is frequently used where ground space is limited or expensive. Engines and boilers are con- nected in a variety of ways through steam headers as shown in the following examples : (1) Single header, Fig. 361, (2) Duplicate headers, Fig. 352, (3 ) Loop or ring header, Fig. 353, (4) The "unit" system, Fig. 357. The single header system is per- haps the most common, since it embodies simplicity, low first cost, and provision for extension. The duplicate system is losing favor, since experience shows that the extra cost of the du- plicate mains will usually give better returns in continuity of operation and maintenance if in- vested in high-grade fittings on a single pipe system. The loop header is well adapted for the power plants of tall office build- DZ4 STEAM POWER PLANT ENGINEERING i* ii nf 1 \1\£ fr 11 Wf >— f-i 1 c w mETPI H i 1? II w UJ *< Fig. 352. Typical Duplicate Header System. VALVE ©: ' rt --) Fig. 353. Typical Loop Header. PIPING AND PIPE FITTINGS 625 ings, Fig. 355, in which a large number of steam engines, elevator pumps, air compressors, and miscellaneous steam-consuming appliances are crowded together in a comparatively small space. Large modern power plants are, by the latest practice, divided into complete and independent units, as in Fig. 440, each prime mover having its own boiler equipment, coal and ash-handling machinery, feed pumps, and piping, operated independently of the rest of the plant, though provision is made whereby any boiler equipment may provide steam for any prime mover. D sas- 3 li rm=° ; wgr Fig. 354. Typical By-Pass System. The power plant of the Manhattan Elevated Railway Company, New York, is practically divided into eight sections each consisting of an engine and eight boilers, the boilers being " double decked " (Fig. 356). The branch pipes from the upper and lower batteries lead into 18- inch headers, the steam from each being conducted to a receiver reservoir 36 inches in diameter and 20 feet long in the engine room basement directly behind each engine, from which the two high-pressure cylinders are supplied. Gate valves are used in each boiler branch, one close to the boiler and another near the header, and also in the steam pipes near the reservoir. The steam headers for each of the eight units are connected by a main which equalizes the pressure and allows a deficiency in one unit to be made up from the others. Figs. 357 and 358 show the general arrangement of the steam piping at the Yonkers power house of the New York Central. The turbines 626 STEAM POWER PLANT ENGINEERING PIPING AND PIPE FITTINGS 627 »WWMtW»Wi»»)»»»>» »»»M>»»M>M »»M^^^ CNCINC ROOM %X LX h hf to i P EI I; £l kr -r Jg 'torn i_y 628 STEAM POWER PLANT ENGINEERING a qnp a .. ( Detail Plan (Power) Fig. 358. Details of Boiler Steam Piping, Yonkers Power House of the New York Central R.R. PIPING AND PIPE FITTINGS 629 are connected in pairs by 14-inch loops, each turbine taking steam from either of two banks of four boilers. The high-pressure steam piping is of mild steel with modified reenforced "Van Stone" joints. The high-pressure valves are of the split disk pattern with semi-steel bodies. Expansion is taken up by the long sweep bends. Fig. 359. Overhead Boiler Piping, Quincy Point Power Plant of the Old Colony St. Ry. Co., Quincy Point, Mass. Plants using superheated steam are ordinarily piped to supply saturated steam to the auxiliaries as illustrated in Fig. 359. The boiler branch E, leading to the main header, normally supplies super- heated steam to the engines. C is an auxiliary main supplying the air pumps, stoker engines, and other auxiliaries with saturated steam from branch pipe D. Steam Piping: Power, Feb. 23, 1909, Nov., 1905, p. 683, April, 1904, p. 213, Feb., 1904, p. 90, Sept., 1904, p. 540, Nov., 1904, p. 677, July, 1903, p. 356; Ir. Tr. Rev., May 29, 1902; Ir. Age, Dec. 11, 1902; Met. Work, Feb. 14, 1903; St. Ry. Rev., Nov. 20, 1904, p. 869; St. Ry. Jour., Oct. 15, 1904, p. 441 ; Am. Elecn., March, 1905, p. 127; Elec. Engr., Lond., Dec. 23, 1904; Engr. U.S., Feb. 15, 1905, Dec. 1, 1904; Elecn., Lond., July 21, 1899; Eng. Rec, 48-90; Trans. A.S.M.E., 15-536; Cass. Mag., June, 1906. 335. Main Steam Headers. — Until quite recently it was the usual practice to make the area of the steam header equivalent to the com- bined areas of the feeders, but the function of the header is now regarded as that of an equalizer rather than a storage reservoir. In the various large power houses recently built in New York City, with ultimate capacities of from 60,000 to 150,000 kilowatts, the largest steam headers are not over 16 inches in diameter. In some recent 630 STEAM POWER PLANT ENGINEERING designs the pipes leading from the header to the engines are two sizes smaller than called for by the engine builders. In this case large receiver separators two to four times the volume of the high-pressure cylinder are provided near the throttle as in Fig. 356. The pipes between receiver and engine are full size. The object of the arrange- ment is to give (1) a constant flow of steam, (2) a full supply of steam close to the throttle, and (3) a cushion near the engine for absorbing the shock caused by cut-off. With moderately superheated steam and boiler pressures from 125 to 150 pounds a velocity of 8000 feet per minute is allowed in the header and as high as 9000 feet per minute between header and receiver. With steam turbines velocities as high as 12,000 feet per minute are permissible, provided the pipe is less than 50 feet in length and practically free from sharp bends. Main headers are ordinarily constructed of mild steel, though cast-iron and cast- steel headers are not uncommon. Cast headers permit of fewer joints and are well adapted to situations where a number of branches are closely grouped as in Fig. 361. Cast-iron headers are employed in the Manhattan Elevated Railway power station, New York. *Tfr|i »oi in which d 4= diameter of the pipe in inches, y = density of the steam in pounds per cubic feet, and W == weight of steam flowing in pounds per minute. In determining the diameter of the steam pipe opening for recipro- cating engines a much lower velocity than 6000 feet per minute is assumed, to allow for the various conditions of operation. Average practice gives the constant in equation (149) a value of 0.3 instead of 0.175 when used in this connection. Equation (149) gives satisfactory results for pipes under 100 feet in length and between 4 and 8 inches in diameter; for larger diameters the velocity could be increased with advantage; for smaller diameters * See author's original paper, Power, June, 1907, .p. 377. PIPING AND PIPE FITTINGS 633 or greater lengths friction and condensation would cause considerable drop in pressure and some one of the approved formulas in Table 83 should be used instead. 1000 2000 3000 4000 5000 6000 SCC0 10000 12000 14000 Mean Velocity, Feet per Minute 16000 13000 Fig. 362. Drop in Pressure for Various Velocities and Pipe Sizes. Initial Pressure 100 Pounds Gauge, Length of Pipe 100 feet. A large drop in pressure means a small pipe and high velocity with consequent decrease in condensation, but a point is soon reached where the economy in the size of pipe is more than offset by the loss in friction. There seems to be no fixed rule for determining the drop 634 STEAM POWER PLANT ENGINEERING op O.co c ° I* + S* 2jS © 18 CO l-H 8 Z w « o ft (-1 P4 g£ C O OS C > rH e8. pqio 3 S3 s u o & o TJ C ja 3 K o If «£l Is <*>; ?1 o c S3 3 "£■£ ■fi-« §^a§ ^ps ^ •Q Si 33 II §1 *M o||n CO b 0) N 1:11 1 § ii ii ii PIPING AND PIPE FITTINGS 635 3 DO o ^ s i |S|» Ik Ilk Ilk 1 1 "5 S 03 5 ^►^^ ^> ^> :r ^> » « 2 § o © o o X 2 c^ (O T3 6 o II o II d II •3 d II d T3 GO ~a CC 1 CD i a 3 5r» « h CO P ^1 a £ col + % + % + % ^r =» S» ?» OS < ©■g w^ 4 -^ CO 55t— i HI- fcl* ^ ^ q ■L CO d o t-~ O CO CD r-i «o CO II fa W 3 to »c r^ CO IO CO ~ w CO CO ? — 02 CO CO CO CO CO ,-H 1 o M o o o o o o o o © o ■ 2 ° o o o o ft, A a q © q o o o q q 1 ►q Q d d d d d d d d 3 fa II II II II II II II II II fC a. 0, ft, 0s ft, ft, ft, ft, ft, H S 3. i % Tt5 ^ 00 OQ OJ col SI- fK ft, ^ H 1 GO as + t3 + + 1 fa 3 3 IIH lih 1*1 l!h IIH ft, ^ ft, ^ ft, ^ s ft7 £ ">* ■^> ^ ^>- ^> >- ^^> ^ m cn •* *°. «o »o IO fa £ 6 CO •* CO »o co' r- l> t^ ■* S U5 »o »c "O 00 00 00 oi s II II II II 11 II II II II fa fe fe fe fe fe ^ fe ^ ^ k o o 3 &T* O ] ^°T^ ■o GQ § eoP »P eoP F 2 2 1 5 + S + 13 ft, + 1 o 1 fa J o (N SO ISIS o 00 o CO co_ c a > 5 i o o > c Ifl c > 3 ► ' ^ (M t^ a> d d »o d d »* i> OJ © II .7 II II i-t II •* ^ ^ ^ ^ >. ^ p^ ^ U {^ ■3 si I* c >5 fe o s: 3 3 £ 00 .3 d to 8 W ffl d w a 3 < Is 3 a 1 I 3 ■8 I •II 1 •j aa( wt* ) 636 STEAM POWER PLANT ENGINEERING most suitable for any given set of conditions. In current practice the drop in pressure between boiler and engine ranges from a fraction of one pound to four pounds per square inch per 100 feet of pipe, with an average between one and two pounds. Table 83 gives a few of the best known formulas for the flow of steam, and Table 84 a comparison between them with respect to velocity, weight discharged, diameter, and the drop in pressure. Formula 11, Table 83, is the most commonly accepted, and the curves in Fig. 362 are based upon it, assuming a steam pressure of 100 pounds absolute and pipe lengths of 100 feet. Within the limit of 12,000 feet per minute velocity and 10 pounds per square inch drop in pressure the curves are sufficiently accurate for all practical purposes, but beyond this range the results are purely conjectural and may not be accurate, as no recorded experiments have been conducted at these high velocities or with pipes of large diameters. Though applicable directly to pipes 100 feet long with mean pres- sure of 100 pounds per square inch absolute, they may be used for any length or pressure. For example, for any length other than 100 feet, multiply the drop given in the curves by the required length in feet and divide by 100. For any pressure other than 100 pounds abso- lute, multiply the drop given in the curves by 0.2271 (density of steam in pounds) and divide by the density of steam at the required pressure. Table 85 is the table ordinarily used in connection with the flow of steam and is calculated from equation 11. Table 86 is based upon equations 4 to 12. The results differ slightly from those in Table 85, though the latter is more comprehensive. The left-hand half of Table 86 gives the discharge in pounds per minute for pipes of various diameters corresponding to drop of pressure as given on the right-hand side in the same horizontal line; e.g., a 6-inch pipe 100 feet long dis- charges 371 pounds of steam per minute for a drop of 16.4 pounds at 100 pounds pressure. 337. Equation of Pipes. — It is frequently desirable to know what number of one sized pipes will be equal in capacity to another pipe. According to the formulas in Group II, Table 84, the weights dis- charged vary with the square root of the fifth power of the diameter; that is, the number of pipes equal in capacity to any given pipe may be determined from the equation N 1 = d^-^dX (150) in which N x = number of pipes of diameter d l equal in capacity to a pipe of diameter d; d, and d in inches. PIPING AND PIPE FITTINGS 637 TABLE 85. FLOW OF STEAM THROUGH PIPES (BABCOCK). Diameter of Pipe, in Inches. Length of each = 240 diameters. sure by Gauge. i 1 l* 2 2* 3 4 Pounds per Square Inch. Weight of Steam per Minute, in pounds, with One Pound Loss of Pressure. 1 1.16 2.07 5.7 10.27 15.45 25.38 46.85 10 1.44 2.57 7.1 12.72 19.15 31.45 58.05 20 1.70 3.02 8.3 14.94 22.49 36.94 68.20 30 1.91 3.40 9.4 16.84 25.35 41.63 76.84 40 2.10 3.74 10.3 18.51 27.87 45.77 84.49 50 2.27 4.04 11.2 20.01 30.13 49.48 91.34 60 2.43 4.32 11.9 21.38 32.19 52.87 97.60 70 2.57 4.58 12.6 . 22.65 34.10 56.00 103.37 80 2.71 4.82 13.3 23.82 35.87 58.91 108.74 90 2.83 5.04 13.9 24.92 37.52 61.62 113.74 100 2.95 5.25 14.5 25.96 39.07 64.18 118.47 120 3.16 5.63 15.5 27.85 41.93 68.87 127.12 150 3.45 6.14 17.0 30.37 45.72 75.09 138.61 Initial sure by Gauge Pounds per Square Inch. 1 10 20 30 40 50 60 70 80 90 100 120 150 Diameter of Pipe, in Inches. Length of each =240 diameters. 10 12 15 18 Weight of Steam per Minute, in Pounds, with One Pound Loss of Pressure. 77.3 95.8 112.6 126.9 139.5 150.8 161.1 170.7 179.5 187.8 195.6 209.9 228.8 115.9 143.6 168.7 190.1 209.0 226.0 241.5 255.8 269.0 281.4 293.1 314.5 343.0 211.4 262.0 307.8 346.8 381.3 412.2 440.5 466.5 490.7 513.3 534.6 573.7 625.5 341.1 422.7 496.5 559.5 615.3 665.0 710.6 752.7 791.7 828.1 862.6 925.6 1009.2 502.4 622.5 731.3 824.1 906.0 979.5 1046.7 1108.5 1166.1 1219.8 1270.1 1363.3 1486.5 804 996 1170 1318 1450 1567 1675 1774 1866 1951 2032 2181 2378 1177 1458 1713 1930 2122 2294 2451 2596 2731 2856 2975 3193 3481 For any other length divide 240 by the given length expressed in diameters and multiply the tabular quantity by the square root of this quotient, which will give the flow for one pound loss of pressure. Conversely, dividing the given length by 240 will give the loss of pressure for the flow given in the table. 638 STEAM POWER PLANT ENGINEERING 3 2 II C w &l ■S s^ 2 J3 o tS < * °* S £ g. £ ~ on O « H SI 00 o »0 ONifl-HOOlOfqO)N^MO»MO^ CO»OiO'>^COCO1"MM«50(0'0^0 HOMOiOiO^^MMNNNHHrtOO ; oo ■*MOOOOOOOMO-itOco^aN(»(N W^t^rHNrHrjfQOMQOMM^OtOCOOM «OMOMlOO>ffiCOONffi©CO(N ^HOomioNocoOMrH©CON©>0>O^COM(NNrH!HFH Q '.'.'.'.'.'.'.'.'.'.'.'.'.'.'.'. '. oooooooooooooooooo (N"*COOOOH«CO^>0(DN(»ffiOiHNrt lOMHffiOOt*©iOTl*COCqiHOffi©OON© CMCONO>ONCOiOOOOOHMi(5(0(» rtMiON04«iOO>M©OMNHTfMfHiO 00(0>0'*NIN'HOOai0500NN«iOiO^I OOfO^NOaXO^NOMtO^NOaXO^ OONOlO'^IN'-lOffiOOOlO^MNOOJOO WM(N(NMMININhhhhhi-ihh 3)Ol 050fflO!05CS050iOi3iOffiOO O O i0HN«0>N>0C0H00N>0MHOM(0^ tQiQT)lT|tCOWCOMWMW(N(NINC0»0'^'*fOMCO __ iHTjt00N>O(NO5ON©©(N0it0«O)t0lN COCOMINNNINMNhhhhhhhh a> to t-i e8 OOOOOOOOOOOOOOOOOO s 00 ©NlOOO'HMTttiflNOOO'HCO'*-COcOiO»O»O"COtO00O5'HCO'H«O00aiHCOTtltt5Q0 WrHO>MOiO^'*M(N'-lOaifflOON«OiO PIPING AND PIPE FITTINGS 639 According to the formulas in Group I, Table 84, the weights dis- charged vary as \ d 5 -s- 1 1 + —j- ) \ and the equation becomes *• " TTIe * "Tie (151 > I d + dj = /J(d,+3.6)\* U' W + 3.6)/ (15i) - J^*™- d53) From (150) and (153) we see that the values of N t are practically the same for either equation when the ratio of d to d ± is small and that they differ widely for large ratios. For example, according to (150), 5.7 eight-inch pipes are equivalent in capacity to one sixteen-inch pipe, whereas (152) gives 6.15. The difference is negligible. Again, according to (150), 180 two-inch pipes are equivalent in capacity to one sixteen-inch pipe, whereas (153) gives 274. The difference is con- siderable. Equation (153) is most commonly accepted and is the basis of Table 87. 338. Friction through Valves and Fittings. — The formulas out- lined in Table 83 are strictly applicable only to well-lagged pipes, free from bends or obstructions of any kind such as valves or fittings, which greatly increase the resistance of the flow of steam. If these obstructions must be considered, it is customary to allow for them by assuming an added length of straight pipe equivalent in resistance to the various fittings and bends. Unfortunately, the few tests which have been made for the purpose of determining the resistance of vari- ous pipe fittings give discordant results, and in the absence of more recent data the rules given by Briggs (" Warming Buildings by Steam") are probably as accurate as any. According to Briggs, the length of pipe in inches equivalent to the resistance of one standard 90-degree elbow is L = 76 d -*- (l + y) < 154 > and to that of one globe valve L = 114 d + (l + y)' < 155 > The resistance of gate valves is not considered. 640 STEAM POWER PLANT ENGINEERING g iCN Tft 00 OS CO i-H OS • I>iO"#CNCN' I i-t i-l CM CO >0 OS CO OS >t>00i ICOrHi i-t 00 IO OS CO ■* CO r-H© co i-H Tft os r^ r^ co o rH>OCOCOOS©CSCOI>t>CSCOCO co-- ■ ' CO CO t- lO CO rH CO O b- OS Tjt >iOiO< )1-t 00 r^ CO CN CO •"* CN i CO O Tjt OS Tft ( tCNCNCOCO'*COOSt^t^©CO OSiOCOlO • CN O CO O 00 - " rHCOrH CN CM CO CO ■* ) rH 00 Ttt OS CO CD 00 OS CO • CO 1^ i-H CO .-t > 00 "* CN rH i-l CN i-h O 00 I- .-h Tjt O IO CO 00 CN CN CD CO IO ) CO t- I> 00 ■* CO c OSrH -t^OI o^*< -cooo< CNi-t -Hi OCOCO CSCNiO • i-l 00 00 CO IO ■* IO 00 00OSCO CNr-COCNCOt^COOCO ->Or-t05CSOCOt^TttiOTjtI>CNOSCNCOOiOCN CDCN.-H i-H i-l CN OS OS i-t CO CO tH •T-tCNCNCOlOCOOOSi-HCO'OOOCOOOt^ COi-li-H • tH i-H i-H rH CN CO CO CMOS'* • CO OS rfi iO O CN 0*OCSCO x-HOOOO • CO "# iO 00 "# CM CO CD tH t> 00 CO CO OS i-H rH tr^ IO CO i-HrHCNCO ll>COi-l • i-H (N CO "* CO 00 O CN IO 00 rH IO OS 00 CN I • rH rH i-t i-H CN CN CN CO CO O-tf COCOCOCN CO rH OS CO CO IOSrH©rHl>CD • • •i-HCNCO-tfCO tCNTjtcOOOrHiOOOCOOCTttOSCO'HHO 1-Hi-HrHCNCNCOCOTttlOr^ COCDCO -rHCOCO CO OS l> O CO >0 -# io OSiOOSi-lOOO •CNO»OOC0iO00i-HC01>CNC0OOrHiO»OHHO'*CD i— I rH i— I CN ^ t^ © ^ COCNOCOCOi-l •CNicHCOO'^OSiOCOrHOCNlOOO ~~ CS^CN • 1-H rH rH CN CO ■* IO CO t^ 00 -tfr*. -C0t>0 COiOCOrHOOOCDt ICOCO ■ CO CO C^ OS CO 00 >0 OS rH CN IO O CN Tt< l> OS CD rH ( rH i-H rH i-H rH CN ^ I ICOi-t •rHCOCOOCOCOCNCN>OOS^ I • rH i-t CN CO ■* IO CD 00 ■COCO rHTjid COrH )rH 00 ICOCO i-Ht^00( rHb- CO OS 00 CO "tf T« O i-t t- rHOOkOiHHCNiOOO'cttOCNiOOSCOt^CNt^OO i-H i-H rH rH CN CN CO CO 'CH JiOrft srttio tCNiO OOCSTjt -CN' IO CO CO • CN tJH I> iO C irHCNCNCO'^iOCOt^OO, Ol-H irH<© ■rHOS "«*■** COCO i>o" 00c0CNOCNt>00CDC000TttCNi I Tft CM (N -<*t 00 Tjt CO """ • t-i CN CO ■* IO t> OS » _ -T i-" rH i-i" CN CN CO* io" os" iO co" co" •OO -&CO< • OS CO O Tft Tfl tH CO c rH CN Tft CO OS CN CO rH r^ CO C OS r- t^ 00 CN CO OS rH CN 00 ■ cm os rH os io r-Ti-Tcvf cn" co"t)h"tjh" io"co"i> o"co"oo"co"oo">o* • • •rHrHCOCOi-ll>rJtcO M CO O CO. CO OOOSlOCOCOCOTttrHt^CO. • CN 1^ j* ^H ^ ,_,* rf fOTfTu^iCoO* o'CN r)*t>*©*c£* CN* iOO*t-' OS* CN '° rH r- rH rH CN CN "# b- CM h- -* i— I rH CN I CN CO CO ■«*« •*< I 2 PIPING AND PIPE FITTINGS 641 642 STEAM POWER PLANT ENGINEERING 339. Exhaust Piping, Condensing Plants. — The exhaust piping in condensing plants is arranged either according to (1) the independent or (2) the central condensing system. In the former each engine is provided with an independent condenser and air pump. In case the vacuum " drops " or it is desired to operate non-condensing, the steam is discharged through a branch pipe with relief valve to the atmos- phere, Figs. 3 and 219. When there are a number of engines in one installation the atmospheric pipes lead to a common free exhaust main, which, on account of its great size, is ordinarily constructed of light- weight riveted steel pipe. The short connection between engine and condenser is usually made with lap-welded steel pipe, since riveted joints are apt to leak, due to the engine vibrations. In a central con- densing plant, Fig. 226, the several engines exhaust through a com- mon main into a single large condenser. An atmospheric relief valve is usually provided in connection with the condenser, and no free exhaust main is necessary. Several arrangements of condenser piping are illustrated in Figs. 219 to 228. 340. Exhaust Piping, Non-Condensing Plant. — Webster Vacuum System. In the majority of non-condensing plants the exhaust steam is used for heating purposes. One of the best-known systems of exhaust steam heating, in which the back pressure on the engine is reduced by circulating below atmospheric pressure, is that known as the Webster combination system. The general arrangement is illus- trated in Fig. 2 and the principles of operation are described in para- graph 3. It has the advantage of affording (1) minimum back pressure on the engine; (2) effective and continuous drainage of condensation from supply pipes and radiators; (3) continuous removal of air and entrained moisture from confined spaces; (4) inde- pendent regulation of temperature in each radiator; (5) continu- ous return of condensation to the boiler; (6) utilization of part of the exhaust steam for preheating the feed water; and (7) automatic regulation. Fig. 363 gives a diagrammatic arrangement of the piping and appurtenances in a typical installation. The characteristic feature of this system is the automatic outlet valve attached to each part requiring drainage, which permits both the water of condensation and the non-condensable gases to be removed continuously. The radiator temperature may be regulated by Varying the quantity of steam sup- plied either by hand, or automatically by thermostatic control. The Webster valve, Fig. 364, enables the vacuum to withdraw the water of condensation as fast as it is formed irrespective of the pressure in the radiator, hence the supply may be throttled to such an extent that the temperature in the radiator is practically as low as that of steam cor- PIPING AND PIPE FITTINGS 643 responding to the pressure in the vacuum line. The small annular space between the inner tube of the float F and the guide H permits of a vacuum in the body of the valve. When the water from the radiator lifts the float the water is drawn into the returns pipe. The S^Wi^!^ OUTLET Fig. 364. Webster Air Valve. Fig. 365. Automatic Vacuum Valve, Illinois Engineering Co; valve then returns to its seat and the escape of steam is prevented except such as finds its way through the annular space around the guide stem H. An improvement on this valve which prevents the escape of steam is illustrated in Fig. 365. When steam is admitted to the radiator the condensation flows into the valve, righting the float A and sealing the outlet B against the passage of steam; as the valve fills with water the buoyancy of the float raises it from its seat and permits the water to be drawn out; the float falls and reseats on the nipple when about a half-inch of water remains in the valve, thus maintaining a water seal. Screen D prevents scale and dirt from entering the valve proper. By-pass H is for emergency use in draining off accumulated water in the radiator in case the valve becomes stopped up ; and permits the bonnet to be removed without trouble from the accumulated water. 341. Exhaust Piping, Non-Condensing Plants. — Paul Heating System. The Paul vacuum system differs from the Webster in that the con- densation, and the air and non-condensable gases are separately handled. Referring to Fig. 366, which gives a diagrammatic arrange- ment of the piping, the condensed steam gravitates to the automatic return tank and pump and is pumped either directly to the boiler or through the heater to the boiler. Air and vapor are withdrawn from the upper part of the radiator by the Paul exhauster or ejector, E, and discharged into the returns tank, which is vented to the atmos- phere for the escape of the non-condensable gases. The exhauster 644 STEAM POWER PLANT ENGINEERING PIPING AND PIPE FITTINGS 645 receives its supply of steam through pipe 0, Fig. 367, which shows the general arrangement of this apparatus. The piping is in duplicate to guard against failure to operate. The suction side of the exhauster is connected with the air pipes sur.TinN SUCTION A, A, Fig. 366. Fig. 368 gives a section through the Paul air or vacuum valve which prevents steam from blowing into the air pipes and permits only air to pass. In Fig. 366 the heat- ing system is piped on what is known as the " one-pipe down-feed system;" i.e., the exhaust steam is first conducted to a distributing header in the attic, from which the various supply pipes are led to the radia- tors. The water of conden- sation returns through these same pipes and gravitates to the returns pump. Both the supply steam and the condensation flow in the same direction. This system is also piped on the " one-pipe up-feed," the " two-pipe up-feed," and the " two-pipe down-feed " principle. The " one-pipe up-feed " differs from the sys- tem just described in that the steam flows upward through the risers and does away with the attic piping. The returns, however, flow against the current of steam and water hammer is more likely to occur than with the down-feed system. In the two-pipe systems the steam supply pipes or risers conduct steam only, and the returns carry the condensation. The one-pipe down-feed is cheaper and simpler and practically as efficient as the two-pipe system under normal conditions. It is objection- able, however, due to the difficulty of draining the radiator with STOP VALVE o STEAM SUPPLY Fig. 367. Paul Exhauster. COMPOSITION FROM RADIATOR TO EXHAUSTER Fig. 368. Paul Vacuum Valve. 646 STEAM POWER PLANT ENGINEERING SSWSSSWASS closely throttled supply valve, since the velocity of the entering steam prevents the water from returning through the same orifice. 342. Automatic Temperature Control. — Experience shows that a considerable saving in fuel may be effected in the heating plants of tall office buildings and similar plants by automatically controlling the temperature. Hano)-controlled valves are usually left wide open, and when the room becomes too hot the temperature is frequently lowered by opening the window, resulting in a waste of heat which may be considerable in modern buildings with hundreds of offices. Many successful methods of automatic temperature control are available, the usual system consisting of thermostats which control the supply of heat by means of diaphragm valves, the latter taking the place of the usual radiator supply valve. Fig. 369 shows a Powers thermostat. The expansible disk U con- tains a volatile liquid having a boiling point of about 50 degrees F. The pressure of the vapor within the disk at a temperature of 70 degrees amounts to six pounds to the square inch, and varies with every change of temperature, causing a variation in the thickness of the disk. The disk is attached by a single screw to the lever Q, which rests upon the screw F as a fulcrum. The flat spring R holds the lever and disk against the mova- ble flange M . Connecting with the cham- ber N are two air passages H and /. The thermostat is attached by means of two screws at the upper end to a wall plate per- manently secured to the wall. This wall plate has ports registering with H and I, one for supplying air under pressure and the other for conducting it to the diaphragm motor which operates the valve or damper. Air is admitted through H under a pressure of about fifteen pounds per square inch, and its passage into chamber N is regulated by the valve J, which is normally held to its seat by a coil spring under cap P. K is an elastic diaphragm carrying the flange M, with escape valve passage covered by the point of valve L. Valve L tends to remain open by reason of the spring. When the temperature rises sufficiently expansion of the disk U first causes the valve to seat, its Fig. 369 Section Through Powers Thermostat. PIPING AND PIPE FITTINGS 647 spring being weaker than that above valve J. If the expansive motion is continued, valve J is lifted from its seat and compressed air flows into chamber N, exerting a pressure upon the elastic diaphragm K in opposition to the expansive force of the disk. If the tempera- ture falls, the disk contracts and the overbalancing air pressure in N results in a reverse movement of the flange M, permitting the escape valve to open and discharge a portion of the air; thus the air pressure is maintained always in direct propor- tion to the expansive power (and tem- perature) of the disk U. The passage / communicates with a diaphragm valve, Fig. 370. The compressed air operates the diaphragm against a coiled spring resistance, so that the movement is proportional to the air pressure and the supply of steam controlled accord- ingly. The adjusting screw G, squared to receive a key, carries an indicator by means of which the thermostat can be set to carry any desired temperature within its range, usually from 60 to 80 degrees. In changing the temperature adjustment lever Q forces the disk U closer to or farther from the flange M . In connecting up the system com- pressed air is carried to the thermostat and diaphragm valves, from a reservoir through small concealed pipes. In the indirect system of heating the dampers are of the diaphragm type and the method of regulation is the same as with the direct system. 343. Feed- Water Piping. — The simplest arrangement of feed- water piping may be found in non-condensing plants, in which the feed water is obtained under a slight head, such as is afforded by the average city supply, and is heated in an open heater by the exhaust steam from the engine to a temperature varying from 180 to 210 degrees F. The hot feed water gravitates from the heater to the pump and then is forced to the boiler, or to the economizer if one is used. If a meter is used it is generally placed on the discharge side of the pump, and should be by-passed to permit it to be cut out for repairs. (Fig. 371.) Plants operating continuously should have feed pumps in duplicate. In some cases the returns from the heating system gravitate to the heater and only enough cold water is added to make up the loss from leakage, etc. In other cases the returns gravi- Fig. 370 Diaphragm 648 STEAM POWER PLANT ENGINEERING tate to a special " returns tank/' from which they are pumped directly to the boiler without further heating. Occasionally a live-steam purifier is used, especially if the water contains a large percentage of Q£> QG ft ILLH » BUILtM ±± — ■ FEEO T MAIN — ■ INJECTOR MAIN — I-J [ FEED PUMP | | FECI I J I | INJECTOR—l X^I 1 4—1 WATER 1*1 ' COLD WATER SUPPLY HEATER / I "•».»!»* X " tass*-- Fig. 371. Feed Water Piping; Now Condensing Plant. calcium sulphate. The feed is then subjected to boiler pressure and temperature and the greater part of the impurity precipitated before it enters the boiler. Closed heaters are often used in place of open heaters. When the supply is not under head a closed heater is usually preferred and is placed between the pump discharge and the feed main. In condensing plants the feed piping is similar to that in non- condensing plants, except that if exhaust steam is used for heating ©0 o© ex n a Fig. 372. Feed Water Piping; Condensing Plant. purposes it is supplied by the auxiliaries, such as feed pumps, stoker engines, condenser engines, and other steam-using appliances. In plants having a number of boilers it is customary to run a feed main or header the full length of the boiler room and connect it to PIPING AND PIPE FITTINGS 649 each boiler by a branch pipe. This main may be a simple header or in duplicate or of the " loop " or " ring " type. Horizontal tubular boilers are frequently arranged in one battery with the feed main run along the fronts of the boilers just above the fire doors. Water-tube boilers are generally set in a battery, and as the arrangement above would block the passageway between the batteries, the main is run either above or under the settings, the former being the more common. Where a single header is used, the feed pumps are sometimes placed so as to feed into opposite ends of the main, which is then cut into sections by valves. Another arrangement is to place the pumps so as to feed into the middle of the header. With the loop arrangement the main is ordinarily cut into sections by valves so that the water may be sent either way from the pumps and any defective section cut out. With duplicate mains a common arrangement is to place one main along the front of the boiler and the other at the rear or both over- head as in Fig. 359. Sometimes one main is placed in the passageway below the boiler setting and the other on top. Standard wrought-iron pipe is usually used for pressures under 100 pounds and extra heavy pipe for greater pressures. The pipes and fittings from boiler to main are frequently of brass, and preferably* so, since brass withstands corrosive action much better than iron or steel. Flanged joints should be used in all cases, since the pockets formed by the ordinary screwed joints hasten corrosion at those points. {Power, June, 1902, p. 4.) Fig. 373, A to E, illustrates the various combinations of check valve, stop valves, and regulating valve in steam boiler practice. The Fig. 373. Different Arrangements of Valves in Feed Water Branch Pipes. simplest arrangement and one sometimes used in plants operating intermittently is shown in A. Here there are but two valves between the boiler and the main, the check being nearest the boiler and the stop valve at the main. The stop valve performs both the function of cutting out the boiler and of regulating the water supply. This 650 STEAM POWER PLANT ENGINEERING arrangement is not recommended, as any sticking or excessive leaking of the check valve will necessitate shutting down the boiler. B shows the most common arrangement. Here the check valve is placed between the regulating valve and a stop valve as indicated. This permits a disabled check to be easily removed while pressure is on the boiler and the main. E shows an arrangement whereby both check and regulating valve may be removed, and is particularly adapted to boilers operating continuously where the regulating valve is subjected to severe usage. In this case the stop valves are run wide open and are subjected to no wear. The regulating valve most highly recom- mended is a self-packing brass globe valve with regrinding disk. The check valve is ordinarily of the swing check pattern with regrinding disk, Fig. 384 (C). Modern practice recommends an automatic water relief valve in the discharge pipe immediately adj acent to each pump to prevent excessive pressure in case a valve is accidentally closed in by-passing or in changing over. 344. Flow of Water through Orifices, Nozzles, and Pipes. — Ber- noulli's theorem is the rational basis of most empirical formulas for the steady flow of a fluid from an up-stream position n to a down- stream position m, thus ("Mechanics of Engineering/ ' Church, p. 706): + TV + Z m = Pn y + + z t fall losses of head 1 -[ occurring between >•> [n and m r — r "m — T - y 2g y 2g in which V = velocity in feet per second at the point considered. P = pressure in pounds per square foot. Z = potential head in feet of the fluid. y = density of the fluid, pounds per cubic foot. g = acceleration of gravity. V 2 (156) Each loss of head will be of the form coefficient of resistance to be determined experimentally head due to skin friction is expressed : 7 v 2 in which / = the coefficient of friction of the fluid in the pipe. I = length of the pipe in feet. d = diameter of the pipe in feet. Other notations as in (156). K— , in which K is the 2g 9 The loss of (157) PIPING AND PIPE FITTINGS 651 Discharge from a circular vertical orifice with sharp corners: Q = CA V2gh } (158) in which Q = cubic feet per second. C = coefficient, varying from 0.59 to 0.65 (Merriman, Treatise on Hydraulics, p. 118). A = area of the orifice, square feet. h = head of water in feet. g = acceleration of gravity =32.2. Discharge from short cylindrical nozzles three diameters in length, with rounded entrance (" Mechanics of Engineering," Church, p. 690) ; Q = 0.815 A V2~gh. (159) Discharge from short nozzles with well-rounded corners and conical convergent tubes, angle of convergence 13J degrees (Church, p. 693): Q = 0M-A\/2gh. (160) Discharge from cylindrical pipe under 500 diameters in length (Church, p. 712): Q = 6 - 3 v/ ( i + £ + ? fl ' (161) in which / = coefficient of friction. Other notations as above. / varies with the nature of the inside surface, the diameter of the pipe, and the velocity of flow. Discharge through very long cylindrical pipes (" Mechanics of Engi- neering," Church, p. 715): Q = 3.15 y/^ • (162) Loss of head due to friction in water pipes. Weisbach's formula is as follows: ou 4+ 0J)lM^ (163) \/V J 5.367 d H=(0 in which H = friction head in feet. V = velocity in feet per second. L = length of pipe in feet. d = diameter of pipe in inches. 652 STEAM POWER PLANT ENGINEERING TABLE OF THE COEFFICIENT / FOR FRICTION OF WATER IN CLEAN IRON PIPES. (Abridged from Fanning.) Velocity in Ft. per Sec. Diam. = | in. Diam. = 1 in. Diam. = 2 in. Diam. = 3 in. Diam. = 4 in. Diam. = 6 in. Diam. = 8 in. = .0417 ft. = .0834 ft. = .1667 ft. = .25 ft. = .333 ft. = .50 ft. = .667 ft. 0.1 .0150 .0119 .00870 .00800 .00763 .00730 . 00704 0.3 .0137 .0113 850 784 750 720 693 0.6 .0124 .0104 822 767 732 702 677 1.0 .0110 .00950 790 743 712 684 659 1.5 .00959 .00868 .00757 .00720 .00693 .00662 . 00640 2.0 .00862 810 731 700 678 648 624 2.5 795 768 710 683 662 634 611 3.0 .00753 .00734 .00692 .00670 .00650 . 00623 .00600 4.0 722 702 671 651 631 607 586 6.0 689 670 640 622 605 582 562 8.0 663 646 618 600 587 562 544 12.0 630 614 590 582 560 540 522 16.0 .00618 .00600 .00581 .00570 .00552 .00530 .00513 20.0 615 598 579 566 549 525 508 Velocity in Ft. per Sec. Diam. Diam. Diam. Diam. Diam. Diam. Diam. = 10 in. = 12 in. = 16 in. = 20 in. = 30 in. = 40 in. = 60 in. = .833 ft. = 1.00 ft. = 1.333 ft. = 1.667 ft. = 2.50 ft. = 3.333 ft. = 5. ft. 0.1 .00684 673 659 . 00669 657 642 .00623 614 603 0.3 .00578 567 0.6 .00504 .00434 .00357 1.0 643 624 588 555 492 428 353 1.5 .00625 .00607 .00572 .00542 .00482 .00421 .00349 2.0 609 593 559 529 470 416 346 2.5 596 581 548 518 460 410 342 3.0 .00584 .00570 .00538 .00509 .00452 .00407 . 00339 4.0 568 553 524 498 441 400 333 6.0 548 534 507 482 430 391 324 8.0 532 520 491 470 422 384 320 12.0 512 500 478 457 412 377 .00313 16 00502 00491 00470 00450 00406 00370 20 498 485 William Cox {American Machinist, Dec. 28, 1893) gives a simple formula which gives almost identical results: H = (4 V 2 + 5 7-2) L 1200 d Notations as in (163). (164) Loss of head due to friction of fittings. Formulas (161) to (164) are based on the flow of water through clean straight cylindrical pipes. PIPING AND PIPE FITTINGS 653 Where there are bends, valves, or fittings in the line the flow is decreased on account of the additional resistance. These frictional losses are conveniently expressed in feet of water, thus: H = C^P- (165) C having the following values: Angles. Class of Valve. 45 degrees. 90 degrees. Gate. Globe. Angle, C 0.182 0.98 0.182 1.91 2.94 Example: Determine the pressure necessary to deliver 200 gallons of water per minute through a 4-inch iron pipe line 400 feet long, fitted with four right-angle elbows and two globe valves. The water is to be discharged into an open tank. A flow of 200 gallons per minute gives a velocity of ?= 5 feet per second (7.48 = number of gallons per 7.48 X 60 X 12.72 cubic foot, and 12.72 = internal area of the pipe, square inches) From the preceding table, / = 0.618 for 7 = 5. From (165), Resistance head of 4 elbows = 0.98 X -=^ X 4 = 1.52 feet. 64.4 Resistance head of 2 globe valves: 1.91 X -^- X 2 = 1.48 feet. 64.4 Resistance head of all fittings: 1.52+ 1.48 = 3 feet. Substitute 7 = 5, L= 400, and d = 4 in (164). (4 X 5 2 + 5 X 5 - 2" "-( 400 1200 X 4 10.25 feet, resistance head of the pipe. Total resistance head = 10.25 + 3 = 13.25 feet of water, or 5.75 pounds per square inch. Example: How many gallons of water will be discharged per minute through above line with initial pressure of 100 pounds per square inch, and what will be the pressure at the discharge end? Since / depends upon the unknown 7, we may put / = 0.006 for a first approximation and solve for 7; then take a new value of / and substitute again, and so on. 654 STEAM POWER PLANT ENGINEERING Substitute /= 0.006, d= -^- ,h = 100X2.3 = 230, and Z = 400 in (162) Q = 3.15 Vo. 33 5 + 230 .006 X 400 = 1.95 cubic feet per second, corresponding to a velocity of 22 feet per second. From the preceding table, / = 0.00548 (by interpolation) for V = 22 feet per second. From (165) the friction of 4 elbows and 2 globe valves is found to be 58 feet for V = 22. From (164) a resistance head of 58 feet of water for V = 22 is found to be equivalent to 136 feet of straight pipe, thus: 4 X 22 2 X 5 X 22-2 58 -■ 1200 X 4 .), L = 136. Substitute /= 0.0548, I = 400 + 136 = 536 in (162): Q=-3.15v/°- 335X230 V 0.0058 X 536 = 1.74 cubic feet per second, corresponding to a velocity of 19.3 feet per second. = 780 gallons per minute. If greater accuracy is necessary determine / and L for V = 19.3 and proceed as above. The total friction head may be determined from (164) thus: 4 X 19.3 2 + 5 X 19.3-2 H - 536 1200 X 4 = 177 feet of water. = 77 pounds per square inch. The pressure at the discharge end will be 100 — 77 = 23 pounds per square inch. Average power plant practice gives the following maximum veloci- ties of flow in water pipes : Size of Pipe in Inches. Velocity, Feet per Minute. Size of Pipe in Inches. Velocity, Feet per Minute. Hoi J to lj 1J to 3 50 100 200 3 to 6 Over 6 250 300-400 PIPING AND PIPE FITTINGS 655 345. Stop Valves. — The valves used to control and regulate the flow of fluids are the most important element in any piping system. A good valve should have sufficient weight of metal to prevent distortion under varying temperature and pressure, or under strains due to connection with the piping; the seats should be easily repaired or renewed; there should be no pockets or projections for the accu- mulation of dirt and scale, and the valve stem should permit of easy and efficient packing. Stop valves are made in such a variety of designs that a brief description will be given of only a few of the best- known types. Fig. 374 shows a section of an ordinary globe valve, so called because of the globular form of the casing. This type of valve is the most Fig. 374. A Typical Globe Valve, Screw-Top, inside Screw. Fig. 375. A Typical Globe Valve, Bolt-Top, outside Screw. common in use. Globe valves are designated as (1) inside screw and (2) outside screw, according as the screw portion of the stem is inside the casting, Fig. 374, or outside, Fig. 375. The top or bonnet may be screwed into the body of the valve, Fig. 374, or bolted, Fig. 375. The smaller sizes, three inches and under, are usually of the screw-top type and the larger of the bolt-top type. Valves with outside yoke and screw are preferable to the other in that they show at a glance whether the valve is open or closed, an advantage in changing from one section to another. The disks are made in a variety of forms, the material 656 STEAM POWER PLANT ENGINEERING depending upon the nature of the fluid to be controlled. Thus, for cold water, hard rubber composition gives good results; for hot water and low-pressure steam, Babbitt metal; for high-pressure steam, copper or bronze; and for highly superheated steam, nickel. The valve bodies are of brass for sizes under three inches, cast iron for the larger sizes and ordinary pressures and temperatures, and cast steel or semi-steel for high temperatures and pressures. Globe valves should always be set to close against the pressure, otherwise they could not be opened if the valves should become detached from the stem. Globe valves should never be placed in a horizontal steam return pipe with Fig. 376. A Typical Gate Valve, Solid-Wedge, Screw- Top, outside Screw. Fig. 377. A Typical Gate Valve, Solid- Wedge, Bolt-Top, inside Screw. the stem vertical, because the condensation will fill the pipe about half full before it can flow through the valve. Globe valves that are open all the time are preferably designed with a self-packing spindle, as in Fig. 375, in which the top of shoulder C can be drawn tightly against the under surface of bonnet S, thus preventing steam from leaking past the screw threads while the spindle is being packed. Figs. 376 to 379 show different types of gate or straight way valves. These valves offer little resistance to the flow of steam or liquid passing through them, and are generally used in the best class of work. Fig. 376 shows a section through a solid-wedge gate valve with outside PIPING AND PIPE FITTINGS 657 screw and yoke. This form of outside screw and yoke with stem protruding beyond the hand wheel is a perfect indicator to show whether the valve is open or shut, as the hand wheel is stationary and the spindle rises in direct proportion to the amount the valve is opened. For these reasons outside screw valves are preferable for high-pressure work and especially for the larger sizes. The seats are made solid, or removable, and of various materials for different pres- sures and temperatures. Fig. 378 shows a section through a split- wedge gate valve with parallel faces and seats. For the sake of illus= Fig. 378. A Typical Gate Valve, Split- Wedge, Bolt-Top, inside Screw. Fig. 379. Ludlow Angle Valve, Gate Pattern. tration this valve is fitted with inside screw. In this design the spindle remains stationary so far as any vertical movement is con- cerned, and the gate or plug being attached to it by means of a threaded nut rises into the bonnet when the spindle is revolved. It is impossible to tell by its appearance whether this form of valve is open or closed. Valves with inside screw are adapted to situa- tions where there is considerable dirt and grit, since the screw is inclosed and protected and excessive wear is thus avoided. Gate valves with split gates are more flexible than those with solid gates, and hence are less likely to leak. Fig. 379 shows the application 658 STEAM POWER PLANT ENGINEERING of the gate system to an angle valve. All high-pressure valves above 8 inches in diameter should be provided with a small by-pass valve, as the pressure exerted against the disk or gate is very great when the valve is closed and the force required to move it is considerable. The by-pass valve also facilitates " warming up " the section to be cut in and is more readily operated than the main valve. 346. Automatic Non-Return Valves. — Fig. 380 shows a section through an automatic non-return valve as applied to the nozzle of a steam boiler. As will be seen from the illustration it amounts to practically a large check valve with cushioned disk. The object of this device is the equali- zation of pressure between the different units of the battery, the valve remaining closed as long as the individual boiler pressure is lower than that of the header. In case a tube blows out the valve closes automatically, owing to the reduction of pressure and prevents the header steam from entering the boiler. It acts also as a safety stop to prevent steam being turned into a cold boiler while men are working inside, because it cannot be opened when there is pressure on the header side only. To be successful, such a valve should not open until the pressure in the boiler is equal to that in the header; it should not stick and become inoperative nor chatter and hammer while performing its work. Referring to Fig. 380, tail rod E insures alignment and hence prevents sticking; steam space C acts as a dashpot to prevent hammering of the valve as it rises, and steam space D acts as a cushion and prevents hammering at closing. Lip F is made to enter the opening in the seat and reduce wire draw- ing across the seat. Fig. 358 shows the installation of a number of non-return valves at the Yonkers power house of the. New York Central Railway Company. 347. Emergency Valves. — In large power plants it is customary to protect the various divisions of the steam piping by emergency valves which may be closed by suitable means at any reasonable distance from the valve. The simplest form of emergency stop is a weighted " butterfly " valve, which is to all intents and purposes a weighted check as illustrated in Fig. 385 (F) . The weight when supported, say Fig. 380. Anderson Non-Return Valve. PIPING AND PIPE FITTINGS 659 by a cord and pulley, holds the valve open; when the cord is cut or released the weight drops and forces the valve shut. The cord may lead to any convenient and safe distance from the valve. In applying this system of control to steam engines the valve is placed in the steam pipe just above the throttle and the weight held up by a lever controlled by the main governor or preferably by a separate gov- ernor. Should the engine exceed a certain speed, as in case of accident to the regular governor, the lever supporting the weight is tripped by the emergency governor and the valve is closed automatically. For high pressures a rotating plug valve or cock is preferred to the butterfly type, since it is balanced in all positions. Gate and globe valves may be converted into emergency valves by 3 having the stems mechanically operated by electric motors, hydraulic pistons, and the like. Fig. 381 shows a section through a Crane hydraulically oper- ated emergency gate valve. A Fig. 381. Crane Emergenc y Valve , Hydraulic. Fig. 382. Anderson Triple- Duty Emergency Valve. Fig. 383. Pilot Valve for Anderson Triple-Duty Emergency Valve. Fig. 382 shows a partial section through an " Anderson triple- duty " emergency valve, and Fig. 383 a section through the pilot valve. A steam connection from the main line to the top of a copper diaphragm holds the pilot valve closed because of the large area above the diaphragm. A steam pipe connection from underneath the emergency piston of the triple-acting valve also leads to the pilot 660 STEAM POWER PLANT ENGINEERING valve. In case a break occurs in the main steam line or branches, the pressure is removed from the top of the pilot valve, causing it to open, thus exhausting the pressure from beneath the emergency piston in the triple-acting valve. The boiler pressure on top of the emer- gency piston causes the valve to close. Pilot valves may be located at any desirable places, thus affording control from different points. In the " Locke automatic engine stop system " the stop valve is operated by an electric motor which is controlled by contact points operated by a speed-limit device. (See Power, August, 1907, p. 471, for a detailed description.) 348. Check Valves. — Fig. 384, A to D, illustrate the different types of check valves in most common use. A is a ball check, B a cup f."'- A'. '1 / \ b sv>ssv l V^jSgr C^&p^ZZZl Tby^rfp "~ "" ~&£$srfff7 (A) (B) (C) Fig. 384. Types of Check Valves. (oj or disk check, C a swing check, and D a weighted check. Occasionally the valve body is fitted with a valve stem and handle for holding the disk against its seat, in which it is designated as a stop check. In A and B the valve seat is parallel to the direction of flow and the valve is held in place by its own weight and by the pressure of the fluid in case of reverse flow. In the swing check the seat is at an angle of about 45 degrees to the direction of flow. The latter construction is preferred as it offers less resistance to flow and there is less tendency for impurities to lodge on the valve seat. By extending the hinge of the swing through the body of the valve, a lever and weight may be attached as in D and the check will not open except at a pressure corresponding to the resistance of the weight. It thus acts as a relief valve and at the same time prevents a reversal of flow. Stop checks are usually inserted in boiler feed lines close to the boiler, and when locked, act as any ordinary stop valve and permit the piping to be dis- mantled or the regulating valve to be reground without lowering the pressure on the boiler. Since the wear on check valves is excessive and necessitates frequent regrinding they are often mounted with regrinding disks, Fig. 384 (C), which may be " ground " against the seat without removing the valve from the line. PIPING AND PIPE FITTINGS 661 349. Blow-off Cocks and Valves. — The requirements of a good blow-off valve are that it shall furnish a free passage for scale and sediment, that it shall close tightly so as not to leak, and that it shall open easily without sticking or cutting. On account of the rather severe service to which such valves are subjected, they are made very heavy, with renewable wearing parts. Fig. 385 gives a sectional view of a Crane ferrosteel valve. The bonnet is easily taken off and the disk removed to be refaced or replaced by a new one. The old disk is repaired by pouring in a hard Babbitt metal and facing it off flush. The seats are of brass and oval on top to prevent scale lodging between them and the disk, and are so made that they may be removed; but it has been found in practice that there is not much cutting of the seat, the damage usually being confined to the softer Babbitt metal which faces the disk. Fig. 385. Crane Ferrosteel Fig. 387. Blow-off Valve. A Typical Blow-off Cock. Fig. 386. Faber Blow-off Valve. Fig. 386 gives a sectional view of a Faber valve. When the disk, which makes a snug fit in the body of the valve, is in the position shown, the boiler discharge is practically shut off and any sediment lying on the seat is cleaned off by a jet of steam or water. Fig. 387 shows a section through a typical blow-off cock of the straightway taper plug pattern with self-locking cam. Plug cocks are often used instead of valves on the blow-off piping. Current practice recommends the use of two valves, or rather one valve and one cock, in the blow-off line of each boiler. In most of 662 STEAM POWER PLANT ENGINEERING S— I L- „_.# ^L If! V- [ ■C ^ r P d < z s 2 X u z 10 1 I * 1 , "0 J 1 f z S — i — i I o \ LI 1" ■v t^M LI AC < hi M _l K «^ — Cr4 — » » i/ 1 ■ '- 1 ^ ! K ! ! ■ —.0-.li » -91 . L LJ 8 -Or*—} Hi ' '* l* ^-.o-.n pld -L ' 1 PIPING AND PIPE FITTINGS 663 the large stations a blow-off valve and a blow-off cock are installed as indicated in Fig. 388. The number and size of blow-off cocks are usually specified by city or state legislation. 350. Safety Valves. — Fig. 389 shows a section through the simplest form of safety valve. The valve is held on its seat against the boiler pressure by a cast-iron weight as indicated. This type has the ad- vantage of great simplicity, and can be least affected by tampering, since it requires so much weight that any additional amount which would seriously overload it can be quickly detected. For high pres- sure and large sizes of boiler this class of valve is entirely too cum- bersome. Fig. 390 shows the general de- tails of the common lever safety valve. The valve is held against its seat by a loaded lever, thereby enabling the use of a much smaller weight than the " dead weight " type, since the resistance is multiplied by the ratio of the long arm of the lever to Fig. 389. " Dead-weight Safety Valve. Fig. 390. Common Lever Safetv Valve. the short one. The proper position of the weight is determined by simple proportion. Fig. 391 shows a section through a typical pop safety valve in which the boiler pressure is resisted by a spring. This type of valve has practically supplanted all other forms. The boiler pressure acting upon the under side of valve V is resisted by the tension in spring S. As soon as the boiler pressure exceeds the resistance of the spring the valve lifts from its seat and the steam escapes through opening 0. The static pressure of the steam plus the force of its reaction in being deflected from the surface A holds the valve open until the pressure in 664 STEAM POWER PLANT ENGINEERING the boiler drops about 5 pounds below that at which the valve is lifted. The additional area of valve exposed to pressure when the valve lifts causes it to open with a sudden motion which has given it its name, and it also closes suddenly when the pressure has fallen. These valves are arranged so that the spring tension may be varied without taking them apart, and pro- vision is made for lifting the seats by means of a lever. The seats are of solid nickel in the best designs, to minimize corrosion. The commercial rating of a safety valve is based upon the area exposed to pressure when the valve is closed. The number and size of safety valves for a given boiler are ordinarily spec- ified by city or state legislation. The logical method for determining the size of safety valves is to make the actual opening at discharge sufficient to take care of all steam generated at maximum load. Most rules, however, are empir- ical and based on the extent of grate surface, thus: According to the Boiler Inspection Department Philadelphia, 22.5 G BOILER CONNECTION Fig. 391. Consolidated Pop Safety Valve. of the city of (166) in which P + 8 -6 A = area of combined safety valves, inches. G = Grate area, square feet. P = boiler pressure, pounds per square inch gauge. According to the rule of United States Supervising Inspectors of Steam Vessels,* q A = — , for lever safety valves. A = C A = — , f or pop safety valves. o Other notations as in (166). Hutton's rule is (" Steam-Boiler Construction/') p. 470 : 4 G Vp ' All notations as above. * Superseded 1908 by the following: A as above A - 0.2074 £ W = weight of steam per hour, lbs. P = Absolute steam pressure. See Power, Mar. 9, 1908, p. 480; Mar. 16, 1909, p. 520. (167) (169) (169) PIPING AND PIPE FITTINGS 665 The Consolidated Safety Valve Company's circular gives the fol- lowing rated capacity of its nickel-seat pop safety valves: Sizes in Inches. 1 1* 1* 2 H 3 3* 4 4* 5 5$ 6 Boiler Horse-power — From 8 10 10 15 20 30 35 50 60 75 75 100 100 125 125 150 150 175 175 200 200 275 275 To 300 351. Back-Pressure and Atmospheric Relief Valves. — These valves are for the purpose of preventing excessive back pressure in exhaust pipes. In non-condensing plants such valves are designated as back- pressure valves and in condensing plants as atmospheric relief valves. In the former the valve is usually adjusted so that a pressure of one to Fig. 392. Foster Back-Pressure Valve. Fig. 393. INLET Davis Back-Pressure Valve. five pounds above the atmosphere is necessary to lift it from its seat; in the latter the valve lifts at about atmospheric pressure. They are practically identical in construction, differing only in minor details. A slight leakage in the back-pressure valve is of small consequence, but in an atmospheric relief valve it may seriously affect the degree of vacuum and throw unnecessary work upon the air pump, hence it is customary to "water-seal" the latter. Fig. 392 shows a section through a typical back-pressure valve. The valve proper consists of a single disk moving vertically. The valve stem is in the form of a piston or dashpot which prevents sudden closing or hammering. The pressure holding the valve against its seat is regulated by a spring. When the back pressure becomes greater than atmospheric plus that added by the spring, the valve raises from its seat and relieves it. Fig. 393 shows a section through a Davis back-pressure valve, in which the resisting pressure is varied by means of a lever and weight. 666 STEAM POWER PLANT ENGINEERING Fig. 363 shows the application of a back-pressure valve to a typical heating system. Fig. 394 shows a section through a typical atmospheric relief valve. Opening B is connected to the exhaust pipe and opening A leads to the atmosphere. Under normal conditions of operation atmospheric pressure holds valve V against its seat. Water in groove S " water- seals " the seat and prevents air from being drawn into the condenser. In case the pressure in pipe B becomes greater than atmospheric it lifts valve V from its seat and is relieved. Piston P acts as a dash- pot and prevents the valve from slamming. Fig. 395 shows a section through an atmospheric relief valve in which the weight of the valve is counterbalanced or even over- balanced by an adjustable weight and lever, thereby permitting the valve to open at or below atmospheric pressure, as may be desired. Fig. 394. Crane Atmospheric Relief Valve. Fig. 395. Acton Atmospheric Relief Valve. 352. Reducing Valves. — It is often necessary to provide steam at different pressures in the same plant, as in the case of a combined power and heating plant. To effect this result the reduction in pres- sure is accomplished by passing the steam through a reducing valve, which is but an automatically operated throttle valve. There are many different forms, the operation of all being based upon the same general principles. In the Kieley valve, Fig. 396, the low-pressure steam acts upon the top of flexible diaphragm D, and the weighted lever L (which may be adjusted to give the desired reduction in pressure) acts upon the other side. The movement of the diaphragm causes the balanced valve V PIPING AND PIPE FITTINGS 667 at the upper end of the spindle to open or close, as may be necessary to maintain the desired lower pressure. Inertia weights T and C prevent chattering. Fig. 396. Kieley Reducing Valve. Fig. 397. Foster Pressure Regulator. Fig. 397 shows a section through a class G Foster pressure regulator or reducing valve. In operation, steam enters at A and passes through the main valve port H to the outlet B. Steam at initial pressure passes through port C to chamber P and thence to the top of piston T through port L, opening the main valve U. Steam at delivery pressure passes through E and raises the diaphragm V against the pressure of spring R, allowing spring W to close the aux- iliary valve X. The pressure in chamber J is then equalized by the reduced pressure in ports G and the under side of piston X, and thus allows spring Y to close the main valve, which is then held to its seat by the initial pressure. Any reduction in delivery pressure is trans- mitted to diaphragm V, and permits spring to open auxiliary valve X, thereby admitting steam to the top of piston T, as previously explained. The delivery pressure is adjusted by screw D; thus increasing the tension of spring R increases the discharge pressure and vice versa. The adjustment once made, the delivery pressure 668 STEAM POWER PLANT ENGINEERING will remain constant, regardless of any variable volume of discharge or of the initial pressure, so long as the latter is in excess of the delivery pressure. W, Fig. 366, shows the application of a reducing valve to an exhaust steam heating system. Live steam is led to the valve through pipe A. It will be noted that the pipe leading from the valve to the heating system is much larger than the high-pressure supply pipe on account of the increase in volume of the low-pressure steam. Reducing valves should always be by-passed to permit of repairs without shutting down the system. Care should be taken in not selecting too large a reducing valve, as the valve lift is very small and the larger the valve the less will be the lift for a given weight of flow and consequently the greater the wire drawing and erosion of the valve seat. 353. Foot Valves. — Whenever a long column of water is to be moved in either suction or delivery pipe it is customary to place a check valve near the lower end of the column to prevent the water from backing up when the pump reverses or shuts down. The check valve placed at the end of the suction pipe is called a foot valve. Any check valve may be used as a foot valve, though practice limits the choice to the disk or flap type as illustrated in Fig. 398. To pre- vent rubbish from destroying the action, a strainer or screen is gener- Fig. 398. Types of Foot Valves. ally incorporated with the body of the valve. A, Fig. 398, illustrates a single-flap, B a multi-flap and C a disk valve composed of a nest of small rubber valves. The single-flap are usually made in sizes J to 6 inches, the multi-flap 7 to 16 inches, and the disk valve in all com- mercial sizes from f to 36 inches. For large sizes, 16 to 36 inches, the multi-disk valve is given preference, since a number of the disks may be disabled without destroying its operation. The Use and Abuse of Globe Valves: Power & Engr., Jan., 1909, p. 10. Gate Valves in Steam Pipe Lines: Power & Engr., Feb. 16, 1909, p. 320. Types of Check Valves and their Operation: Power & Engr., July 6, 1909, p. 11. CHAPTER XVL LUBRICANTS AND LUBRICATION. 354. General. — The losses due to the friction of the working parts of machinery include considerably more than the mere loss of power, namely, the depreciation resulting from wear of bearings, guides, and other rubbing surfaces, and the expense arising from accidents traceable to excessive friction. The power absorbed in overcoming friction varies with the type of plant and the character of machinery and is seldom less than 5 per cent and often greater than 30 per cent of the total power developed. In large central stations these losses approximate 8 per cent and in weaving and spinning mills will average as high as 25 per cent. (Trans. A.S.M.E., 6-465.) These figures refer to properly lubricated plants operating under normal conditions. The proper selection of lubricant is therefore a very important problem, since, besides the cost of the lubricant itself, the loss in power and in wear and tear to machinery is no small item. A change of lubricant may frequently result in marked increase in economy of operation. The lubricants most commonly met with in power plant practice are conveniently classified as oils, greases, and solids, and are of animal, mineral, or vegetable origin. Reference Books: Archbutt and Deeley, Lubrication and Lubricants; Redwood Lubricants; W. M. Davis, Friction and Lubrication; Gill, Oil Analysis; Robinson, Gas and Petroleum Engines; Thurston, Friction and Lost Work; Gill, Engine Room Chemistry. 355. Vegetable Oils. — Except for certain special purposes and for compounding with mineral oils these possess lubricating properties of little practical value, since they decompose at comparatively low tem- peratures and have a tendency to become thick and gummy. The vege- table oils sometimes employed are linseed, cottonseed, rape, and castor. Vegetable Oils: Power, May, 1906, p. 300; Archbutt and Deeley, Lubrication and Lubricants, p. 232; W. M. Davis, Friction and Lubrication, p. 28; Gill, Oil Analysis. 356. Animal Fats. — Many animal fats have greater lubricating power than pure mineral oils of corresponding viscosity but are objec- tionable on account of their unstable chemical composition. They decompose easily, especially in the presence of heat, and set free acids 669 670 STEAM POWER PLANT ENGINEERING which attack metals. They are seldom used in the pure state and are usually compounded with mineral oils. The animal products used in this connection are tallow, neat's-foot oil, lard, sperm, wool grease, and fish oil, the first named being the most important. In cylinder lubrication, especially in the presence of moisture, the addi- tion of 2 to 5 per cent of acidless tallow seems to make the oil adhere better to the metal surfaces and increases the lubricating effect, while the proportion is so small that ill effects from corrosion or gumming are scarcely perceptible. Animal Fats: Archbutt and Deeley, Lubrication and Lubricants, p. 323; Gill, Oil Analysis, p. 44; Wright, Analysis of Oils, p. 193; Andes, Animal Fats. 357. Mineral Oils. — These are all products of crude petroleum and form by far the greater part of all lubricants. They present a wider range of lubricating properties than those derived from animal or vegetable sources, the thinnest being more fluid than sperm and the thickest more viscous than fats and tallows. They are not easily oxidized, do not decompose, become rancid, or contain acids. Crude American petroleum of specific gravity 0.802 may yield the following commercial products. (" Gas and Petroleum Engines," W. Robinson.) Average Percentage. Specific Gravity. Boiling Point, Degrees F. Light Oils. C Cymogene Petroleum ether < Rhigolene r Gasoline ( C. Naphtha Petroleum spirit < B. Naphtha ( A. Naphtha (benz.) Burning oils, kerosene. jordtnaTy kerosene". '. '. '. '. Fuel oils {For making oil gas; fuel C Lubricating oils Heavy oils < Paraffin wax ( Residium traces 0.1 1-1.5 10 2-2.5 2-2.5 12-20 40-55 17.5 2 5-10 0.590 625-. 631 635-. 658 680-.700 717-.72 742-. 745 780-.785 800-.810 0.85 885-. 920 908 at 60 deg. F. 32 64 85-155 140-212 175-245 212-265 300-570 300-680 and up- wards 480 and upwards Mineral lubrication oils may be classified as (1) Distilled oils, which are produced by distillation from crude petroleum and made pale, amber colored, and transparent by treat- ment with acid and alkali. LUBRICANTS AND LUBRICATION 671 (2) Natural oils, which are prepared from crude petroleum, from which grit, suspended and tarry impurities have been removed. They are dark and opaque and are rich in lubricating properties. (3) Reduced oils, or heavy natural oils, from which the lighter hydro- carbons have been evaporated and from which the tarry residue has been removed by nitration. Mineral Lubricants: Engr. U.S., July 1, 1904, pp. 466, Vol. 44 (1907), pp. 241, 369, 542, 585; National Engr., Jan., 1905, p. 19; Eng. Mag., June, 1904, p. 455; Power, March, 1906, p. 146. 358. Solid Lubricants. — Dry graphite, soapstone, and mica are sometimes used as lubricants, though they are usually mixed with grease or oils. They cannot easily be squeezed or scraped from between the surfaces, and are consequently suitable where very great weights have to be carried on small areas and when the speed of rub- bing is not high. The coefficient of friction of such lubricants is high, and when economy of power is essential better results may be secured by the use of liberally proportioned rubbing surfaces and liquid lubri- cants. Under certain conditions of pressure and speed these lubri- cants will sustain, without injury to the surfaces, pressures under which no liquid would work. Graphite: Trans. A.S.M.E., 13-374; Engng., Aug. 16, 1907; Sci. Am., May 11, 1907; National Engr., Jan., 1904; Am. Mach., Dec, 1907, pp. 784, 934; Horseless Age, Jan., 1904, June 11, 1902, p. 712; Power, Dec., 1906, p. 758. 359. Greases. — Under this name may be included the various - compounds which consist of oils and fats thickened with sufficient soap to form, at ordinary temperatures, a more or less solid grease. Those usually employed are lime, soda, or lead soaps, made with various fats and oils. " Engine " greases are thickened with a soap made from tallow or lard oil and caustic soda, and often contain neat's-foot oil, beeswax, and the like. For exceptionally heavy pres- sures, graphite, soapstone, and mica are sometimes added to the grease. Greases: Jour. Eng. Soc. West. Penn., March, 1904, p. 112; Railroad Gazette, July 8, 1904, p. 131; St. Ry. Jour., July, 1905, p. 95; see also text-books given in references at beginning of chapter. 360. Qualifications of Good Lubricants. — A good lubricant should possess the following qualities: (1) Sufficient "body" to prevent the surfaces from coming into contact under conditions of maximum pressure. (2) Capacity for absorbing and carrying away heat. 672 STEAM POWER PLANT ENGINEERING (3) Low coefficient of friction. (4) Maximum fluidity consistent with the " body " required. (5) Freedom from any tendency to oxidize or gum. (6) A high " flash point " or temperature of vaporization and a low congealing or " freezing point." (7) Freedom from corrosive acids of either metallic or animal origin. Lubricating oils are identified by certain tests which «,re used by refiners in grading and classifying the oils and by consumers in buy- ing them. These tests usually cover the following: (1) Identification of the oil, whether a simple mineral, animal or vegetable oil or a mixture. (2) Density or gravity. (3) Viscosity. (4) Flash point. (5) Burning point, fire test. (6) Acidity. (7) Coefficient of friction. (8) Cold test. 361. Identification of Oil. — The chemical analysis of oils lies in the province of the chemist, but some of the characteristics may be readily determined by a few simple tests. To detect admixtures of fatty oils in mineral oil a small quantity is heated in a test tube for 15 minutes with small pieces of either metallic sodium or caustic potash. If fatty oil is present, saponification takes place and the soap formed will rise to the top as a semi-solid mass and the amount may be estimated. Tarry matter may be detected by dissolving a small quantity of oil in from 10 to 20 times its bulk of gasoline; the tar and other insoluble matter will separate and collect at the bottom. Oil Testing and Specifications: Power, May, 1904, p. 302, Vol. 24 (1904), pp. 139, 240, 302, 526, Vol. 26 (1906), pp. 145, 222, 300, 331, 407; Am. Mach., April 11, 1907, p. 525; Engr. U.S., Oct. 15, 1904, p. 724, Oct. 2, 1905, p. 657; Marine Engng., June, 1903, p. 303; Chem. Engr. Nov., 1905, p. 10, Dec, 1905, p. 87, Jan., 1906, p. 141; Am. Gas Light Jour., Jan. 23, 1905; U.S. Cons. Repts., June, 1905; Sci. Am. Sup., Jan. 14, 1905. 362. Gravity. — The density or specific gravity is conveniently determined by means of a hydrometer, which, in the oil trade, is graduated according to the Baume scale. The relationship between specific gravity and degrees Baume at a temperature of 60 degrees F. may be expressed: Specific gravity = 130 + degrees Baume LUBRICANTS AND LUBRICATION 673 Table 88 gives the specific gravity and gravity Baume of a number of lubricating oils. Gravity: Power, March, 1904, p. 139; Robinson, Gas and Petroleum Engines, p. 474; Archbutt and Deeley, Lubrication and Lubricants, pp. 172-185; W. M. Davis, Friction and Lubrication, p. 34. TABLE 88. SPECIFIC GRAVITY AND GRAVITY BAUME OF A NUMBER OF LUBRICANTS. Water Cylinder oil Cylinder oil Heavy engine oil. . Medium engine oil Light engine oil. . . Castor machine oil Lard oil Sperm oil Tallow oil Cottonseed oil.. . . Linseed oil Castor oil (pure) . . Palm oil Rape-seed oil Spindle oil Specific Grav- ity. 1.000 .9090 .8974 .9032 .9090 .8917 .8919 .9175 .8815 .9080 .9210 .9299 .9639 .9046 .9155 .8588 Gravity Baume\ 10 24. 26 25. 24 27 27 23 29 24. 22 19 15 25 23 33 Flash Test, Degrees F. 575 540 411 382 342 324 505 478 540 518 505 405 312 363. Viscosity. — Viscosity may be denned as the degree of fluidity or internal friction of an oil. It is sometimes called the " body." It is determined by a viscosimeter. There are a number of different instruments for this purpose but no recognized standard instrument or method, so that " viscosity " conveys no meaning unless the name of the instrument, the temperature, and the amount of oil tested are given. Nearly all instruments are of the orifice type; that is, the viscosity of an oil is taken as the number of seconds required for a given amount, usually 50 cubic centimeters, to flow through an orifice at a given temperature. By " specific viscosity " is meant the ratio of the time required for the oil to run out to that of an equal quantity of water at 60 degrees F. The viscosity of engine oils is usually taken at 70 degrees F. and of cylinder oils at 212 degrees F. Viscosity: Trans. A.S.M.E., 9-369; Engr., Lond., Sept. 7, 1906, p. 344, June 12, 1900, p. 633; Eng. Mag., June, 1907, p. 455; Machinery, May, 1903, p. 484; Power, May, 1904, p. 303, May, 1907, p. 293, March, 1906, p. 146. 364. Flash Point. — The flash point is determined by heating a sample of oil in an open or closed cup at the rate of 15 degrees F. per minute until a spark will ignite the vapor. The temperature at which 674 STEAM POWER PLANT ENGINEERING this occurs is the flash point. So much depends upon the extent of oil surface exposed, size of spark, distance spark is held from the oil at the time of ignition, and the dimensions of the cup, that there may be con- siderable variation in the flash point as obtained by different experi- menters. Flash Test: Power, April, 1906, p. 222; Robinson, Gas and Petroleum Engines, pp. 482-488; Archbutt and Deeley, Lubrication and Lubricants, pp. 187-191; W. M. Davis, Friction and Lubrication, p. 34; Gill, Oil Analysis, p. 36. 365. Burning Point, or Fire Test. — By continuing the application of heat and noting the temperature at which the oil takes, fire and continues to burn, the burning point is obtained. The higher the temperature under which the oil must work the higher the fire test required, so that it will not decompose or volatilize. Too high a fire test gives an oil that does not atomize readily enough to reach all parts of the cylinder. Consult references under " Flash Test." 366. Acidify. — The presence of free acid is determined by shaking up equal quantities of oil and water and testing with litmus paper. Another simple test is as follows: A small quantity of oil is placed in a test tube with a little cupric oxide (Cu 2 0) and subjected to a gentle heat for three or four hours. The reaction with the copper turns the solution green if fatty acid is present and blue if vegetable acid is present. Acidity: Power, April, 1906, p. 222; Archbutt and Deeley, Lubrication and Lubricants, pp. 215-218; Gill, Oil Analysis, p. 74. 367. Cold Test. — The " cold test " is the temperature at which the oil will just flow. The sample is solidified by means of a freezing mixture and the temperature noted when it softens sufficiently to flow. Cold Test: Robinson, Gas and Petroleum Engines, p. 481; Archbutt and Deeley, Lubrication and Lubricants, pp. 195, 200-6; W. M. Davis, Friction and Lubrication, p. 28; Gill, Oil Analysis, p. 34; Redwood, Lubricants, p. 3; Power, March, 1906, p. 146. 368. Friction Test. — The coefficient of friction as determined from friction-testing machines is useful in obtaining a comparison of oils under the test conditions, but gives little information concerning the action of the oil under the widely different conditions found in actual practice. Table 89 gives the physical properties of a number of lubricating oils, with their particular zone of application. Friction and Lubrication: Trans. A.S.M.E., 1-74, 6-136; Am. Mach., July 21, 1904, p. 956, Jan. 23, 1902, p. 113, Sept. 10, 1903, p. 1303; Am. Elecn., Nov., 1905, p. 557; Engr., Lond., June 19, 1903, p. 631; Power, Dec, 1905, p. 748; National Engineer, Jan., 1905, p. 19; Mech. Engr., Nov. 30, 1907; Pro. Inst. Civ. Engr., 1901, p. 146; Machinery, Aug., 1903, p. 631; Proc. A.S.M.E., Nov., 1909, p. 1099. LUBRICANTS AND LUBRICATION 675 369. Atmospheric Surface Lubrication. — In a general sense all journals, slides, and "atmospheric" surfaces should be lubricated with straight mineral oils (as free from paraffin as possible), except when in contact with considerable water, in which case it is advisable to add 20 to 30 per cent of lard oil. Vegetable oils, paraffin oils, and animal oils (except lard oil as above stated) are not recommended for general engine and dynamo service. The test requirements of a number of classes of lubricants are outlined in Table 89 and represent current practice. Bearings, guides, and all external rubbing surfaces may be lubricated in a number of ways. (1) They may be given an inter- mittent application of oil, as, for example, with an oil can; (2) they may be equipped with oil cups with restricted rates of feed; and (3) they may be flooded with oil. The relative lubricating values of the systems have been estimated approximately as follows {Power, December, 1905, p. 750) : Intermittent. . . . Restricted feed. Flooded bearing. Coefficient of Fric- tion. Comparative Value. 0.01 and greater 72 and less 0.01 to 0.012 79 to 86 0.00109 100 370. Intermittent Feed. — Intermittent applications are ordinarily limited to small journals, pins, and guides which are subject to light pressures and which do not easily permit of oil or grease cups, as, for example, parts of the valve gear of a Corliss engine, governors and link work. On account of the labor attached and the frequent doubt about the oil reaching the wearing surfaces this method of lubrication is limited as much as possible even in the smallest plants. 371. Restricted Feed. — In the- average power plant the major part of the lubrication is effected by means of oil cups which are filled at intervals by hand or by mechanical means, the oil being fed from the cup by drops, according to the requirements. 372. Oil Bath. — In large power plants the principal journals and wearing parts are supplied with a continuous flow of oil which com- pletely " floods " the rubbing surfaces. The oil is forced to the various parts either by gravity from an elevated tank or by pressure from a pump. After the oil leaves the bearings it flows into collecting pans, thence into a receiving and filtering tank, and finally is pumped back into an elevated reservoir and used over and over again. The little lost by leakage and depreciation is replenished by the addition of new oil to the system. 676 STEM! POWER PLANT ENGINEERING TABLE 89. PHYSICAL CHARACTERISTICS OF A NUMBER OF LUBRICANTS. (Power, December, 1905, p. 750.) Kind of Oil. Use and Adaptation. O 8 Q to $ Viscosity at 70 De- grees. High-pressure cylinder oil. For steam cylinders using dry steam at pressures from 110 to 210 pounds. 25 to 24.5 30 600 to 610 645 to 660 175 to 205 General cylinder oil . . For steam cylinders using dry steam at 75 to 100 pounds. For air compressor cylinders when made from steam-re- fined mineral stock and when viscosity is 200. 26 to 25.5 30 550 to 585 600 to 630 180 to 190 Wet cylinder oil. (Remark 1.) For use where the steam is moist, especially in compound and triple expansion engines. 25.8 to 25.3 30 560 to 585 600 to 630 150 to 185 Gas engine cylinder oil. (Remark 2.) For gas engine cylinders. Neu- tral mineral oil compounded with an insoluble soap to give body. 26.5 30 320 350 300 Automobile gas engine oil. (Remark 3.) For automobile gas engines and similar work. 29.5 30 430 485 195 Heavy engine and machinery oils. For heavy slides and bearings, shafting, and horizontal sur- faces. 30.5 to 29.5 30 400 440 to 450 170 to 195 General engine and machine oils. For high-speed dynamos and machines. 30.8 to 30 30 400 to 420 450 to 470 175 to 190 Fine and light machine oils. For fine work, from printing presses to sewing machines and typewriter oils. With a cold test of 25° to 28° and a viscosity of 140° this makes an excellent spindle oil. 32.5 to 30.2 30 400 440 110 to 160 Cutting and heat dis- sipating oils. (Remark 4.) For cutting tools, screw cutting and similar work. 27 to 23 30 410 to 420 475 to 480 210 to 175 For ice machinery 30.2 200 225 165 Wet service and marine oils. (Remark 4.) For marine service, or where a great deal of moisture must be handled. 28 30 430 475 230 They are used in special work and for heavy pressures mov- ing at slow velocities. Remark 1. — May contain not over 2 to 6 per cent of refined acidless tallow oil in the high- pressure oils and not over 6 to 12 per cent in the low-pressure oils. Remark 2. — The reason for using an insoluble soap such as oleate of aluminum is that it is impossible to decompose the soap with a high heat ; the soap, although not a lubricant, is a vehicle for carrying some oil. Remark 3. — Owing to a lack of body, this oil will not interfere with the sparking by depos- iting carbon on the platinum point. Remark 4. — May contain 30 to 40 per cent of pure strained lard oil. LUBRICANTS AND LUBRICATION 677 373. Oil Cups. — Fig. 399 illustrates the application of sight-feed oil cups to the crosshead and slides of a reciprocating engine. The oil is fed into the cups by hand and gravitates to the rubbing surfaces, the rate of flow being regulated by a needle valve. Cups A and B feed directly to the crosshead guides, but the oil from cup D flows to the bottom orifice 0, from which it is wiped by a metal- lic wick S and carried by gravity to the wrist pin. 374. Telescopic Oiler.— Pig. 400 shows the application of a tele- scopic oiler to a crosshead and guides. and C are sight-feed oil cups, the former feeding directly to the top guide through the tube S. The oil from C flows by gravity through the swing joint into the telescopic tubes P, R and thence to the pin through the lower swing joint as indicated. As Fig. 399. Oil Cup Lubrication, Hand Filled. Fig. 400. Nugent 's Telescopic Oiler. the crosshead moves back and forth, the pipe P slides into and out of pipe R, the oil being thus conducted directly to the pin without wasting. 678 STEAM POWER PLANT ENGINEERING A device of this type installed on a high-speed automatic engine at the Armour Institute of Technology has been in operation for three years without cost for repair or renewal. 375. Ring Oiler. — Small high-speed engines are often oiled by the oil-ring system ; as illustrated in Fig. 401. The shaft is encircled by M/MMUWBZZ* rj 4,^ . , / Fig. 412. Forced-Feed Cylinder Lubrication. 386. Siegrist System. — Fig. 413 shows an application of the Siegrist system of cylinder and engine lubrication. There are two storage tanks on the engine-room floor, one for cylinder oil and the other for engine oil, the distributing arrangements being the same in each case. The oil is pumped from each tank into a main pipe extending the length of the engine room and provided with branches at each point requiring lubrication. The oil pumps are actuated by steam and are of the duplex direct-acting type, provided with auto- matic governors which regulate the speed to suit the demand for oil. 686 STEAM POWER PLANT ENGINEERING 5-j 05 !T-^-r^i.-Jr-_t: ii-.l--J I i — — i _] 1 i V2H— i-. j ^b ■ h - r 'Fol LUBRICANTS AND LUBRICATION 687 TO STEAM FEED REGULATOR Fig. 414. Siegrist Sight- Feed Lubricator. The cylinder oil is forced through a special sight-feed lubricator, Fig. 407, under a pressure of about 25 pounds in excess of the steam pressure. Referring to Fig. 414, diaphragm valve D, in the bottom of the lubri- cator, is kept closed by the steam pressure admitted through pipes B. Thus the inlet pressure must be greater than that of the steam before the valve will open and admit oil to the engine. The oil, after enter- ing, passes upward through the sight- feed glass and downward through the hollow arm A to the steam pipe. The engine oil is forced by the pump to the various points under a pressure of about 20 pounds. The waste oil is caught in suitable re- ceptacles and, after being filtered, is returned to the storage tank by a steam pump. This pump is con- nected so that it can supply the storage tank either from the filter or with fresh oil from a large oil tank in the basement. By this arrangement all handling of oil in the engine room is done away with. 387. Oil Filters. — After oil has been applied to machinery its lubricating properties become impaired on account of (1) contami- nation with anti-lubricating material, such as dust, metallic particles from wear, gum, acid, and resin; and (2) exposure to heat and the atmosphere which drives off part of the more volatile constituents and decreases the fluidity of the oil. In many small plants no attempt is made to reclaim oil that has once been used, since the quantity is so small that the cost and trouble involved would more than offset the gain. Where large quan- tities of oil are used, considerable saving may be effected by using it over and over again. To render the oil fit for reuse it must be thoroughly purified. The anti-lubricating matter is removed by pre- cipitation and filtration. Fig. 415 shows a section through a " White Star" oil filter and purifier. The apparatus consists of a cylindrical sheet-iron vessel divided into two compartments by a vertical partition. These two compartments are connected near the top by valve B. The smaller chamber is provided with a funnel A and a steam coil for heating the contents. The large chamber contains a cylindrical wire screen covered with several folds of filtering cloth. Impure oil is poured into funnel A, the upper part of 688 STEAM POWER PLANT ENGINEERING WATER LEVEL WATER '_-_ ~ which is provided with a removable sieve or strainer, and is discharged below the surface of the water through holes in the foot of the tube. The thin streams of oil rise vertically to the surface of the water and the heavy particles of grit and dirt gravitate to the bottom. The steam coil heats the oil and water and facilitates precipita- tion of the solid matter by thinning out the streams of oil. When the oil in the smaller chamber reaches the level of valve B it flows in- to the filter bag, which re- moves the remaining im- purities and permits the purified products to flow into the -large compartment from which it may be drawn at will. All parts are access- ible and readily removed for cleaning purposes. The accumulated sediment in the bottom of the small chamber is dis- charged to waste at intervals by means of a suitable drain. When the PERFORATED PLATE FILTERING MATERIAL PERFORATED PLATE PERFORATED PLATE FILTERING MATERIAL PERFORATED PLATE WATER STEAM COILS Fig. 415. White Star Oil Filter. SECTION 1 SECTION 2 SECTION 3 Fig. 416. Turner Oil Filter. SECTION 4 filter cloth is to be removed, valve B is closed and the wire cylinder is disconnected and lifted out. Any oil remaining in the filter is returned to funnel A. The filter cloth is held against the screen by cords and hence is readily removed. LUBRICANTS AND LUBRICATION 689 Fig. 416 shows a section through a Turner oil filter, illustrating the type of filter usually installed in large stations where continuous fil- tration is desired. This apparatus consists of a rectangular tank divided into four compartments. The returns from the lubricating system flow into section 1 through a screened funnel and discharge into the water space at the bottom of the compartment. The oil rises through the water, passes, under pressure of the head in the funnel, through a layer of filtering material resting on a perforated plate, and collects in an inverted cone. Through perforations round the top of the cone it passes into a dirt chamber, where most of the heavy impuri- ties are deposited, and then, still rising, passes through another per- forated plate and more filtering material. The partially cleaned oil, which issues, overflows into the second compartment and thence into the third, the same cycle of operations being repeated in these two. The overflow from the third compartment descends through a final filter in the fourth compartment and collects at the bottom, from which it is withdrawn by the oil pump. Forced-Feed Lubrication: Am. Elecn., Aug., 1902, p. 402, Dec, 1905, p. 608; Automobile, Nov. 1, 1906, p. 572; Mech. Engr., April 20, 1907, p. 552. Cylinder Lubrication: Power, Dec, 1902, p. 30, Jan., 1905, p. 36, March, 1906, p. 163; Engr., Lond., 1905, Vol. 96, pp. 55, 108, 132, 155; St. Ry. Jour., June 22, 1907, p. 1103; Engr. U.S., Oct. 15, 1906, p. 682; Am. Gas Light Jour., Jan. 23, 1905, p. 130; Horseless Age, Sept. 24, 1902, p. 676. Miscellaneous. — Measurement of Durability of Lubricants: Trans. A.S.M.E., 11-1013. Valuation of Lubricant by Consumer: Trans. A.S.M.E., 6-437. Suit- ability of Lubricants: Power, Nov., 1906, p. 673. Oil Required for Lubricators: Elec World, May 5, 1906, p. 934. Gumming Tests: Jour. Am. Chem. Soc, April, 1902, p. 467. Valuation of Lubricants: Jour. Am. Chem. Ind., April 15, 1905, p. 315. Lubrication, General: Power, March, 1903, p. 135; Mech. Engr., June 30, 1906, p. 919; Prac Engr., Dec. 15, 1905, p. 915; Elec Engr., Lond., Sept. 7, 1906, p. 344. ■ Oil Purification: Elec. Engr., Lond., Jan. 13, 1903, p. 51; Elec. World, Dec. 1, 1906, p. 1053. Economy in Lubrication of Machinery: Trans. A.S.M.E., 4-315. Theory of Finance of Lubrication: Trans. A.S.M.E., 6-437. Experiments, Formulas, and Constants for Lubrication of Bearings; Am. Mach., Sept. 10, 1903, pp. 1281, 1316, 1350. Lubricators and Lubricants: Power & Engr., Sept. 21, 1909, p. 486. j Selection of an Oil for Lubrication: Power & Engr., July 27, 1909, p. 137. CHAPTER XVII. FINANCE AND ECONOMICS— COST OF POWER. 388. Records. — Few engineers realize the importance of a detailed system of accounting, or the saving which may be effected in cost of operation by careful study of the daily records of performance. Many regard graphical load curves, meter readings, and similar records as interesting but of little economic value. During the past few years the author has made a close study of the cost of power in a large num- ber of central and isolated stations in Chicago, and found, without exception, that the highest economy was effected by the engineers who kept the most systematic records; the poorest results were obtained where records were kept indifferently or not at all. In some small plants the numerous duties of the engineer prevented him from devoting the necessary time, but in the majority of cases the absence of records was due entirely to lack of interest. Power-plant records to be of value must be closely studied with a view to improvement. The mere accumulation of data to be filed away and never again referred to is a waste of time and money. Records should cover not only the daily operation of the plant but also, as permanent statistics, a complete analysis of each item of equipment. The value of such data cannot be overestimated. The engineer will frequently find it greatly to his interest to have avail- able at a moment's notice the complete details of his engines, boilers, generators, and other machinery, especially when it is required to renew a broken or worn-out part. 389. Output. — The periodical output of a power plant may be expressed in terms of (1) Steam plant. Indicated or brake horse power. Indicated or brake horse-power hours. (2) Electric lighting plant. Electrical horse power or kilowatts. Electrical horse-power hours or kilowatt hours. Lamp hours. 690 FINANCE AND ECONOMICS — COST OF POWER. 691 (3) Electric railway plant. Electrical horse power or kilowatts. Electrical horse-power hours or kilowatt hours. Car miles. When a plant is operating at practically constant load it is suffi- ciently accurate for most purposes to express the output in horse power or kilowatts per year. When the output fluctuates from day to day it is best expressed in horse-power hours or kilowatt hours, or by specifying the load factor along with the periodical output in horse power. For example, 1 horse power per year, 24 hours per day and 365 days per year, is equivalent to 365 X 24 = 8760 horse-power hours. If the full power is used throughout this time, it matters little whether the charge is based on horse power or horse-power hours; if, however, the power is used say only half the time, the yearly cost per horse power will remain unchanged but the cost per horse-power hour will be just double. As will be shown later the load factor (ratio of actual to rated load) exerts a marked influence on the cost of pro- ducing power, and for this reason the output is usually expressed as horse-power hours, kilowatt hours, lamp hours, or the like. 390. Load Factor. — The yearly load factor or simply load factor, as it is usually called, is the ratio of the actual yearly output to the rated yearly output measured on a 24-hour basis. Thus: For a steam plant, Load factor = Yearly output, horse-power hours . ( Rated horse power X 8760 For an electric station, Load factor = Yearly output kilowatt hours _ ( Rated capacity, kilowatts, X 8760 (8760 = number of hours in one year.) The curve load factor or station load factor is the ratio of the actual yearly output to the rated output, based upon the number of hours the plant is in actual operation. Thus for an electric station, Curve load factor- Yearly output, kilowatt hours _ ( Rated capacity X hours plant is in operation In any plant the great desideratum is a high load factor with con- sequent greatest return on the investment. All the factors of expense included in the cost of power are then operating at maximum economy. High peak loads and low average loads necessitate large machines which are but little used and greatly increase the fixed charges. 692 STEAM POWER PLANT ENGINEERING In any system the total fixed charges per year are constant irre- spective of the load factor, since interest, taxes, depreciation, insurance, and maintenance go on whether the plant is in operation or not. The total fixed charges for a specific case are illustrated in Fig. 417 by a 1.4 280000 cfr*^ 1.2 u 3 S V6 s* . v*> P°> 240000 \^c * ^ 3^ O w & 1.0 \V & ^2^ L, Dollar O ^ 0.8 <^ L* v ^i 1 8 tal Yearly Cos J# 2 g 0.6 o ^< Ay Oper v tij ost; «*?> Cer ts PerKw. Hr. kp^ 8 80000 o O 04 <5 0^ X S < — Total Fixed Charges, Dolls irs s >• 2 --i£ ^C harg . e « J >er Kfl. H-r 40000 uts 0.0 < ) 1 2 3 Ye 4 arly L0£ 5 idF acto G r-P erC 7 ent £ 9 1C Fig. 417. Influence of Load Factor on the Cost of Power at the Switchboard. (5000 Kilowatt Electric Light and Power Station.) straight line. The cost per kilowatt hour, however, decreases as the load factor increases. For example, with the plant operating con- tinuously at rated load (100 per cent load factor) the fixed charges per kilowatt hour are — ^955 — = $0.00148. 5000 X 8760 With 30 per cent load factor these charges are 65 ' 0QQ = $0.00445 kilowatt hour. 0.3 (5000 X 8760) The higher the load-factor the greater is the amount of power produced and the longer does the apparatus work at best efficiency. But the greater the power produced the larger will be the fuel consumption and the oil and supply requirements. The labor charges will be prac- tically constant. The total operating cost per year increases as the load factor increases, but not directly. (See Fig. 417.) The cost per FINANCE AND ECONOMICS — COST OF POWER 693 kilowatt hour, however, decreases as the load factor increases. For example, the operating costs per year with plant operating contin- uously at full load are $230,200. This gives 230 ' 2QQ — = $0.00525 per kilowatt hour. 5000 X 8760 F With 30 per cent load factor the yearly operating charges are $87,980, which gives , 87,98Q = $0.0067 per kilowatt hour. 0.3 (5000 X 8760) F Table 107 shows the influence of the load factor on the cost of power in two isolated stations of the same rated capacity, one operat- ing with the unusually high load factor of 80 per cent and the other operating with the low load factor of 17 per cent. The former fur- nishes current for a large electro-chemical concern in which the load is practically constant. In general, the higher the load factor the greater becomes the ratio of the operating to the fixed charges, and extra investment may become advisable to secure the greatest economy possible. On the other hand, when the load factor is low the fixed charges are the governing factor in the cost of power, and extra expenditures must be carefully considered, particularly if fuel is cheap. 391. Cost of Operation. — The cost of operation of power plants is conveniently divided into two parts: (1) Fixed charges. (a) Investment costs. (b) Administration costs. (2) Operating costs. 392. Fixed Charges. — These cover all expenses which do not expand and contract with the output. In very large plants they are usually divided into two parts, (a) the investment costs, which include interest, rental, depreciation, taxes, and insurance, and a reserve fund to cover depreciation of the investment, and (b) the administration costs, which include rental of offices, annual salaries of officers, and all other expenses not directly chargeable to the power plant. In the average plant the fixed charges comprise interest, rental, depreciation, taxes, insurance, and sometimes maintenance, though the latter is ordinarily included in the operating costs. 393. Interest. — The rates of interest on borrowed money vary with the nature of the security. In the case of power plants the form of security is usually a mortgage on the plant and equipment. If a 694 STEAM POWER PLANT ENGINEERING builder has sufficient funds to construct the plant without borrowing, he should charge against the item " interest " the income which the sum involved would bring if placed out at interest or if invested in his own business. In estimating the interest charges 5 per cent of the capital invested is ordinarily assumed unless specific figures are available. TABLE 90. APPROXIMATE USEFUL LIFE OF VARIOUS PORTIONS OF STEAM POWER PLANT EQUIPMENTS. Years. Buildings, brick or concrete 50 Buildings, wooden or sheet iron 15 Chimneys, brick '. : 50 Chimneys, self-sustaining steel 25 Chimneys, guyed sheet-iron 10 Boilers, water-tube 25 Boilers, fire-tube 15 Engines, slow-speed 25 Engines, high-speed 15 Turbines 25 Generators, direct-current 25 Generators, alternating-current 30 Motors 20 Pumps 25 Condensers, jet 35 Condensers, surface 20 Heaters, open 30 Heaters, closed 20 Economizers 20 Wiring 20 Belts 7 Coal conveyor, bucket 15 Coal conveyor, belt 10 Transformers, stationary 30 Rotary converters 25 Storage batteries 15 Piping, ordinary 12 Piping, first class 20 NOTE. — So much depends upon the design and the conditions of operation that no fixed values can be definitely assigned and the above figures should be used with caution. Practice shows that most power-plant appliances become obsolete long before the limit of their useful life is reached. 394. Depreciation. — This charge represents the gradual deterio- ration of a plant, resulting in its eventually wearing out. It is also assumed to represent the superannuation of a plant or the rate at which the apparatus is becoming obsolete. Thus, under the first assumption, if the useful life of an engine is 40 years, the rate of depreciation, neglecting interest, is 2.5 per cent; if, however, it is assumed that the engine will become obsolete in 20 years and uneco- nomical for further operation, the rate of depreciation will be 5 per cent. It is difficult to assign a fixed rate of depreciation against any FINANCE AND ECONOMICS — COST OF POWER 695 piece of apparatus, due to possible new developments which cannot be reckoned with in advance in computing the useful life of the appa- ratus. Again, depreciation cannot always be separated from current repairs and is a variable factor even in the parts of the same machine. It is therefore more or less of an approximation. The average life of various parts of a steam power plant is outlined in Table 90, but on account of the inability to assign fixed values to the useful life of any apparatus, and on account of the great number of appliances in even a small plant, it is customary to charge a fixed rate of depreciation against the entire plant and thus avoid confusion and complexity. This very crude method usually results in overestimation in well- designed, well- operated plants and underestimation in poorly designed and badly managed installations. One of the largest power plant designing concerns in Chicago charges 7J per cent against deprecia- tion and finds this figure none too small. The Pennsylvania Railroad uses 7 per cent to cover depreciation charges on their power-house equipments. TABLE 91. RATE OF DEPRECIATION. (Per Cent of First Cost.) Rate of Interest, per Cent. 2 2.5 3 8.5 4 4.5 5 5.5 6 7 8 9 10 2 49.50 49.37 49.27 49.14 49.02 48.90 48.78 48.66 48.54 48.31 48.07 47.84 47.62 3 32.67 32.51 32.35 32.19 32.03 31.87 31.72 31.56 31.41 31.10 30.80 30.51 30.21 4 24.26 24.08 23.90 23.72 23.55 23.39 23.20 23.03 22.86 22.52 22.19 21.84 21.55 5 19.21 19.02 18.83 18.65 18.46 18.28 18.10 17.91 17.73 17.40 17.04 16.73 16.37 6 15.85 15.65 15.46 15.26 15.08 14.89 14.70 14.52 14.33 13.97 13.63 13.29 12.96 B9 7 13.45 13.25 13.05 12.85 12.66 12.46 12.28 12.09 11.91 11.15 11.20 10.87 10.55 3 8 11.65 11.44 11.24 11.05 10.85 10.66 10.47 10.28 10.10 9.74 9.40 9.06 8.74 3 9 10.25 10.04 9.84 9.64 9.45 9.26 9.07 8.88 8.70 8.34 8.00 7.68 7.36 10 9.13 8.92 8.72 8.52 8.33 8.14 7.95 7.76 7.58 7.23 6.90 6.58 6.27 < 11 8.21 8.01 7.80 7.61 7.41 7.22 7.04 6.85 6.68 6.33 6.00 5.69 5.40 o 12 7.45 7.25 7.04 6.85 6.65 6.46 6.28 6.10 5.92 5.60 5.27 4.97 4.69 >s J OS O 5 W ? 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ooooooooooo !CNJ(MO-nx3MniOHO >ooooooooooo KNMO*CDOO. t^ 00 00 00 c oooooooioioooooot OOOOOO :££o3o38888 •a^j '^uiqj jo azig «Si28g|S|||S| . £ 01 I* I 2 w © •I ^ .Z © 9 - 2 o o §JS6 +3 V. 01 , . o> 03 S-° £•2° 11 ^ jo — "03 c 2 •2o 2 sH ?3 O^ o> T3 c« o q . o o o3 .2 > «*-l 03 0> 03 0«-( 0> 3 oj £ £?$? J— o> 60 03 Sr 1 ^ T3 o 01 ^>TJ S § l a ©J £ '"S " ^- s- s ^ g o3 ^ &s S 5 ® «c « a t- © I °s S3 FINANCE AND ECONOMICS — COST OF POWER 703 determined by dividing the total water fed into the boiler per year by the total consumption of coal gave only 5.2 pounds. Current practice gives an average efficiency (based on yearly operation) of boiler and furnace of 70 per cent for pumping stations running at practically full load, 65 per cent for large lighting and power stations with yearly load factor of 0.50 or more, and 60 per cent for similar stations with load factor between 0.35 and 0.45. For very low load factors, 0.25 and under (as in connection with manufacturing plants, tall office building and other plants operating on a 10-hour basis) this efficiency seldom exceeds 50 per cent. With these figures as a guide the cost of fuel per unit output may be roughly approximated. 400. Oil, Waste, and Supplies. — These items approximate 2 to 10 per cent of the total operating expenses. Tables 95 to 107 give some idea of current practice in different classes of power plants. 401. Repairs and Maintenance. — This item ordinarily refers to the cost of keeping the plant in running order over and above the cost of labor or attendance, and depends upon the age and condition of the plant and the efficiency of the employees. Tables 95 to 107 give the cost of repairs and maintenance for a wide range in power-plant practice. 402. Cost of Power. — The actual cost of producing power depends upon the geographical location of the plant, the size of apparatus, the design, conditions of loading, system of distribution, and the method of accounting. Tables 95 to 108 compiled from various sources give the detailed costs of a large number of central and isolated stations. Table 95. Operating costs per kilowatt hour for a number of typical British electric light and power plants. Table 96. Operating costs per kilowatt hour for a number of United States electric power plants for street-railway, light, and general power service. Table 97. Average operating costs per kilowatt hour for all stations of the Boston Elevated. Table 98. Operating costs for the year 1907 of the mechanical plant of the First National Bank Building, Chicago. Table 99. Costs, fixed and operating, of producing one brake horse power per year, simple non-condensing engine, etc. Table 100. Cost of power for compound condensing engine plants. Table 101. Costs of different sizes and types of plants and annual costs per brake horse power, average working conditions. Tables 102, 103, and 104. Cost, fixed and operating, of producing one electrical horse power per year for different sizes and types of plants. 704 STEAM POWER PLANT ENGINEERING Table 105. Influence of load factor on cost of electrical power in isolated stations. Tables 106 and 107. Cost of furnishing heat, light, and power for a number of isolated stations in New York City, tall office buildings, loft buildings, apartment houses, hotels, and club buildings. 135 130 125 120 115 03 | 110 pq £ 105 o W 100 90 ^ Curves showing Range in Cost of Power in 200 Mfg. Plants Middle-Western States 1 \ [ \ \ \ \ \ \ I \ \ \ \ \ ^ /, 0l ;& •flyj ' llllllil Size of Plant, Horse -Power Fig. 418. o o 8 g Table 108. Cost of furnishing heat, light, and power for the year 1907, First National Bank Building, Chicago. FINANCE AND ECONOMICS — COST OF POWER 705 20 40 60 PER CENT LOAD-FACTOR Fig. 418a. Cost of Power in Large Power Plants with Maximum Load over 30,000 Kilowatts. Coal at $3.00 per Ton. 14,500 B.T.U. per Pound. 706 STEAM POWER PLANT ENGINEERING 130 50 l.=Keciprocating steam-plant. 2.=- Steam-turbine plant. 3. -Reciprocating-engine and low-pressure turbine plant. 4. = Gas-engine plant. 5 = Gas-engine and steam-turbine plant. 6 . = Hydraulic-plant \y 2^ 5^ ^ ^ ^ 1F^ ^2 20 Per Cent Load-Factor Fig. 418b. Cost of Power in Large Power Plants with Maximum Load over 30,000 kilowatts. Coal at $1.50 per Ton. 11,000 B.T.U. per Pound. FINANCE AND ECONOMICS — COST OF POWER 707 Table 94a, and Figs. 418a, 418b and 418c give the fundamental relations between the various items entering into the cost of power for various types of plants of over 30,000 kilowatts rated capacity. These data are taken from a paper presented by H. G. Stott at a meeting of the Toronto section of the American Institute of Electrical Engineers, Toronto, Ont., December, 1908. The figures have been brought up to date (June, 1910) by Mr. Stott and show what is actually being done to-day in large plants of the size stated above. TABLE 94a. DISTRIBUTION OF MAINTENANCE AND OPERATION COSTS IN POWER PLANTS HAVING A MAXIMUM LOAD OVER 30,000 KILOWATTS. . (H. G. Stott.) Recip- rocating Steam Plant. Steam Tur- bine Plant. Recip- rocating Engines and Low- pressure Steam Tur- bines. Gas En- gine Plant. Gas En- gines and Steam Tur- bines. Hy- drau- lic. Maintenance. 1. Engine room, mechanical 2. Boiler or producer room 2.59 4.65 0.58 1.13 61.70 7.20 6.75 7.20 2.28 1.07 2.54 1.78 0.30 0.17 100.00 100.00 125.00 11% 0.51 4.33 0.54 1.13 55.53 0.65 1.36 6.74 2.13 0.95 2.54 0.35 0.30 0.17 77.23 75.00 93.75 11% 1.55 3.55 0.44 1.13 46.48 0.61 4.06 5.50 1.75 0.81 2.54 1.02 0.30 0.17 69.91 80.00 100.00 11% 5.18 1.16 0.29 1.13 26.52 3.60 6.76 1.81 1.14 0.54 2.54 1.80 0.30 0.17 52.94 110.00 137.50 12% 2.84 1.97 0.29 1.13 25.97 2.16 4.06 3.05 1.14 0.54 2.54 1.07 0.30 0.17 47.23 96.20 120.00 11.5% 0.51 3. Coal- and ash-handling apparatus. 4. Electrical apparatus 1 13 Operation. 5. Coal 6. Water 7. Engine room, labor 1.36 8. Boiler or producer room, labor . . . 9. Coal- and ash-handling, labor .... 10. Ash removal 11. Electrical labor 2.54 12. Engine room, lubrication 13. Engine room, waste, etc 14. Boiler room, lubrication, etc Relative operating cost, per cent Relative investment, per cent ........ Probable average cost per kilowatt . . . Probable fixed charges 0.20 0.20 5.94 100.00 125.00 11% For steam-turbine plants larger than 60,000 kw. the cost per kilowatt may be reduced to $75.00. 08 STEAM POWER PLANT ENGINEERING i 1 1 1 FIXED CHARGES 1 1 1 i 10 1. Reciprocating steam plant. Cost $125.00 per Kw. .1 I*' 1 J 14 2| " $75.00 4 » 3. Eng no and low pressure turb nc plant. 12 4. Cost poo.uu per Kw. Gas engine plant. Cost $137.50 per Kw. 5. Gas engine and steam turbine plant. 10 CO UJ o 1\ 6. Cost Hyd $15*0.00 per raulic plan Kw. 1 t. Cost $125.00 I per Kw. < llV 1 1 TOTAL OPERATING CHARGES | 8 o Q |\ 1-6 A l'P' s above, coal (& $3.00- dittO «nn.lL CO o 2 u CO °g a HO. O U U uj L. O) a. u_ u si O 5 o o u U.CA Off Id t-V- W< O Id OI O SIMPLE non-ccnd, H.B, 100 30 100 2 2880 2300 1000 1292 92 F.T. 2 02 1904 2680 280 160 COMPOUND NON-CCNB. M.6, 120 24 100 2 4600 2300 1000 1430 74 F.T. 2 74 1840 2510 230 130 COMPOUND CONO. H. 8. 120 20 100 2 4600 2300 1000 1400 1430 66 F.T. 2 66 1500 2414 200 125 DE I AVAL TURBINE 120 18.6 100 2 7400 1000 1200 910 62 F.T. 2 62 1425 2360 200 120 © o CM SIMPLE NON-COftO. H.8. 100 30 100 3 4290 3450 1500 1830 185 F.T. 3 93 2&50 3850 530 235 COMPOUND NON-COND. M.S. 120 24 1C0 3 C300 8450 1500 2040|l46 F.T. 3 74 2460 3600 440 200 COMPOUND CONO. M. 6. 120 20 100 3 6900 8450 1500 2060 2040 132 F.T. 3 66 2250 3440 400 180 OE LAVAL TURBINE 120 18.6 100 3 11100 1500 1940 1270 12j F.T. 2 124 2480 2825 390 180 o § o o SIMPLE NON-COND. M. 8. 100 30 200 3 7020 6630 3000 2450 370 F.T 4 123 4920 5312 1000 400 COMPOUND NON-COND. M.S. 120 24 200 3 10620 6630 3000 2698 1297 F.T. 4 99 3960 5098 840 335 COMPOUND COND. H. 8. 120 20 200 3 10620 6630 3000 3700 2700 |264 F.T. 4 88 3670 4970 740 312 OE LAVAL TURBINE 120 17 200 3 18960 8000 3250 1620 22c F.T. 3 113 3390 4000 650 275 VERTICAL COND. L. 6. 150 13 400 2 20400 10600 4000 2000 4300 153 W.T 2 153 6100 2300 500 665 200 HOR. COND.L.S. 150 13 400 2 13600 15400 4000 2600 5440 153 W.T. 2 153 6100 2300 500 6C5 200 SIMPLE NON-C0N0. M.S. 100 30 200 4 9360 8840 4000 3190 555 FT. 6 111 6660 . 7070 1460 560 COMPOUND NON-COND. M.8. 120 24 220 4 141C0 8840 4000 3512 445 F.T. 5 111 5550 6440 1200 4U0 COMPOUND COND H.8. 120 20 200 4 14160 8840 4000 5050 3512 397 F.T 5 99 4950 6296 1070 420 OE LAVAL TUROI" 1 ? T20 17 2u0 4 25280 4000 4500 2094 339 F.T. 4 113 4520 5240 930 380 VERTICAL COND. L. 6. 150 13 300 3 29550 12600 4500 3600 6225 230 W.l 2 230 8460 2530 650 1150 280 HOR. COND. 1. 8. 150 13 300 3 19800 19920 4500 3600 7CC0 i3& W.T 2 230 8460 2533 650 1150 2$0 PARbONS TORPINF 150 13 COO 2 35010 6000 3600 1470 230 W.T 2 230 8460 2530 650 1150 280 o o CM PIMPIE NON-COND. M.8. 100 30 oeNoooioso)00 HHrtHHHrtlHHHINIM t^t^CO>--ii-h.-iO©000000000 OOOOiOOOONNN>OWW 00000000000000 OO^HO«OeOiOiCTti»Ot^t^t^T-i»o OlOHHHNOOOOHHrtNN OOOOOOOOOOOOOO OOOOOTJiTjH^OtO-^OCMOO !>. t- t>- t>- t>. 00 rH_ O »0 O CO CO CO CN .Tto«:oeotoo 10 0>C3®NNC>50ieo©OONO Tt< >-ii-icsieqoo«oaoocMt^eo»-ie3;Ot»'* «oosoOO"3»N00 0i0s»O!O i-tr-*»-«l-I^Hi-ll-lT-H-(i-l^i>. ^i>- j>^ cq c co c<^ »o >o c^ 05eo©t~i-l HIMWN N^OOOSCO— i eo «© t— i as oooooooooooooo oooooooooooooo » cq h n n n o o> q oo q q So •ot>-oo>ooo«ooo>oioo< t-I^Hi-ICO»0(>»0*0so^ooaJ3j"5«ocoicio lOiMOQOOOlrtO^CO iN*«oecofirt6o COr;-HNCi3(NMDN C^l C4C4C4 ^h , OhO0U*h( 00 O OS 00 OS •'f I _ JONOOt^'^eOOCSOOt^^. >oo i— -asi*'os»-iOOOO o o< > t^oo»ooo' i—i i—i l— ( CO U3 1 cj to « S3 3 « o-3 8-2 o o +- c 3 "J I C S3 Q.S'3 0>'3 s3 -g »Ses tH > ^dg C «2 d « § u 5 > bp ^ m b. O s a) *> «+» * S3 £ O '■5 , C"3 J ss 718 STEAM POWER PLANT ENGINEERING •s^uao — jajj-eg; Jadoox^ianj v[%va jnojj "Ai^; jad sasuadxg; i^oj, s^uaQ — jajj^gjad s^uao qi jan^ q^iAi jnojj 'A^X J3d sasuadxg p^oj, s;u93 — jajj^gjad s^uaoosianj^iAi jnojj -m^ jad sasuadxg pnoj, •mx J9d fan j jad iq%vj& P U13 'saqddng 'joqi^i 'saSj^qo paxij •HO jo jajj^g jad sjnojj - Ai5[ e;uaQ — jajj-eajad 00"l«I ar M w^ jno H -Ai X J8Cl easuadxg j^oj, s;uag — jajjisgjad s;uao 9^, ianj q^M jnojj mjj iad sasuadxg i^ox s^uaQ — jajj^g jad s;uaQ 05 janj q^iAi. jnojj '*^ jad easuadxg l^oj, •s^uaQ — jajj-Bg J9d 00'T$^ Bjno H ■aij jad jan^j s^uaQ — jnojj -m ^ jad ja;i3A\ pu-B ' saqddng 'joq^^ 'saSj'Bq^ P^xi^j I!0 J° P-wbH jad sjnojj avjj •M.51 — p^U jo uj tjJ ci ec OMNrHOO)OON rjJ ec e<» c0r- (OCti- i00«S»O TjneoetO«0»0"3»0 OOfflNrtNOOiHOO'tWOJM t^r~t^ooooc^coeo-*ti>.i>.oo W : :ffl :q : i o « o 6 d d 6 6 §£8888888§8§; ^ H ^ Hr -,eoiot^.o5525o< HHNUJOO FINANCE AND ECONOMICS — COST OF POWER 719 NNOOOINNONNiOifMNH ITS Tji Tl* CO CO CQ C^ «H «-< rH r4 i-l r-i ,-i i^WMhOO Lft ■<# CO CO -ie>5NO> OOOONO^NiOWINHHOOlX ■QOa>0>OOX«OcD>Q lTtt(Mt~»>CO»O HNU50CDC >(MrJHC» ! -IOOC3COt-t--g>00 *3< CO t-coi-n>CMi-iOJCj»oooot-t^«o ■**iT*-£^.l>.0000-iiot^g?ooooooior^ t^ «> ic »o T*i ^i co csi -»ot— nooin cd ui ■*< t*" eo co csi c4 eq i-< t-4 )rt«5iO>0©i0O(Nffi00 )0>MO(Ort00010>t^ a>tot--oio^-«*00!ONlOrH O5»ft00t^00CT>C.»ft-*-Oi-iCO MOOrHlONOOrtOONN OJlO.-ir^OOr^.-ia>0005CO ib "** t»J eo eo e-J — » t-4 ^h NiDOOCCNOlOOOSNCt NClOm-^^OOiO-HtO CN| C<1 SO lO Oi SO i— H~-lOlO-«l« »o "* co eo .t~<© co 10 ^h Tti ■**< im a> « VCNNNHHI e b = or C a ~ z I T C I £ c c £ a. t a I p i 1 1 I C C £ c '-3 b c '5 c a X = c a — c £ i O Oi O 7) o o o o o .' L^ ! t< : :M :o • • St2SSSSSS< lrtlSlSN< 2 « a a o, > o J2 •tt 9. S°o fl S£- s 3 £&^S cv • F fc< < ts c § £ o »>>« 3=3-3.2 1.1^ S" git eg a-* ft § 3 o 8 3-S«o.S o-* 3 > -« bd-53'O'r » a °«ft?pgs O « 03 uc co ft^pi5. 2 2 £ _,pq os o 5 ^ 3bS. - KM* SfttJ sg-,2 ^-^ g S3 03 « »T3— 3 ■s ft'2 & 5^3 « ^Joo' 3 S ^ 720 STEAM POWER PLANT ENGINEERING TABLE 105. COST OF POWER. Examples of Isolated Station Practice. Rated capacity kilowatts Yearly capacity kilowatt hours Actual load kilowatt hours Yearly load factor per cent Curve load factor per cent Large Office Building. 500 4,380,000 670,000 17 17 Small Office Building. 50 438,000 40,820 9.1 24.7 Manufactur- ing Plant, Electro- plating. 500 4,380,000 3,500,000 80 80 Operating Charges, per Year. Labor Coal and ashes $6,050.00 6,342.00 642.00 168.00 395.00 69.00 $1,400.00 960.00 75.00 90.00 41.00 182.00 $12,300.00 9,100.00 Water Oil and waste 210 00 Lamps Repairs and renewals 50.00 1,008.00 Total $13,666.00 $2,748.00 $22,668.00 Fixed Charges, per Year. Interest (5 per cent) $3,500.00 4,200.00 350.00 1,050.00 900.00 $325.00 628.00 30.00 90.00 $4,500.00 Depreciation (6 per cent) 5,400.00 450.00 1,350.00 Insurance (£ per cent) Taxes (11 per cent) Total $10,000.00 $1,073.00 $11,700.00 Cost per Kilowatt Hour, Cents. Operating charges Fixed charges Total cost ,65 .33 .98 FINANCE AND ECONOMICS — COST OF POWER 721 00 *j o o — i a) CO o o o © O "< © W. CO © © © 4% © g 2S CO o •3 ^ 5 S CO • CN •V © 3 A CN ^ O iO t^ B CN rH 5 Office 12 40X 100 -20 H.P. 6 floors Simple 1 40 .65 240 «5 yOb 3 § co rH PQ ^ B rM 0) p. 1 OS : 45 Si £ c b is 'E PC 1 — M '1 1 1 rH c b C ■d c s 1 > a a jl 'Z t a a 1 i a s i = c 1 ft i p a £ E 1 a »; c 1 <£" R 1 > J 1 I • t c , c TZ s l( — a < J a 1 o I -° 2: «*h c a) a a ~ >> c H PC 1 a a ! PC 1, c 1 + B X g B > 1 si ] c ?1 u 9 ) ! | fe CM O o o t^ CO X a CO • © r^ >o •* 1? CO • © Tfl CO • rH (M t OS >0 • © © 00 ■* ^ m CO «3 CO o O CO O CO c o> • CN t^ © • t^ CO o o 05 • CO CO I> rH rH Tt* • © © 00 _i "* ■<* CN I © o r-f »s 99 fe ■* CN t^ co •hj io ^ ur. ■* CO NOOO N>f rH O) o •<* Oi t^ CN CO CN CO CO CO CO • o CN (N CN "3 rH TtH If} O CO rH 1— 1 •*" 00 00 co a H ^ (O g ^ ©8 ee 53 °° rf< CO CO O lOlO o rH © CN o o CXI CN CO O t> • © CO rH • rH 0> r-< "■# CO CO • CO © © 00 rH © OS K> CN CN C 33 0 NN w o3 ^ CN as -4-a" C3 22 © CO CO 00 C75 O t* OS CO CN ^f O rH O 5 lO CD lO l> CO CN CO CO •<* CO rH CO If} • w "g I^ CO I> ^ CO "* CO o OS OS 05 O CN OS O co 1> co" OS if} O CO* r^ CN OS CO »o S M h iO w -(j" s «* a> OS ft! 53 ^ la O CO t* o o o lH co tj< t^ Tt< r-- o «4H t^ l^ CO O f} CN CO CO "# i> Tt< o • o o o S t N (O w 03 C*? 1-i rH CN o as Q 53 © 00 "* rH O CO o c CO ^ O* NO) "O g © iO IC CO CN 00 "* t- CO O CO CO lO rH • 05 ^ I> > X5 8 1 "8 I 73 a a K .1 c K 3 o 1 J li ■3 X 3 o cS "^ 1 1 — light, kilowatt — power, kilowat — total kilowatt I 5 c c _ p. 8 3 c3 p. c ips. airs rest 3 3 3 ft ° -d jp c ^, a C! ft ft ft .. 03 3 E 'f 1 C a — C c r^ P 5 1 4; J 3 = >oc 1 1 3 P Ph i 722 STEAM POWER PLANT ENGINEERING CQ o Q W ° i 1— < 1-1 o m * P* C H CQ c t^ £ t © _ CN C LO X c W5 "5 »-* d O • «J «0 t> -H -£ X O0» u x : "* 2 : m co ffl CO JO CO ©Ph'scooooo ■ o iO • ^ 1-4 3 n 8 B -< srsg X o OS s §^- CO « £ 2 .3 o «o IO CN c o |°. ** "3 O • TO CN CO O 00 - c f 13 o cc O • CN CO «fl r-i •£• ff CD <© o M P4 • a d § :| s t-a -* 2 a p O CU 3 « m O lO Q. L- 10 v S S 1 S M -d £ co be "33 5f £ c S * = O f f. c 5 .5 .11 7 c c - p £ .S fc O 3 u o f o * o I E Q 1 - E be o H 3 1 - 3 a, o cS! ^ -d to o 3 ca PC > E- •? 2 * 2 c E- > < ^ > E- '* PC •2 3 c co o CO o c '■f r)< CO • CO ^r co CD g O -O OJ » ^_ «3 rt ^ (N rH o> OS o> N CO t^ 1-( PH x x • I> «o I> •o >o O * CO •o" X* s'g " ^ «» 93 CO t_ O iO b« CO •* N O ■* ON(D h SOWH lOMrtO CO O iO CO • PNMH l-H tH ■* 00 © ■* I-H *«,H i-H id" §8 ^ ^. 9& pH O flOO CO CD CO IO I> l> r- os »-i tj< - CN X CO X o o o ~ ■ • CO CO X Tj< CO • i-i CO o * a m of *g « «o o O IO CO CN CO IO CO IO (M CO CO CN IO I> - CN to CN l-H i-H «c o O H t-. • m - CO CN E t^ CO CO • CO .• CO cog - iO 2 O O CC o •■# O iO Tt O l> CO ■"* S 9B0C a£ co »o ■* co >o o ■«# ■«* O CO CN ■ CO CO Hi o CT O O X CN Sci H " H " co" «o" ee t.' w ° ° c O X CN CC CN O O O 05 o O Ol N iO CD OS ■* IO X cs o CO ii O • ^ p ^ CO X 05 ©_ OOOM d ^= C5 IO S o ~ ~ CN m -. O ^ i-l T3 3 3 -OCOCOOCOOCO "^(OO^CCON ^00iOMNH«n x o o o io os n< •o e o»ioh • 3 CC nouooioo O — co »o CN CN (NON n H" ^h x «o CO 2 ^ -i CN ^ w -3 e# ooOcO-cUCNXCOiOC r- • iO O CN -H OCNCOCOI> ■* © C ■c* •O Nh . ^ g CO OS ■* c- CN • CO i-h CO iO : « d N 'o « s» co" -t-3 X O year ,280 ,995 164 CO X >0 CN o c o o -* o »o CO CN ■<* t^ cn no n© n • i-l K> CN N iiH o"5 "O _. CN >-" 9& "■ •O CO CN _ e# us cn x ^ K hours tt hour hours . , u cc 1 a> a £ °3 +» *J CO 3 1 o (T^ C ilowa kilovs lowat hour, med , lowat en TO o .1 - 3 '> > a 1 a^ d .§ c _ — light, k — power, — total ki r kilowatt coal consu coal per ki CP > 1 •3 -d Z a a cc CO co -^ CC 8 • S _, t_, CO t. and ips. airs rest d d a & ° T3 CO o- c p. a d. .. co a a E" j: ! 1 — C 5 p - C 3 c 3 c 1 1 1 Pi FINANCE AND ECONOMICS — COST OF POWER 723 TABLE 107a. YEARLY OPERATING COSTS IN FOUR TYPICAL CENTRAL STATIONS, STATE OF MASSACHUSETTS Year ending June, 1909. Type of Prime Mover Rated station capacity, kw.. Output, millions of kw. hrs.. . Yearly load factor, per cent. . Total station operating force Cost of fuel, dollars per ton . . Coal per kw. hr Salem Elec- tric Light Co. 6 Engines 2500 3.106 14.2 14 4.51 3.3 Fitchburg Gas & Electric Co. 3 Engines 2000 4.006 22.9 12 4.52 3.28 Haverhill Electric Co. 2 Turbines 1 Engine 2300 3.721 18.5 13 3.97 3.27 Maiden Electric Co. 1 Turbine 3 Engines 4.715 14 3.78 3.02 Operating Costs, Cents per Kilowatt Hour. Coal 0.740 0.025 0.027 0.410 0.034 0.158 0.011 0.740 0.015 0.025 0.308 0.017 0.041 0.072 0.024 0.650 0.190 0.003 0.285 0.063 0.073 0.019 0.040 0.565 Oil and waste 0.020 Water 0.045 Wages 0.320 Station building repairs. .... Steam equipment repairs. . . . Electrical equipment repairs. Miscellaneous 0.023 0.072 0.14 0.21 Total 1.412 1.242 1.152 1.08 TABLE 107b. COST OF POWER, CENTS PER KW. HOUR. STEAM-ELECTRIC CENTRAL STATIONS. Year ending June 30, 1908. Fuel Oil and waste Water Wages Station repairs Steam repairs Electrical repairs Miscellaneous Total Cost of fuel per ton Output, millions kilowatt hours per year Capacity of station, thou- sands of H.P Bos- ton. .462 .008 .024 .192 .015 .042 .056 .023 .822 3.99 88.5 73.5 Worcester. .703 .027 .034 ,360 ,012 055 055 000 1.246 4.79 5.4 5.90 Lowell .710 .009 .008 .262 ,020 ,020 ,009 022 1.060 4.75 9.4 7.39 Fall River. .880 .032 .012 .538 .012 .037 .029 ,080 620 68 4.0 4.43 Mai- den. .635 .017 .032 .342 .035 .072 .014 .033 1.180 4.49 4.6 4.87 Cam- bridge. .690 .019 ,055 ,347 ,021 059 046 000 1.237 4.40 6.0 6.75 Lynn. .618 .012 .040 .296 .052 .147 .045 .000 1.210 3.60 8.7 8.2 724 STEAM POWER PLANT ENGINEERING TABLE 108. COST OF POWER (1907), FIRST NATIONAL BANK BUILDING, CHICAGO. Total weight of coal burned tons Total weight of water evaporated gallons Total electrical output kilowatt hours Water actually evaporated per pound of coal as fired 14,956 22,100,000 1,546,600 6.1 Electric-Light Plant. (Cost of Power only.) Pounds of coal per kilowatt hour Cost of coal per kilowatt hour cents Cost of labor per kilowatt hour cents Cost of supplies per kilowatt hour cents Total '. 6.57 0.88 0.78 0.11 1.77 All expenses of entire plant charged against switchboard. Pounds of coal per kilowatt hour Cost of coal per kilowatt hour cents Cost of labor per kilowatt hour cents Cost of supplies per kilowatt hour cents Total 19.34 2 23 1 59 1 18 5.00 Elevator Plant. Passengers handled Total car miles Cost of labor per car mile cents Cost of material per car mile cents Cost of power * per car mile cents Total 2,016,300 92,700 8.14 * Approximate. BIBLIOGRAPHY. COST OF POWER. * Electrical. Rented Power for Electric Railways, American Electrician 10: 329 July, 1898 Cost of Pumping Station in New York, Electrical World 43: 820 April 30, 1904 Cost of Electric Power in Comparison with Steam for Traction, Engineer (London) 90: 600 Dec. 14, 1900 Cost of Steam and Electricity, Engineering 74: 667 Nov. 21, 1902 Cost of Power, Engineering 76: 706 Nov. 20, 1903 Economy of Isolated Plant (I. D. Parsons), Engineering Magazine 22: 573 Jan., 1902 720 Feb., 1902 * See page 729 for bibliography 1908-1910. FINANCE AND ECONOMICS — COST OF POWER 725 Electrical — Continued. Cost of Energy in Electric Supply (A. D. Adams), Engineering Magazine 24 : Street Railway Review 12: Data on Electric Power Generation in Glasgow, Engineering Record Cost of Generating Electric Power (E. J. Fox), Engineering Record 49 : Relative Cost of Electric Power for Three Types of Plants (R. D. Mushon), Engineering Record 49: Cost of Electric Power for Street Railways (R. W. Conant), Power 18 : Street Railway Review 8: Cost of Power in New Orleans Railroad Company Power Plant, Street Railway Journal 18: Cost of Power, Union Traction Company of Indiana, Street Railway Journal 18: Relative Costs of Steam and Polyphase Traction, Street Railway Journal 21 : Cost of Power at Newcastle-on-Tyne, Street Rail- way Journal 22 : Charges for Rented Power, Street Railway Review. . . 8 : Cost of Power for Electric Railways, Street Railway Review 8 : 9: 181 Nov., 1902 149 March, 1902 478 April 22, 1899 388 March 26, 1904 411 April 2, 1904 8 Oct., 1898 631 Sept., 1898 668 Nov. 2, 1901 827 Dec. 7, 1901 737 May 16, 1903 207 Aug. 8, 1903 236 April, 1898 43, 97, 186, 224, 340, 385, 461, 760, 886, 1898 35, 123, 185, 261, 319, 459, 529, 595, 749, 851 10: 11, 223, 399, 521, 735 11:123,416, 418, 1901 Cost of Niagara Power at Buffalo, Street Railway Review 8: 339 May, 1898 Statistics on the Cost of Power, Street Railway Review 12: 77 Oct., 1902 Analysis of Cost of Generation and Distribution of a Unit of Electricity (C. W. Rice), Western Electrician 22: 574 June 25, 1898 Cost of Electric Power at Lachine Rapids, Canada (W. L. Walbank), Western Electrician 23: 24 July 9, 1898 Suggestions Relative to Determining Cost of Elec- tric Supply (M. E. Turner), Western Electrician. . . 23: 143 Sept. 10, 1898 Cost of Power (C. S. Brown), Western Electrician ... 28: 127 Feb. 23, 1901 Cost of Power (C. Grey), Western Electrician 30: 211 March, 1902 Cost of Power (M. J. Eichorn), Western Electrician. . 31 : 69 Aug. 2, 1902 Graded Costs of Electrical Supply (M. E. Turner), Western Electrician 35: 204 Sept. 10, 1904 Some Notes on the Cost of Generating Electrical Energy (E. J. Fox), Engineer (London) 27: 219 Feb. 26, 1904 269 March 11, 1904 Effect of Load on the Cost of Power (E. M. Archi- bald), Engineer (United States) 42: 315 May 1, 1905 Cost of Power in Street Railway Service, Machinery 11:317 Feb., 1905 726 STEAM POWER PLANT ENGINEERING > Electrical — Continued. Cost of Electric Power at the Switchboard (C. H. Hile), Power 25: 662 Nov., 1905 Power Plant Economies (H. G. Stott), Engineer (United States) 43: 191 March 1, 1906 Power Costs (Charles E. Lucke), Electrical Review (New York) 50: 797 May 18, 1907 Systems of Charging for Electricity Supply (W. A. Toppin), Electrical Engineer (London) 39: 42 Jan. 11, 1907 Rates of Charge for Electricity and Their Effect on Cost (J. S. Codman), Proceedings of the American Institute of Electrical Engineers 26: 31 April, 1907 The Principles of Modern Rate-Making for Electric Light and Power, Electrical World 49: 1086 June 1, 1907 Methods of Computing Central Station Rates in Boston, Electrical World 49: 1090 June 1, 1907 The Present Tendency of Charging for Electricity (W. A. Toppin), Electrical Review (London) 60: 945 June 7, 1907 Electric Power Tariffs (C. S. Nesey-Brown), Cassier's Magazine 32: 304 Aug., 1907 The Sale of Electricity for Lighting Purposes (L. E. Bucknell), Electrical Engineer (London) 40: 370 Sept. 13, 1907 Rates and Systems of Charging (Jacques), Elec- trical Review (London) 61: 1074 Dec. 27, 1907 Gas. Cost of Pumping Station in New York, .Electrical World 43: 820 April 30, 1904 Comparative Cost of Power Generated by Steam Engine, Water Turbine, and Gas Engine, Engineer (London) 88: 320 Sept. 29, 1899 Comparative Cost of Generating Power by Steam Engine, Water Turbine, and Gas Engine (J. B. C. Kershaw), Engineering 70: 351 Sept. 14, 1900 390 Sept. 21, 1900 Cost of Gas Power for Central Station, Engineer .... 71 : 27 Jan. 4, 1901 Gas Power for Central Stations (J. R. Bibbins), Engineering Record 49: 11 Jan. 2, 1904 Street Railway Journal 1089 Dec. 26, 1903 Power 24: 100 Feb., 1904 Comparative Cost of Steam and Gas Plant, Engineering Record 49: 310 March 5, 1904 Is Gas Power More Economical than Water Power ? (H. C. T. Horace), Power 25: 599 Oct., 1905 Cost of Steam Power (Edit), American Electrician . . 10: 114 March, 1898 Cost of Steam Electrical Generating Plant (R. C. Carpenter), Electrical World 43: 1016 May 28, 1904 Economy of Power Installations (C. Weiss), Engineering 66: 59 July 8, 1898 Estimates for an Electric Light Plant in New York City, Engineering News 52: 583 Dec! 29, 1904 FINANCE AND ECONOMICS — COST OF POWER 727 Gas — Continued. Cost of a Power Station in Europe, Street Railway Journal 20:210 Aug. 9, 1902 Power House Cost, Louisville Electric Railway, Street Railway Review 9: 592 Sept., 1899 Miscellaneous. Improvements in Economy of the Steam Engine (W. F. Durand), American Electrician 11: 13 Jan., 1899 68 Feb., 1899 Cost of Power (Edit), Engineer (London) 96: 285 Sept. 4, 1901 Estimating the Cost of Power, Engineer (United States) 36:285 Dec. 1, 1899 Investigation of the Cost of Power (C. G. Gray), Engineer (United States) 39:43 Jan. 1, 1902 Efficient Use of Steam and Labor in Isolated Plants (P. R. Moses), Engineering Magazine. ... 16: 99 Oct., 1898 Cost Determination in Isolated Plants (P. R. Moses), Engineering Magazine 20: 1082 March, 1901 Cost of Pumping at a Colliery (R. V. Norris), Engineering News 49: 228 March 12, 1903 Economic Power Production (R. H. Thurston), Engineering Record 47: 35 Jan. 3, 1903 Cost of Power (Carpenter), Power 24: 425 July, 1904 Decreasing Costs in the Steam Plant (Edit), Engineer (United States) 42:412 June 15, 1905 The Economy of a Small-Sized Coal for the Power Plant (P. R. Moses), Engineering Magazine 28: 783 Feb., 1905 Analysis of Central Station Costs and Revenues (H. S. Knowlton), Engineering Magazine 29: 238 May, 1905 Cost of Operating Buildings, Engineering Record 48: 759 Dec. 19, 1903 Power Plant Supervision and Accounting (F. W. Ballard), Engineering Record 51 : 687 June 17, 1905 Power Plant Economy, Power 25: 602 Oct., 1905 Relative Efficiency and Desirability of Various Types of Engines (A. W. Richter), Street Railway Review 10: 162 March 15, 1900 Steam. Comparative Cost of Power Generated by Steam Engine, Water Turbine, and Gas Engine, Engineer (London) 88: 322 Cost of Electric Power in Comparison with Steam for Traction, Engineer (London) 90 : 600 Investigation of the Cost of Power, Engineer (United States) 39: 161 Cost of Steam Power per Horse Power Year (J. M. Whitman), Engineer (United States) 40: 741 Cost of Steam Raising (J. Holliday), Engineering ... 68: 739 Sept. 29, 1899 Dec. 14, 1900 March 1, 1902 Oct. 1, 1903 Dec. 8, 1899 Dec. 22, 1899 T28 STEAM POWER PLANT ENGINEERING Steam — Continued. Comparative Cost of Generating Power by Steam Engine, Water Turbine, and Gas Engine (J. B. C. Kershaw), Engineering 70: 351 Sept. 14, 1900 390 Sept. 21, 1900 Cost of Steam and Electricity, Engineering 74: 667 Nov. 21, 1902 Comparative Cost of Steam and Water Power (W. O. Webber), Engineering Magazine 15: 923 Sept., 1898 Steam Costs in an Industrial Combination (W. D. Ennis), Engineering Magazine 28: 86 Oct., 1904 Fuel Economy of Engines in Electric Railway Power Stations (Carpenter), Engineering News. . . 42:234 Oct. 12, 1899 Reduction in Cost of Steam Power from 1870 to 1897, Engineering Record 37: 12 Dec. 4, 1897 Economy in Use of Coal for Production of Power (I. N. Hollis), Engineering Record 46: 491 Nov. 22, 1902 Economy of Fuel in Electric Plants (Edit), Engineering Record 48: 233 Aug. 29, 1903 Economical Steam Making (Edit), Engineering Record 48: 385 Oct. 3, 1903 Cost of Fuel and Power in the South, Power 18:13 Nov., 1898 Economical Production of Steam with Special Reference to the Use of Cheap Fuel, Power 19: 19 June, 1899 Suggestions for Steam Economy (W. M. Kay), Engineer (United States) 42: 655 Oct. 2, 1905 Yearly Cost of One Steam Horse Power, Machinery . 9 : 374 March, 1903 Water. Cost of Water and Electric Power (G. E. Walsh), American Electrician 16: 331 July, 1904 Comparative Cost of Power Generated by Steam Engine, Water Turbine, and Gas Engine, Engineer (London) 88: 322 Sept. 29, 1898 Investigation of the Cost of Power, Engineer (United States) 39: 64 Jan. 15, 1902 Comparative Cost of Generating Power by Steam Engine, Water Turbine, and Gas Engine (J. B. C. Kershaw), Engineering 70: 351 Sept. 14, 1900 390 Sept. 21, 1900 Cost of Water Power in France, Engineering 76: 571 Oct. 23, 1903 Comparative Cost of Steam and Water Power (W. O. Webber), Engineering Magazine 15: 923 Sept., 1898 Cost of Hydraulic Transmission of Power (E. B. Ellington), Engineering Magazine 17: 233 May, 1899 Cost of Hydraulic Power in Switzerland, Engineer- ing Record 41 : 182 Feb. 24, 1900 Analysis of the Commerical Value of Water Power (A. F. Nagle), Engineering Record 46: 540 Dec. 6, 1902 Costs of Pumping Water, Power 20:12 Nov., 1900 FINANCE AND ECONOMICS — COST OF POWER 729 COST OF POWER. (1908-1910.) Approximate Cost of Gas Power (M. P. Cleghorn), Power and Engineer April 7, 1908 Central Station vs. Private Plants, Engineering Feb. 26, 1909 Comparative Cost of Power Production, Electrical Age 40:63 March, 1909 Electrical World 53: 792 April 1, 1909 Electrical Review and Western Electrician 55: 773 April 28, 1909 Cost of a Gas Engine and of a Combined Steam 1 Plant, Engineering Record 60: 272 Sept. 4, 1909 Cost of a Kilowatt-Hour (R. A. Day), Electrical World 54 : 853 Oct. 7, 1909 Cost of Power (H. G. Stott), Pro. Am. Inst. Elec. Engrs 28: 283 April, 1909 Cost of Power in a 3000 Kw. Turbine Plant, Elec- trical Review and Western Electrician 55: 62 Oct. 2, 1909 Cost of Power in a 1500 Kw. Central Station, Engi- neering News : 61 : 471 April 29„ 1909 Cost of Power in Small Plants (W. E. Snow), Engi- neering Magazine 33: 169 May, 1908 Cost of Power in Four Central Stations, Electrical World 55: 813 March 31, 1910 Cost of Power for Various Industries (C. T. Main), Jour. Assoc. Eng. Soc 44: 151 March, 1910 Engineering Record 60: 711 Dec. 25, 1909 Cost of Power in Varying Units (W. O. Webber), Engineering Magazine 35: 562 July, 1908 Cost Systems and Time Keeping, Columbus, O., Muni- cipal Electric Lighting Plant, Engineering News Dec. 3, 1908 Electric Power Costs in Small Stations, Engineering Record 59: 30 Jan. 9, 1909 First Cost of Plant and Cost of Generating and Distributing Electricity for Lights, Brooklyn Edi- son Co., Engineering Contractor 33: 393 April 6, 1910 Isolated Power Plant Costs and Their Relation to Central Station Service(W.F.~Lyod),ElectricalWorld 53:323 Feb. 4, 1909 Isolated Station Records and Accounting Power .... April 28, 1908 Operating Costs of Large Units, Power and Engi- neer 31: 981 May 31, 1910 Power Costs for Factories, Engineering Record 60: 604 Nov. 27, 1909 Power Plant Waste (P. R. Moses), Cassier's Maga- { 36: 497 Oct., 1909 zine . . . ( 37: 12 Nov., 1909 Relation of Load Factor to Power Costs, Jour. Wes. Soc. Engr 14: 241 April, 1909 Engineering Record 59: 702 June 5, 1909 Representative Data from Electric Power Plant Operation (H. S. Knowlton), Engineering Magazine 36 : 833 Feb., 1909 Systems of Charging for Electrical Energy (W. T. Ryan), Engineering Magazine April, 1909 The Small Station and its Economical Operation, Western Electrician 43: 10 July 4, 1908 The Valuation of Steam Power Plants (C. T. Main), Electrical Age 39: 228 Oct., 1908 Useful Figures from Practical Power Plant Opera- tion, Electrical World 54: 781 Sept. 30, 1909 Working Results from a Gas-Electric Power Plant (J. R. Bibbins), Pro. Am. Inst. Elec. Engrs 27: 1123 July 1, 1909 CHAPTER XVIII. TESTING AND MEASURING APPARATUS. 403. General. — The importance of maintaining a system of records has been referred to in paragraph 388. The various items which may be recorded and the instruments and appliances used in this con- nection are outlined in the accompanying chart. In large stations a full complement of indicating, recording, and integrating instruments may prove to be a good investment if intelligently and closely studied by the operating engineer with a view to locating and eliminating unnecessary losses. The instruments should be inspected and cali- brated at intervals, since many of them are delicately constructed and are apt to become inaccurate after a few months' service. Steam gauges, thermometers, and pyrometers, and particularly water meters are subject to appreciable error after considerable use. Voltmeters, ammeters, and other switchboard instruments are easily deranged, espe- cially when subjected to continuous vibration or to high temperature. 404. Weighing the Fuel. — In most small plants the delivery tickets of the coal dealer are depended upon for the weight of coal used, no attempt being made to determine its evaporative value, and the economy of the plant is judged by the size of the coal bill. In such cases a considerable saving can be effected by keeping a daily record covering at least the coal and water consumption. The coal can be conveniently weighed on ordinary platform scales. In a num- ber of large stations the weight of coal is determined by suspended weighing hoppers, which may be stationary, as in Fig. 109, or mounted on a traveling truck, as in Fig. 110. The scales of such devices are made indicating, autographic, integrating, or a combination of the three, the latter costing but little more than the simple indicating or recording devices. 405. Measurement of Water. — The most accurate means of measur- ing water is by the use of two or more tanks resting upon scales, arranged to be filled and emptied alternately. This method, however, involves considerably more time than is ordinarily at the disposal of the fireman or engineer. The usual practice is to place a hot-water meter on the pressure side of the feed pump, with provision for calibration without shutting off the feed supply to the boilers. 730 TESTING AND MEASURING APPARATUS 731 TESTING AND MEASURING APPARATUS STEAM PLANT. Weights , Pressures. Tempera- tures Power. Fuel. Water .(Water meters Steam High Low. Flue gas analysis Moisture. Fuel analysis Platform scales, indicating and autographic. Suspension hoppers, indicating and auto- graphic. Platform scales and tanks. Piston. .. .") Rotary . . . ^Integrating. Disk J Venturi, indicating and autographic. Weirs. 'Weighing condensed steam. Steam meters I St * John,s > autographic, steam meters. • | Burnham ^ indicating. ( Bourdon gauge, indicating and autographic, t Manometers, mercurial, indicating. ''Manometers — mercurial, indicating, and autographic. Manometers — water, indicating, and auto- graphic. .Diaphragms, indicating and autographic, f Mercurial thermometers, indicating. Up to 800 deg. F. 4 Expansion thermometers, indicating and [ autographic. Expansion thermometers, indicating and autographic. Resistance thermometers, indicating and autographic. Thermo-electric thermometers, indicating and autographic. 'Optical pyrometer, indicating and auto- graphic. ^Platinum or clay ball pyrometer. /T _j.-- „ + _j (Indicators, hand manipulated. f inaicaiea j Indicators, continuous autographic. {Rope brake. Prony brake Absorption dynamometers. Electric generator. TOrsat apparatus. J Arndt's econometer, indicating. ] Sarco and Ados recorder, autographic. LUehling gas composimeter, autographic. fin air Hygrometer, indicating and autographic. 1 1n steam Calorimeters. . - (Separating. ■ 800 to 2500 deg. F. < Over 2500 deg. F. - Coal calorimeters Gas calorimeter. "• /Throttling. fMahler bomb. J Carpenter. 1 Thompson. IParr. . .Junker. ELECTRICAL PLANT. Voltage Voltmeters, A.C. and D.C., indicating and autographic. Current Ammeters, A.C. and D.C., indicating and autographic. Output Wattmeters, A.C. and D.C., integrating and autographic. Power factor .Power factor meters, A.C. only, indicating and autographic. Frequency Frequency meter, A.C. only, indicating. Synchronism . .Synchronizers, A.C. only, indicating. 732 STEAM POWER PLANT ENGINEERING There are several types of meters in common use. Fig. 419 illus- trates the piston type, in which reciprocating pistons are displaced by a definite volume of water; Fig. 305, the rotary type, depending upon Fig. 419. A Typical Piston Water Meter. the displacement of rotating impellers; Fig. 420, the disk type, in which impellers are given a combined rotating and tilting motion. When periodically calibrated, water meters give satisfactory results. When graduated to read in pounds the accuracy de- pends upon the temper- ature range of the water; thus the density of water at 62 degrees F. is 62.36 pounds per cubic foot and at 212 degrees it is 59.76, a range of 2.6 pounds per cubic foot. Hence a meter calibrated to read correctly at 62 degrees F. will have an error of about 4.2 per cent if used to measure water at 212 degrees. The average range in feed tern- Fxg.420. A Typical Disk Water Meter. perature fc se l d om greater than 40 degrees, and if the meter is calibrated for the mean tem- perature the error is somewhat less than one per cent. The Venturi meter, Fig. 421, is frequently employed for measuring large volumes of water, as in city waterworks, and in connection with condensing plants.* It amounts practically to a constriction in the diameter of the pipe, is readily installed, and the total absence of * Tests on a Venturi Meter for Boiler Feed. Proc. A.S.M.E., Mid.-Oct ., 1909, p. 1065. TESTING AND MEASURING APPARATUS 733 working parts is a great advantage. The meter is supplied with either indicating or autographic manometer. With water at constant temperature the error in the readings should not exceed one per cent. The pitometer is a simple adaptation of the well-known pitot tube, and is used for measuring the flow of water through pipes where it is IPES TO MANOMETER =U=4=i Fig. 421. Principles of the Venturi Meter. n^?n MANOMETER impracticable to insert a meter. It is only necessary to drill a small hole in the pipe for the introduction of the tube. The volume flowing may be calculated from the readings of the manometer or may be autographically recorded. In measuring large volumes of water flowing in open channels the measurements are made by weirs of suitable proportions or by current meters. Water Measurement, General: Eng. Rec, Feb. 15, 1902. Water Meters: Trans. A.S.M.E., 18-134, 14-676, 5-63; Engng. News, Jan. 3, 1907, March 9, 1905, June 16, 1904, p. 569; Eng. Rec, Nov., 1903; Stevens Ind., Jan., 1901; Jour. New Eng. Waterworks Assn., June, 1907; Eleen., Lond. May 8, 1908. Venturi Meter: Revue Technique, Feb. 10, 1905, Eng. News, Feb. 28, 1901; Prac. Engr., Feb. 15, 1907; Pro. A.S.M.E., Nov., 1906; Trans. A.S.C.E., Nov., 1907, 57-531 (1906) ; Engng., Feb. 22, 1907, p. 236. Pitometer: Jour. Franklin Inst., Dec, 1907, p. 425; Tech. Quar., June, 1907; Jour. New Eng. Waterworks Assn., June, 1906; Trans. A.S.M.E., 25-184; Sib. Jour. Engng., May, 1902; Jour. Assn. Eng. Soc, Aug., 1901; Eng. News, March 31, 1904, Dec. 21, 1905; Pro. Engrs. Soc. of West. Penn., Dec, 1906. Pitot Tube: Am. Mach., Aug. 9, 1906, p. 175; Trans. A.S.C.E., 47-6, 57-265, 25-184; Eng. News, March 31, 1904, p. 318, Dec 21, 1905, p. 660; Progressive Age, June, 1906, p. 63; Jour. Assn. Eng. Soc, Aug., 1901, p. 35; Sib. Jour., May, 1902; Cal. Jour, of Tech., May, 1905. Weirs: Engr. (Lond.), June 5, 1903, p. 562, Aug. 17, 1906; Am. Soc. Civ. Engrs., 44-160; Eng. Rec, July 13, 1901, p. 32. >OWER PLANT ENGINEERING Current Meters: Cal. Jour, of Tech., April, 1904; Eng. News, March 7, 1907, p. 263, Feb. 12, 1902; Eng. Rec, Dec. 19, 1903; Pro. Am. Soc. Civ. Engrs., Sept., 1901, Nov., 1901, Dec, 1901. Piezometers: Power, Aug., 1907, p. 569; Pro. Am. Soc. Civ. Engrs., 44-34, 49-112, 51-252; Eng. News, Sept. 13, 1906, p. 271; Power, Aug., 1907, p. 569. 406. Measurement of Steam. — In surface-condensing plants the weight of steam consumed by the engines is conveniently obtained by weighing the condensed steam; in practice, however, this method Fig. 422. St. John's Steam Meter. Fig. 422a. Principles of the Burnham Steam Meter. is adopted only when testing the plant, the feed-water measurement sufficing for general recording purposes. When steam is supplied to various points and the weights cannot be readily determined by condensing, steam meters are sometimes used. The St. John's, Fig. 422, is the best known in this country. This apparatus records the weight of steam passing through the seat of an TESTING AND MEASURING APPARATUS 735 automatically lifting valve which rises and falls as the demand for steam increases or diminishes. When the maximum fluctuation in steam pressure is less than 10 pounds per square inch and the moisture in the steam is practically constant, this apparatus is said to register within two per cent of the actual weight flowing. The Burnham steam meter,. Fig. 422a, a recently patented device, offers the advantages of low first cost and simplicity of installation. This apparatus is attached to the steam pipe in a manner similar to a simple hydrostatic lubricator and occupies about the same space. It is based on the principle of the pitot tube, and the weight of steam flowing per unit of time is read from a graduated scale. Half-inch pipe fittings are used in connecting up. In Europe the principles of the Venturi meter have been success- fully applied to the measurement of steam. jfGluckauf, Dec. 9, 1905.) Steam Meters: Proc. A.S.M.E., Mid.-Nov., 1909, p. 1239. DIFFERENTIAL DRAFT GAUGE Fig. 423. Different Forms of Manometer Pressure Gauges. SIMPLE O TUBE 406a. Pressure Gauges. — The Bourdon type of gauge, either autographic or indi- cating (Fig. 424), is the most familiar and satisfactory means of measuring pressures up to 1500 pounds per square inch or more, although diaphragm gauges are also used and both are employed as vacuum gauges. For the latter purpose, however, the mercurial vacuum gauge has the ad- vantage of greater accuracy and is not subject to derangement. Bourdon gauges should be frequently standardized by com- parison with a gauge of known accuracy, a mercury column, or a gauge tester. For measuring very low pressures, such as are found in boiler flues Fig. 424. Bourdon Pressure Gauge. 736 STEAM POWER PLANT ENGINEERING or gas mains, indicating or recording diaphragm gauges may be had, but some form of U tube manometer is generally employed, the design best adapted to the purpose depending upon the accuracy required. The simple U tube (Fig. 423) when filled with mercury may be used for pressures limited only by the inconvenience due to length of tubes, or, with water as the fluid, for pressures only a fraction of an ounce per square inch. Where greater accuracy is required than can be obtained with the simple U tube, some modification may be employed, such as the Eames draft gauge with one inclined leg which magnifies the reading several times. A form of sensitive gauge is sometimes used which depends upon the use of two fluids of different specific gravity, as oil and water. Pressure Gauges, General References: Mech. Engr., Aug. 17, 1907; Am. Elecn., July, 1901; Engng., Aug. 23, 1907; Elec. World, Feb. 2, 1907, p. 258; Power, March, 1905, p. 184. Recording Pressure Gauge: Trans. A.S.M.E., 11-225, 14-325; Elec. World, April 28, 1906, p. 886. Draft Gauges: Trans. A.S.M.E., 21-123; Engr. U.S., Feb. 15, 1907, p. 218; Mech. Engr., Oct. 27, 1906. 407. Measurement of Temperature. — For power-plant purposes mercurial thermometers are most convenient for measuring tempera- tures up to 400 degrees F., and are inexpensive. For higher tempera- ture, up to say 800 degrees F., they are also adapted, but must be made of special glass and the space above the mercury filled with nitrogen under pressure to prevent vaporization of the mercury. Such thermometers must be used intelligently, and should be standardized from time to time, since they are subject to considerable change. The Bureau of Standards at Washington, D.C., is prepared to furnish certificates for which a nominal charge is made. Fig. 425 shows a form of thermometer which is much used where a continuous autographic record is required. It depends for its oper- ation upon the pressure produced by a fluid, liquid or gaseous, contained in a small bulb and exposed to the temperature to be measured. The pressure is transmitted to the recording mechanism through a flexible capillary tube which may be of . considerable length. Such thermometers are suitable for feed water, flue gas, and tempera- tures not exceeding 1000 degrees F. Fig. 426 illustrates a form of electrical pyrometer employing thermo- couples which has come into wide use as a reliable means of measur- ing temperatures up to 2600 degrees F. The couples most frequently used are composed of platinum and platinum-rhodium, platinum and platinum-iridium, copper and copper-constantan, and copper and nickel, TESTING AND MEASURING APPARATUS 737 the first named being adapted to the higher ranges of temperature. The electro-motive force set up, when the thermo-j unction is heated, is pro- portional to the temperature and is measured by means of a sensi- W//////MM? HH/WWM1L Fig. 425. Bristol Recording Pyrometer. Fig. 426. Bristol Thermo-Electric Pyrometer. tive millivoltmeter which is usually graduated to read temperature directly. Thermo-couples may be made to give an autographic record by means of a thread recorder. 738 STEAM POWER PLANT ENGINEERING Fig. 427 shows the element of an electrical thermometer based upon the change in resistance of a platinum wire when subjected to change in temperature. The resistance, in terms of temperature, is measured by a Whipple indicator, a convenient and portable form of Wheat- Fig. 427. Element for Callendar Resistance Pyrometer. stone bridge, or may be autographically recorded by means of a Callen- dar recorder. Resistance thermometers of this type are very sensitive and accurate, not easily deranged, and are limited in range only by the fusing points of the platinum and the porcelain protecting sheath. For higher temperatures and for obtaining the temperatures of inclosed spaces above about 900 degrees F., such as boiler furnaces, annealing ovens, and kilns, various forms of optical and radiation pyrometers have been devised. In such devices no part of the instru- OIFFUSING GLASS FLAME GAUGE l^AMYL-ACETAT LAMP Fig. 428. Wanner Optical Pyrometer in Position for Standardizing. ment is exposed to the temperature to be measured and hence suffers no injury from this cause. Optical pyrometers are based upon the measurement of the brightness of the hot body by comparison with a standard. The Wanner optical pyrometer is shown in Fig. 428. TESTING AND MEASURING APPARATUS 739 After standardizing by comparison with an amyl-acetate lamp, it is only necessary to focus the instrument upon the source of heat to be measured and the temperature is read on the graduated scale. TABLE 109- TYPES OF THERMOMETERS IN GENERAL USE. Range in Degrees F. Principle of Operation. Type. for which they can be used. Expansion .... Those depending on the Gas — 400 to +2900 — 35 to +950 change in volume or Mercury, Jena glass, length of a body with and nitrogen temperature. Glass and petrol ether. — 325 to +100 Unequal expansion of to 950 metal rods. Transpiration and cosity. vis- Those depending on the flow of gases through The Uehling to 2900 capillary tubes or small apertures. Thermo-electric . . .... Those depending on the electro-motive force developed by the dif- ference in temperature of two similar thermo- electric junctions op- posed to one another. Galvanometric — 400 to +2900 Electric resistance. . . . .Those utilizing the in- Direct reading on indi- — 400 to +2200 crease in electric resist- cator or bridge and ance of a wire with galvanometer. temperature. Radiation , . . . Those depending on the Thermo-couple in focus of mirror. 300 to 4000 bodies. Bolometer — 400 to Sun Optical . . . . Those utilizing the Photometric compari- son. change in the bright- ness or in the wave Incandescent filament 1100 to Sun length of the light in telescope. emitted by an incan- Nicol with quartz plate descent body. and analyzer. Calori metric . . . .Those depending on the Platinum ball with water vessel. 32 to 3000 specific heat of a body- raised to a high tem- perature. Fusion . . . . Those depending on the Alloys of various fusi- bilities. 32 to 3350 unequal fusibility of various 'metals or earthenware blocks of varied composition. Radiation pyrometers depend upon the measurement of the heat radiated from the hot body. The Fery radiation pyrometer, Fig. 429, 740 STEAM POWER PLANT ENGINEERING is the best-known instrument of this type. When focused upon the source of heat a cone of rays of definite angle is reflected by means of the mirror upon a thermo-couple located in its focus. The electro- motive force set up is measured in terms of the temperature of the TO GALVANOMETER Fig. 429. Fery Radiation Pyrometer. source of heat by a milli volt meter. Neither the couple nor any part of the instrument is ever subjected to a temperature much above 150 degrees F. The indications are practically independent of the distance from the source of heat, and the range is without limit. Table 109 embodies in outline the principles and temperature ranges of the various types of thermometers in use. Temperature ranges verified by U. S. Bureau of Standards. Indicating and Recording Thermometers, Expansion Type: Sci. Am. Sup., Dec. 16, 1905; Trans. A.S.M.E., 22-143; Jour. Am. Chem. Soc, 16-396; Jour. Soc. Chem. Ind., 13-61; Philosoph. Mag., 50-251, 1900. Indicating and Recording Pyrometers, Thermo-Electric: Jour. West. Soc. Engrs., Sept., 1907; Cassier's Mag., Aug., 1905; Elec. Rec., Jan. 12, 1901; Elec. Chem. and Met., June, 1901. Indicating and Recording Pyrometers, Electric-Resistance: Engng., May, 1899; Jour. Chem. Soc, 1890, 1895; Jour. Iron and Steel Inst., 1892; Pro. Royal Inst., Vol. XVI, 1901; Bureau of Standards, 3-641, 1907; Electrician, March 17, 1905, p. 880. Indicating and Recording Pyrometers, Optical: Elecn., Lond., Aug. 17, 1906; Am. Mach., Vol. 28, 160-29; Sch. of Mines Quarterly, April, 1907; Bulletin No. 2, Bureau of Standards, Wash., 1905; Jour, de Phys., Serjt., 1904; Engng., Sept. 6, 1907, Oct. 18, 1907; Cal. Jour, of Tech., Aug., 1907; Bureau of Standards. Bulletin No. 2; Iron Age, 73-24. Miscellaneous References: Engng. Times, March, 1904; Engng., Feb. 17, 1903, March 6, 1904; Sci. Am. Sup., July 22, 1905; Min. Rept., Aug. 8, 1901; Iron Age, Feb. 7, 1907; Iron and Coal Tds. Rev., May 10, 1907; Am. Elecn., May, 1904 ; Physical Rev., 8-193; Roy. Soc. of Lond., 66-86, 1900. TESTING AND MEASURING APPARATUS 741 408. Power Measurements. — The indicated horse power of recip- rocating engines is usually obtained by means of the steam-engine indicator. There are several reliable types to be had, including the continuous indicator, which permits of several diagrams being taken successively on the same paper. Among other devices may be mentioned mean pressure indicators and those giving the horse power directly. The developed horse power is determined by some form of absorp- tion dynamometer. For description of such dynamometers see Appen- dix C, article XV, A.S.M.E. code for conducting steam-engine tests. Power Measurements: Trans. A.S.M.E., 13-531; Am. Mach., Vol. 30, No. 27, Vol. 31, No. 5; Mechanical Engr., Feb. 23, 1907; Engng., June 14, 1907, p. 768. Indicators, Continuous: Trans. A.S.M.E., 18-1020; Power, Jan., 1907, p. 26. Prony Brakes: Trans. A.S.M.E., 15-62; Am. Mach., July 27, 1905, p. 127; Eng. News, Vol. 44, p. 216. Water Absorption Dynamometers: Prac. Engr., Sept. 14, 1906; Trans. A.S.M.E., 11-958; Eng. News, Vol. 51, p. 475; Prac. Engr., Sept. 14, 1906, p. 326. Fig. 430. Orsat Apparatus for Flue Gas Analysis. 409. Flue-Gas Analysis. — The simplest device for the analysis of flue gases is the Orsat apparatus (Fig. 430). In this apparatus a 742" STEAM POWER PLANT ENGINEERING measured volume, representing an average sample of the gas, is forced successively through pipettes containing solutions of caustic potash, pyrogallic acid, and cuprous chloride in hydrochloric acid, respectively, thus removing the carbon dioxide, the oxygen, and the carbon monox- ide, the contraction of volume being measured in each case. Orsat Apparatus: Trans. A.S.M.E., 18-901 ; Steam Boilers, Peabody and Miller, Chap. II; Power, Aug., 1907, p. 532; Engr. U.S., Jan. 1, 1907, p. 71. For most practical purposes it is sufficient to determine the carbon dioxide. A simple and efficient device for continuously indicating the per cent oT C0 2 is Arndt's econometer, Fig. 431. This apparatus rM±M Fig. 431. Arndt's Econometer. is a gas-weighing balance consisting essentially of a sensitive beam from one end of which is suspended a glass globe, closed at the top and open at the bottom, of about one pint capacity, and from the other end a compensating rod and scale pan. When not in operation the globe is filled with air and the scale pan and globe are in perfect balance, the indicator pointing to zero. When in operation the flue gases, thoroughly dried and filtered, are introduced in a continuous flow into the body of the hollow globe by means of a glass tube. The larger the per cent of C0 2 present in the contents of the globe the greater will be the deflection of the pointer, since C0 2 is about TESTING AND MEASURING APPARATUS 743 50 per cent heavier than atmospheric air. The scale is graduated to read from to 21 per cent C0 2 , and the results obtained check closely with those of the Orsat apparatus. Arndt's econometer is not portable, though it may be placed almost anywhere where it can be easily seen by the fireman. When there are a number of boilers, and it is not desired to have a separate instrument for each, the eco- nometer is connected with the breeching of each boiler by suitable piping, the gas from one boiler at a time being analyzed. For descriptive details see circular issued by Joseph Wickes, 106 Fulton Street, New York. MM///////7777n///// ///S//////A "" Bj^^^gpiU 2^|I- - -|n 3 HtJ V^ J □ I B^^Si 1 jS^^::?:^JH i !■ ^— 1[ V^ In .__ _____ 111 IB ^M l=n 1 B^^ : '--imB !_ \ V^S^SSK^ 1 j III E&H| X^ | J 9 r-' 1 In ' f_*_n( /^- 1 .E^!S 4 fi \mr1i trT\v*^^^^^ m 1 !~! \\w! ,,: r i l_JJT | mil V^; 1 a J 1 1 'EHi ^ IqK SI —HI ^^ ' ^^ .J is laig^Miiiit}^ 9 3 p m A TYPICAL CENTRAL STATION 777 778 STEAM POWER PLANT ENGINEERING A TYPICAL CENTRAL STATION 779 and carried to an ash bin directly over the coal track. Illinois screen- ings furnish the greater part of the fuel. Provision is also made for outside storage. Boilers. — The boiler plant is divided into five sections, each section consisting of sixteen 500-horse-power B. & W. boilers arranged in bat- teries of eight and equipped with B. & W. chain grates. The settings are installed back to back, as illustrated in Fig. 440. Each boiler has two 42-inch steam drums, approximately 5000 square feet of heating surface, and about 1000 square feet of superheating surface. Steam is generated at a pressure of 200 pounds per square inch with super- heat of 150 degrees F. The ratio of water-heating surface to grate surface is approximately 55 to 1, and the ratio of the total heating surface to grate surface is about 66 to 1. When burning Illinois screen- ings, an average depth of 7 inches is maintained on the grate, with speed of grate of 5 inches per minute. The grates are driven by Krehbiel oscillating engines, two engines being provided for each section but only one being in use at a time. The boilers are supported by reen- forced girders of the main building structure. A gallery is placed in front of the settings, 8 feet above the floor, to facilitate cleaning of tubes. Galleries are also placed between the batteries and on top of them. Spaces of 5 feet are provided between the sides and rears of the batteries, and 18 feet 8 inches in front. The furnaces are similar to the one illustrated in Fig. 62. The outside of the setting is finished with red pressed brick. Each drum is fitted with a four-inch pop safety valve. The blow-off main is 4 inches in diameter and discharges into the river. There are four blow-off connections to each boiler, each being provided with a blow-off cock and an angle valve; three of the connections are fitted to the mud drum and the other to the super- heater drain. Chimneys. — One stack is provided for each section of 16 boilers. The shaft is supported by the steel work of the boiler setting, as shown in Fig. 442, an arrangement which commends itself where space is limited and real estate values are high. The stacks for all units are 259 feet 6 inches in height above the grate surface, and are 18 feet in internal diameter. The lining is of radial fire brick and varies from 4 inches to 13 inches in thickness. The steel sections are 5 feet high and vary in thickness from f inch to \ inch. . There are two flues, one 32 feet long and the other 63 feet, which enter the stack on opposite sides. Turbines. — The prime movers are vertical five-stage Curtis turbines with base condenser and are rated at 12^000 kilowatts each. The normal speed is 750 r.p.m. The average steam consumption, including all 80 STEAM POWER PLANT ENGINEERING ~^fON mooji Satiij" A TYPICAL CENTRAL STATION 781 auxiliaries is approximately 15 pounds per kilowatt hour, corresponding to a coal consumption of 3 pounds per kilowatt hour (Illinois screenings, 10,400 B.T.U. per pound). Special tests have shown as low as 12.8 pounds per kilowatt hour, initial pressure 200 pounds gauge, 150 de- grees superheat, absolute back pressure \ inch of mercury. Each pair of units has a pair of duplicate pumps, an accumulator and a storage tank for supplying oil, the step-bearing pressure being maintained at 750 pounds per square inch. When the accumulator falls below a certain point a motor-driven pump is automatically started. Generators. — The generators are 2300- volt, 25-cycle, three-phase General Electric alternators mounted over the vertical shaft as illus- trated in Fig. 185. Exciting current is furnished by 2 50-kilowatt motor-driven generators. 2 75-kilowatt motor-driven generators. 2 1 50-kilowatt motor-driven generators. 2 75-kilowatt steam-driven generators. 2 1 50-kilowatt steam-driven generators. Part are held in reserve, though no particular units are maintained for the purpose. The high-tension cables lead -from the generator through an underground tunnel to the switch house, located about 50 feet west of the main building. The oil switches, wattmeters and other instruments are located on the first floor, while the bus-bars and other high-tension connections are in the basement. The station switch- board or operating gallery in the main building is equipped with only such devices as are necessary for the control of the machines, all other instruments being located in the switch house. From the switch house the high-tension current is conducted through oil switches to the various substations, where it is converted to direct current by rotary converters, or transformed from 25 to 60 cycles by motor generator sets. The Twenty-second Street substation is located at the north end of the property (Fig. 438). In this substation are installed two motor generator sets and one rotary converter, the latter supplying direct current to the neighboring district and to the main station. Boiler and Turbine Piping. — Immediately below each boiler section is an apartment called the " header room," where the steam pipes from the various boilers join the main header, which increases in size from 6 inches at the most remote boiler to 10 inches at the middle boiler, and finally to 14 inches where it leaves the nearest boiler and passes to the turbines. The pipes are of wrought iron, with welded flanges, and are packed with copper gaskets. The feeder from each boiler is 6 inches in diameter. An angle stop valve and a check valve are placed at the 782 STEAM POWER PLANT ENGINEERING boiler nozzle and a globe valve at the header. A motor-operated throttle valve and strainer are placed at the turbine, and a hydraulically oper ated valve, controlled from the operating gallery, is located in the header room. The main header is not anchored at any point, the entire weight being carried by roller supports. The only drain in the header is a f-inch bleeder on the turbine side of the hydraulically operated valve. The bleeder is connected to a trap which discharges into either boiler or superheater blow-off main. All branches or feeders are drained and discharged into the superheater blow-off. Condensers and Auxiliaries. — Each unit has its own condensing apparatus, feed-water heater, hot well and feed pumps. The condensers are of the Worthington " base " type with 25,000 sq. ft. of cooling surface each, composed of 5900-6000 1-inch tubes 16 feet long. Cooling water is taken from the east slip through concrete tunnels and is dis- charged from the condenser into similar tunnels which empty into the west slip. (See Fig. 438.) The dry vacuum pumps are of the rotative type, with cylinders 26 X 24, r.p.m. 100-120, and are driven by a 75-horse-power Corliss engine. The circulating pumps are of the volute single-stage centrifugal type and are mounted on an extension of the main shaft of the engine driving the dry vacuum pump. They are rated at 22,500 gallons per minute each. The hot-well pumps are of the two-stage centrifugal type, driven by 20-horse-power direct-current motors. The feed pumps are of the Dean vertical single-cylinder pattern and are installed in duplicate for each unit. The steam cylinders are 24 inches in diameter, water cylinders 14 inches in diameter, and common stroke 24 inches. Feed water is drawn in by suction from the hot well at a temperature of about 100 degrees F. and is forced through closed heaters having 3000 sq. ft. of heating surface, and its temperature is raised to 180 degrees. The heater receives the steam exhausted from the steam-driven auxiliaries. From the heater this water is forced through a 5-inch feed main to the different boilers in the section. The branches from main to boiler drum are 3 inches in diameter and fitted with an angle stop valve, a regrinding check and a grate-regulating valve. There is a 5-inch auxiliary main which supplies cold water to the boiler in case the main header is shut down. A TYPICAL CENTRAL STATION 783 &2 a V//ty>M,Y/.)$ZZ. 1 s ' & S Ni jj/iix / \ 4\ /iK'^tr y aop pei^X; i; /||\ B*|uluoi\, ^^Tt || H)f 100 X [(212 - t) + 966 + 0.48 (T - 212)] (l = temperature of air in the boiler room, T = that of the flue gases). Loss due to moisture formed by the burning of hydrogen = per cent of hydrogen to combustible -f- 100 X 9 X [(212 - t) + 966 + 0.48 (T - 212)]. * Loss due to heat carried away in the dry chimney gases = weight of gas per pound of combustible X 0.24 X (T - t). CO f Loss due to incomplete combustion of carbon = per cent C in combustible , X — X x0 ; 150. C0 2 +CO Loss due to unconsumed hydrogen and hydrocarbons, to heating the moisture in the air, to radiation, and un- accounted for. (Some of these losses may be sepa- rately itemized if data are obtained from which they may be calculated.) Totals. Per Cent. * The weight of gas per pound of carbon burned may be calculated from the gas analyses as follows: 11 C0 2 +80+7 (CO + N) Dry gas per pound carbon = ■ . in which C0 2 , CO, O, and N are 3 (COo + CO) the percentages by volume of the several gases. As the sampling and analyses of the gases in the present state of the art are liable to considerable errors, the result of this calculation is usually only an approximate one. The heat balance itself is also only approximate for this reason, as well as for the fact that it is not possible to determine accurately the percentage of unburned hydrogen or hydrocarbons in the flue gases. (See Appendix XXXII.) The weight of dry gas per pound of combustible is found by multiplying the dry gas per pound of carbon by the percentage of carbon in the combustible and dividing by 100. f C0 2 and CO are respectively the percentage by volume of carbonic acid and carbonic oxide in the flue gases. The quantity 10,150 = No. heat units generated by burning to carbonic acid one pound of carbon contained in carbonic oxide. XXIII. Report of the Trial. — The data and results should be reported in the manner given in either one of the two following tables, omitting lines where the tests have not been made as elaborately as provided for in such tables. Additional lines may be added for data relating to the specific object of the test. The extra lines should be APPENDIX B 833 classified under the headings provided in the tables, and numbered as per preceding line, with sub letters a, b, etc. The Short Form of Report, Table No. 2, is recommended for commercial tests and as a convenient form of abridging the longer form for publication when saving of space is desirable. For elaborate trials, it is recommended that the full log of the trial be shown graphically, by means of a chart. (Appendix XXXVIII.) TABLE NO. 1. Data and Results of Evaporative Test, Arranged in accordance with the Complete Form advised by the Boiler Test Com- mittee of the American Society of Mechanical Engineers. Code of 1899. Made by of boiler at ■ to determine Principal conditions governing the trial Kind of fuel * Kind of furnace State of the weather . Method of starting and stopping the test ("standard" or "alternate," Art. X and XI, Code) 1. Date of trial . 2. Duration of trial hours. Dimensions and Proportions. A complete description of the boiler, and drawings of the same if of unusual type, should be given on an annexed sheet. (See Appendix X.) 3. Grate surface width length area square feet 4. Height of furnace inches. 5. Approximate width of air spaces in grate inch. 6. Proportion of air space to whole grate surface per cent. 7. Water-heating surface square feet. 8. Superheating surface " 9. Ratio of water-heating surface to grate surface — to 1. 10. Ratio of minimum draft area to grate surface 1 to — . Average Pressures. 11. Steam pressure by gauge lb. per sq. in. 12. Force of draft between damper and boiler in. of water. 13. Force of draft in furnace " " 14. Force of draft or blast in ash pit " " * The items printed in italics correspond to the items in the " Short Form of Code." 834 STEAM POWER PLANT ENGINEERING Average Temperatures. 15. Of external air degrees. 16. Of fire room " 17. Of steam " 18. Of feed water entering heater " 19. Of feed water entering economizer " 20. Of feed water entering boiler " 21. Of escaping gases from boiler " 22. Of escaping gases from economizer " Fuel. 23. Size and condition 24. Weight of wood used in lighting fire pounds. 25. Weight of coal as fired * , " 26. Percentage of moisture in coal f per cent. 27. Total weight of dry coal consumed pounds. 28. Total ash and refuse 29. Quality of ash and refuse 30. Total combustible consumed pounds. 31. Percentage of ash and refuse in dry coal per cent. Proximate Analysis of Coal. (App. XII.) Of Coal. Of Combustible. 32. Fixed carbon per cent. per cent. 33 Volatile matter 34. Moisture " 35. Ash " 100 per cent. 100 per cent. 36. Sulphur, separately determined Ultimate Analysis of Dry Coal. (Art. XVII, Code.) Of Coal. Of Combustible. 37. Carbon (C) per cent. per cent. 38. Hydrogen (H) 39. Oxygen (O) 40. Nitrogen (N) 41. Sulphur(S) 42. Ash " — 100 per cent. 100 per cent. 43. Moisture in sample of coal as received * Including equivalent of wood used in lighting the fire, not including unburned coal withdrawn from furnace at times of cleaning and at end of test. One pound of wood is taken to be equal to 0.4 pound of coal, or, in case greater accuracy is desired, as having a heat value equivalent to the evaporation of 6 pounds of water from and at 212 degrees per pound. (6 X 965.7 = 5794 B.T.U.) f This is the total moisture in the coal as found by drying it artificially, as described in Art. XV of Code. APPENDIX B 835 Analysis of Ash and Refuse. 44. Carbon per cent. 45. Earthy matter " Fuel per Hour. 46. Dry coal consumed per hour pounds. 47. Combustible consumed per hour . 48. Dry coal per square foot of grate surface per hour 49. Combustible per square foot of water-heating surface per hour. . . Calorific Value of Fuel. (Art. XVII, Code.) 50. Calorific value by oxygen calorimeter, per pound of dry coal B.T.U. 51. Calorific value by oxygen calorimeter, per pound of combustible 52. Calorific value by analysis, per pound of dry coal * 53. Calorific value by analysis, per pound of combustible Quality of Steam. (App. XV to XIX.) 54. Percentage of moisture in steam per cent. 55. Number of degrees of superheating degrees. 56. Quality of steam (dry steam = unity). (For exact determination of the factor of correction for quality of steam see Appendix XVIII) Water. (App. I, IV, VII, VIII.) 57. Total weight of water fed to boiler f pounds. 58. Equivalent water fed to boiler from and at 212 degrees 59. Water actually evaporated, corrected for quality of steam 60. Factor of evaporation J pounds. 61. Equivalent water evaporated into dry steam from and at 212 degrees. (Item 59 X Item 60.) " Water per Hour. 62. Water evaporated per hour, corrected for quality of steam 63. Equivalent evaporation per hour from and at 212 degrees 64. Equivalent evaporation per hour from and at 212 degrees per square foot of water-heating surface * See formula for calorific value under Article XVII of Code, f Corrected for inequality of water level and of steam pressure at beginning and end of test. t Factor of evaporation = ,__ _ in which H and h are respectively the total heat 965.7 in steam of the average observed pressure, and in water of the average observed temperature of the feed. 836 STEAM POWER PLANT ENGINEERING Horse Power. 65. Horse power developed. (34| pounds of water evaporated per hour into dry steam from and at 212 degrees equals one horse power.)* ... horse power. 66. Builders' rated horse power " 67. Percentage of builders' rated horse power developed per cent. Economic Results. 68. Water apparently evaporated per pound of coal under actual condi- tions. (Item 58 -s- Item 25.) 69. Equivalent evaporation from and at 212 degrees per pound of wet coal. (Item 61 -f- Item 25.) 70. Equivalent evaporation from and at 212 degrees per pound of dry coal. (Item 61 -j- Item 27.) 71. Equivalent evaporation from and at 212 degrees per pound of com- bustible. (Item 61 -f- Item 30.) (If the equivalent evaporation, Items 69, 70, and 71, is not cor- rected for the quality of steam, the fact should be stated.) pounds. Efficiency. (See Art. XXI, Code.) 72. Efficiency of the boiler; heat absorbed by the boiler per pound of com- bustible divided by the heat value of one pound of combustible f . . . per cent. 73. Efficiency of boiler, including the grate; heat absorbed by the boiler, per pound of dry coal fired, divided by the heat value of one pound of dry coal X " Cost of Evaporation. 74. Cost of coal per ton of pounds delivered in boiler room $ 75. Cost of fuel for evaporating 1000 pounds of water under observed conditions $ 76. Cost of fuel used for evaporating 1000 pounds of water from and at 212 degrees $ Smoke Observations, (App. XXXIV and XXXV.) 77. Percentage of smoke as observed per cent. 78. Weight of soot per hour obtained from smoke meter ounces. 79. Volume of soot per hour obtained from smoke meter cubic inches. * Held to be the equivalent of 30 pounds of water per hour evaporated from 100 degrees F. into dry steam at 70 pounds gauge pressure. (See Introduction to Code.) f In all cases where the word combustible is used, it means the coal without moisture and ash, but including all other constituents. It is the same as what is called in Europe "coal dry and free from ash." J The heat value of the coal is to be determined either by an oxygen calorimeter or by calculation from ultimate analysis. APPENDIX B 837 Methods of Firing. 80. Kind of firing (spreading, alternate, or coking) 81. Average thickness of fire 82. Average intervals between firings for each furnace during time when fires are in normal condition 83. Average interval between times of levelling or breaking up Analyses of the Dry Gases. 84. Carbon dioxide (C0 2 ) per cent. 85. Oxygen (O) 86. Carbon monoxide (CO) " 87. Hydrogen and hydrocarbons " 88. Nitrogen (by difference) (N) 100 per cent. TABLE NO. 2. Data and Results of Evaporative Test, Arranged in accordance with the Short Form advised by the Boiler Test Committee of the American Society of Mechanical Engineers. Code of 1899. Made by on boiler, at to determine Kind of fuel Kind of furnace Method of starting and stopping the test (" standard" or "alternate," Art. X and XI, Code) Grate surface square feet. Water-heating surface " Superheating surface " Total 1. Date of trial 2. Duration of trial hours. 3. Weight of coal as fired * pounds. 4. Percentage of moisture in coal * per cent. 5. Total weight of dry coal consumed pounds. 6. Total ash and refuse " 7. Percentage of ash and refuse in dry coal per cent. 8. Total weight of water fed to the boiler pounds. 9. Water actually evaporated, corrected for moisture or superheat in • steam " 10. Equivalent water evaporated into dry steam from and at 212 * See foot-notes of Complete Form. 838 STEAM POWER PLANT ENGINEERING Hourly Quantities. 11. Dry coal consumed per hour pounds. 12. Dry coal per hour per square foot of grate surface " 13. Water fed per hour " 14. Equivalent water evaporated per hour from and at 212 degrees corrected for quality of steam " 15. Equivalent water evaporated per hour per square foot of water- heating surface " Average Pressures, Temperatures, etc. 16. Average boiler pressure lb. per sq. in. 17. Average temperature of feed water degrees. 18. Average temperature of escaping gases " 19. Average force of draft between damper and boiler in. of water. 20. Percentage of moisture in steam, or number of degrees of super- heating Horse Power. 21. Horse power developed (Item 14 -f- 34^) horse power. 22. Builders' rated horse power " 23. Percentage of builders' rated horse power developed per cent. Economic Results. 24. Water apparently evaporated per pound of coal under actual conditions. (Item 8 -5- Item 3.) pounds. 25. Equivalent water actually evaporated from and at 212 degrees per pound of wet coal. (Item 9 -r- Item 3.) " 26. Equivalent evaporation from and at 212 degrees per pound of dry coal. (Item 9 -f- Item 5.) " 27. Equivalent evaporation from and at 212 degrees per pound of combustible. [Item 9 -*- (Item 5 — Item 6).] " (If Items 25, 26, and 27 are not corrected for quality of steam, the fact should be stated.) Efficiency. 28. Heating value of the coal per pound B.T.IL 29. Heating value of the combustible per pound " 30. Efficiency of boiler (based on combustible) per cent. 31. Efficiency of "boiler, including grate (based on coal) " Cost of Evaporation. 32. Cost of coal per ton delivered in boiler room $ 33. Cost of coal required for evaporation of 1000 pounds of water from and at 212 degrees $ APPENDIX B 839 LIST OF APPENDICES TO CODE * No. of Appendix. I. Relative Weights of Water and Fuel c. e. e. II. Object of the Test. (I, 1885 Code) j. c. h. III. General Observations. (II, 1885 Code) c. t. p. IV. Precautions to be Observed in Making a Boiler Test. (Ill, 1885 Code) c. e. e. V. Weighing the Coal. (IV, 1885 Code) j. c. h. VI. Weighing the Coal. (V, 1885 Code) c. t. p. VII. Weighing the Water. (VI, 1885 Code) j. c. h. VIII. Measuring the Feed Water. (VII, 1885 Code) c. t. p. IX. Keeping Time of Observations. (VIII, 1885 Code) j. c. h. X. Description of Boiler. (XXIII, 1885 Code) . . .c. e. e. XI. Determining the Moisture in Coal w. k. XII. Proximate Analyses of Coal . . w. k. XIII. Coal Calorimeter g. h. b. XIV. Comparative Calorimetric Tests of Coals w. k. XV. Determination of the Moisture in the Steam w. k. XVI. Correction for Radiation from Throttling Calorimeters g. h. b. XVII. Combined Calorimeter and Separator g. h. b. XVIII. Corrections for Quality of Steam c. e. e. XIX. The Quality of Superheated Steam g. h. b. XX. Efficiency of the Boiler w. k. XXI. Distribution of the Heating Value of the Fuel . . w. k. XXII. Observation Blanks. (Amendment to XXIV, 1885 Code) . . . .c. e. e. XXIII. Horse Power. (XXV, 1885 Code) j. c. h. XXIV. Steam Units. (XXVI, 1885 Code) c. e. e. XXV. Discrepancy between Commercial and Experimental Results. .. .c. e. e. XXVI. Recording Steam Gauge. (IX, 1885 Code) j. c. h. XXVII. Pyrometer. (XIII, 1885 Code) c. t. p. XXVIII. Draft Gauge. (XIV, 1885 Code) j. c. h. XXIX. Draft Gauge g. h. b. XXX. Draft Gauge w. k. XXXI. Sampling Flue Gases. (XVI, 1885 Code) j. c. h. XXXII. Computation of the Weight of Chimney Gases from the Analysis by Volume of Dry Gas w. k. XXXIII. The Orsat Apparatus for Analyzing Flue Gases g. h. b. XXXIV. Smoke Measurements g. h. b. XXXV. The Ringelmann Smoke Chart w. k. XXXVI. Starting and Stopping a Test w. k. XXXVII. Starting and Stopping a Test g. h. b. XXXVIII. Chart Showing Graphically the Log of a Trial g. h. b. XXXIX. Continuous Determinations of Carbonic Acid in Flue Gases. . . .g. h. b. XL.fMeasuring Radiation from Certain Types of Boilers .r. s. h. XLI.f Determination of the Moisture in Steam Flowing through a Horizontal Pipe d. s. j. * Only a few of the appendices are reprinted. t Contributed by members of the Society and accepted by the Committee for publication in the Appendix. 840 STEAM POWER PLANT ENGINEERING APPENDIX XX. Efficiency of the Boiler. The efficiency of the boiler, including the grate, or the efficiency based on coal, is the quotient arising from dividing the heat absorbed by the boiler by the heating value of the total amount of coal supplied to the boiler, including the coal which falls through the grate. It may be conveniently calculated by multiplying the number of pounds of water evaporated from and at 212 degrees F. into dry steam per pound of dry coal by 965.7 and dividing the product by the heating value in British thermal units of one pound of dry coal. The efficiency of the boiler, not including the grate, or the efficiency based on combustible, is the quotient arising from dividing the heat absorbed by the boiler by the heating value of the combustible burned. It may be calculated by multiplying the number of pounds of water evaporated from and at 212 degrees F. into dry steam per pound of combustible by 965.7 and dividing the product by the heating value in British thermal units of one pound of combustible; the term " com- bustible " being denned as coal dry and free from ash, or the coal sup- plied to the boiler less its moisture and the ash and unburned coal which falls through the grate or is otherwise withdrawn from the furnace. The efficiency of the boiler, not including the grate (or the efficiency based upon combustible), is a more accurate measure of comparison of different boilers than the efficiency including the grate (or the efficiency based upon coal), for the latter is subject to a number of variable con- ditions, such as size and character of the coal, air spaces between the grate bars, skill of the fireman in saving coal from falling through the grate, etc. It is, moreover, subject to errors of sampling the coal for drying and for analysis, which affect the result to a greater degree than they do the efficiency based upon combustible, for the reason that the heating value of one pound of combustible of any sample selected from a given lot, such as a car load, of coal is practically a constant quantity and is independent of the percentage of moisture and ash in the sample; while the sample itself, upon the heating value of which the efficiency based on coal is calculated, may differ in its percentage of moisture and ash from the average coal used in the boiler test. When the object of a boiler test is to determine its efficiency as an absorber of heat, or to compare it with other boilers, the efficiency based on combustible is the one which should be used; but when the object of the test is to determine the efficiency of the combination of the boiler, the furnace, and the grate, the efficiency based on coal must necessarily be used. APPENDIX B 841 It has been proposed that in reporting the efficiency of a boiler when the fuel used contains hydrogen, the efficiency should be considered to be the sum of the percentage of the heating value of the fuel which is utilized by the boiler in making steam and of the percentage of that heating value which is lost in the shape of latent heat in the moisture in the chimney gases, which moisture is formed by the burning of the hydrogen. This latent heat may amount to over three per cent of the total heating value of the fuel. The reason assigned for this proposal is that, since it is impossible for this heat to be utilized by the boiler because the gases are discharged at a temperature above 212 degrees F. it should not be charged against the boiler. The writer does not con- sider it advisable that this method of reporting the efficiency should be adopted (1) because it is opposed to the generally accepted definition of efficiency, which is the useful work received from an apparatus divided by the work (or heat value of the fuel) put into it; (2) because in order to calculate it it is necessary to know both the percentage of hydrogen in the coal and whether or not all of this hydrogen has been burned to H 2 0, the first requiring an analysis of the coal, which is not always obtainable, and the second an analysis of the gases for hydrogen, which cannot be obtained with any approximation to accuracy with our present methods of sampling and analyzing gases; and (3) because it is opposed to the almost universal custom in reporting boiler tests. It is true that the latent heat of the H 2 in the chimney gases cannot be utilized (unless an economizer which discharges its gases below 212 degrees is used), and it is not the fault of the boiler that it cannot be utilized. It may be considered the misfortune of the boiler, when tested with hydrogenous coal, similar to the misfortune under which an engine labors when it is tested while supplied with a condenser which gives a vacuum of less than 30 inches of mercury. The engine might give a higher efficiency with a vacuum of 30 inches than it would with one of 27 or 28 inches; but it is not customary to credit the engine with the efficiency which it loses on account of the imperfect vacuum. Since it is well understood that a boiler cannot show quite as high an efficiency (as commonly defined) when using bituminous coal high in hydrogen as when using anthracite nearly free from hydrogen, no harm is done, and much confusion is avoided, by reporting the efficiency as the percentage of the heating value of the coal which is actually utilized in making steam. The fact that bituminous coal is used is always stated in the report of a test made with that coal. If desired a state- ment may also be made in the " heat balance" of the approximate or esti- mated percentage of heat which is lost in the latent heat of the moisture in the chimney gases, together with the loss due to moisture in the coal. 842 STEAM POWER PLANT ENGINEERING APPENDIX XXV. Discrepancy between Commercial and Experimental Results. The final result sought by manufacturers, in initiating tests of steam or other machinery in actual use, is the value of the work done measured in dollars and cents. In some cases the broad question is raised as to the saving that may be accomplished by installing improved boilers, engines, or other machinery; but more generally it is desired to ascer- tain what can be done to produce saving with the apparatus already in place under the actual conditions that prevail at the particular location. In both these cases it is necessary to ascertain the average cost of the work done commercially previous to the test. Frequently, in fact generally, this important fact will not be ascertained by an elaborate trial, for the reason that everything will be put in order for the test, and all details of the trial be conducted so carefully that the losses due to average carelessness or want of skill in the past will be eliminated, the engineer making the test will not receive proper credit, and the owners on seeing the report may conclude that they are already doing very well, and perhaps continue old methods with fancied security. If the cost of the output of the factory for a given time were ascertained in terms of the coal burned during the same time, and compared with the corresponding cost for the time of the trial, the latter would fre- quently be found to be one-eighth to one-third less than the former, and it might not be possible to tell what had caused the difference; for instance, whether it was due to putting in order the machinery prior to the tests, to greater care exercised by the fireman under the spur of careful watching, or whether, as is usually claimed, the coal was different, etc., etc. The losses are generally due in the main to the carelessness of the firemen. It follows, therefore, that the cost of the power under average conditions must be obtained in some quiet way preliminarily. Frequently the comparison of the output of the factory with the coal burned will not be sufficiently accurate, and it will be necessary to devise some corresponding check which will not interfere with the regular routine of the establishment. The work of the boilers may be checked by arranging a meter so as to continuously measure the feed water; and its record, compared with the total weight of coal purchased, will frequently give the check desired. Such a check becomes more difficult when it is desirable to ascertain the performances of particular boilers, and the coal supply is common to all boilers; but by assigning particular weighed car loads of coal to the particular boilers, without any intimation to the firemen that they are being watched, it may be APPENDIX B S43 possible to ascertain the average performance of the boilers used for the particular purpose. Preliminary experiments of this kind conducted without notice to employees, and continued through a long period, will furnish a basis for comparison with elaborate tests, and it will then be possible to point out clearly where the several losses have taken place, and the testing engineer will get the credit for the saving shown. C. E. E. APPENDIX XXXV. The Ringelmann Smoke Chart. Professor Ringelmann, of Paris, has invented a system of determining the relative density or blackness of smoke, which has been communi- cated to the writer by Mr. Bryan Donkin, of London, and published in Engineering News of November 11, 1897. In making observations of the smoke proceeding from a chimney four cards ruled like those in the cut, together with a card printed in solid black and another left entirely white, are placed in a horizontal row and hung at a point about 50 feet from the observer and as nearly as convenient in line with the chimney. At this distance the lines become invisible, and the cards appear to be of different shades of gray, ranging from very light gray to almost black. The observer glances from the smoke coming from the chimney to the cards, which are numbered from to 5, determines which card most nearly corresponds with the color of the smoke and makes a record accordingly, noting the time. Observations should be made continu- ously during say one minute, and the estimated average density during that minute recorded, and so on, records being made once every minute. The average of all the records made during a boiler test is taken as the average figure for the smoke density during the test, and the whole of the record is plotted on cross-section paper in order to -show how the smoke varied in density from time to time. A rule by which the cards may be reproduced is given by Professor Ringelmann as follows: Card — All white. Card 1 — Black lines 1 mm. thick, 10 mm. apart, leaving spaces 9 mm. square. Card 2 — Lines 2.3 mm. thick, spaces 7.7 mm. square. Card 3 — - Lines 3.7 mm. thick, spaces 6.3 mm. square. Card 4 — Lines 5.5 mm. thick, spaces 4.5 mm. square. Card 5 — All black. The cards as printed in Fig. 452 are much smaller than those used by Professor Ringelmann, but the thickness and the spacing of the lines are the same. w. "K" 844 STEAM POWER PLANT ENGINEERING No.l No.2 No.3 No.4 Fig. 452. Ringelman Smoke Chart. APPENDIX B 845 APPENDIX XXXVIII. Chart Showing Graphically the Log of a Trial. The well-known method of plotting observations and data on cross- section paper and making a chart applying to the test is a useful means of representing the exact uniformity of conditions existing Lbs CHART SHOWING LOG OF BOILER TEST. '■"• 0CO ^ "^ 70 LOS. 1 1 1 A r-\ l/!\ j y ' /"' \ eo /\ ^St r omOau 8 «/\ A / \ A / \" Z ^=o ^N v^lC^^V^ Z l. ZIv Z L3 ^ i_ ^^ l±2w r ^ tz^s £ m °°° / 0EG . V | ° - 14 2 \W* V ■t r- ^ 1 ^ 140.000 , 2 -T_ a I 7± -_; Zflt _«^s ^ 3 L^ .a,^ S ~ 7 2i2-% ^%Xd£^ V Z ^ \ IX Ik- ' •^ 2 ^ 5 £ ^ 5 ^2 ^ Feed Terrtp \ ] .>■' \ 1 \ i t f 3 I ] mooo ^ |- r--^ a,^<- _j_ ^ ? 3 / ' ■ » „<•" \-^'° E • te^ ^ - "• oro " 80000 - - - - **£ i x 80.000 jr ^/ 16,000 £ / ,/ ^ ™' m Tfes. 7 1"' _g . 11.000 600 ^ •> ,0,000 _ _ . .. ^ FlueTemp . ' ^ 60,000 - p's-_~» ^r:_^2._^=2 = =2 = ^U« ! g^_=: = =2 = =:^^: _^^.» 10.000 5l? f i r_; /.. ' ^'=' . "^-^-°"" C 8000 «o.ooo -3 T'-s.*^ + ** 1 = **"''° ^o-^^-o^ 4 ^«g -3.000 K000 %r ^\ i-\_ +-*~~ W.0W ^ r -"j: 2 * N*^ /" _. »» "* 4000 SO.000 - - ^^^ . _< — ?:_ _ _ — _ _ 1 1 M 1 1 1 1 1 1 1 1 1 1 1 1 1 1 1 1 1 1 1 1 1 o Fig. 453, during a trial. Such a chart is illustrated in Fig. 453, in which the abscissae represent times and the ordinates, on appropriate scales, the various observations and data. g. h. b. APPENDIX C. RULES FOR CONDUCTING STEAM-ENGINE TESTS * Code of 1902. I. Object of Test. — Ascertain at the outset the specific object of the test, whether it be to determine the fulfillment of a contract guar- antee, to ascertain the highest economy obtainable, to find the working economy and defects under conditions as they exist, to ascertain the performance under special conditions, to determine the effect of changes in the conditions, or to find the performance of the entire boiler and engine plant, and prepare for the test accordingly. II. General Condition of the Plant. — Examine the engine and the entire plant concerned in the test; note its general condition and any points of design, construction, or operation which bear on the objects in view. Make a special examination of the valves and pistons for leakage by applying the working pressures with the engine at rest, and observe the quantity of steam, if any, blowing through per hour. If the trial has for an object the determination of the highest efficiency obtainable, the valves and pistons must first be made tight, and all parts of the engine and its auxiliaries, and all other parts of the plant concerned, should be put in the best possible working condition. III. Dimensions, etc. — Measure or check the dimensions of the cylinders in any case, this being done when they are hot. If they are much worn, the average diameter should be determined. Measure also the clearance, which should be done, if possible, by filling the spaces with water previously measured, the piston being placed at the end of the stroke. If the clearance cannot be measured directly, it can be determined approximately from the working drawings of the cylinder. Measure also the dimensions of auxiliaries and accessories, also those of the boilers so far as concerned in attaining the objects. It is well to supplement these determinations with a sketch or sketches showing the general features and arrangement of the different parts of the plant. * From the final report of the committee appointed to standardize a system of testing steam engines. Trans. A.S.M.E., Vol. XXIV. Greatly abridged. 846 APPENDIX C 847 IV. Coal. — When the trial involves the complete plant, embracing boilers as well as engine, determine the character of coal to be used. The class, name of the mine, size, moisture, and quality of the coal should be stated in the report. It is desirable, for purposes of com- parison, that the coal should be of some recognized standard quality for the locality where the plant is situated. V. Calibration of Instruments. — All instruments and apparatus should be calibrated and their reliability and accuracy verified by comparison with recognized standards. Such apparatus as is liable to change or become broken during a test, as gauges, indicator springs, and thermometers, should be calibrated before and after the test. The accuracy of scales should be verified by standard weights. When a water meter is used, special attention should be given to its cali- bration, verifying it both before and after the trial, and, if possible, during its progress, the conditions in regard to water pressure and rate of flow being made the same in the calibrations as exist through- out the trial. VI. Leakages of Steam, Water, etc. — In all tests except those of a complete plant made under conditions as they exist, the boiler and its connections, both steam and feed, as also the steam piping leading to the engine and its connections, should, so far as possible, be made tight. If absolute tightness cannot be obtained (in point of fact it rarely can be), proper allowance should be made for such leakage in determining the steam actually consumed by the engine. This, how- ever, is not required where a surface condenser is used and the water consumption is determined by measuring the discharge of the air pump. In such cases it is necessary to make sure that the condenser is tight, both before and after the test, against the entrance of circu- lating water, or if such occurs to make proper correction for it, deter- mining it under the working difference of pressure. Should there be excessive leakage of the condenser it should be remedied before the test is made. When the steam consumption is determined by measuring the discharge of the air pump, any leakage about the valve or piston rods of the engine should be carefully guarded against. Make sure that there is no leakage at any of the connections with the apparatus provided for measuring and supplying the feed water which could affect the results. All connections should, so far as pos- sible, be visible and be blanked off, and where this cannot be done, satisfactory assurance should be obtained that there is no leakage either in or out. 848 STEAM POWER PLANT ENGINEERING VII. Duration of Test. — The duration cf a test should depend largely upon its character and the objects in view. The standard heat test of an engine, and, likewise, a test for the simple determination of the feed- water consumption, should be continued for at least five hours, unless the class of service precludes a continuous run of so long duration. It is desirable to prolong the test the number of hours stated to obtain a number of consecutive hourly records as a guide in analyzing the reliability of the whole. Where the water discharged from the surface condenser is measured for successive short intervals of time, and the rate is found to be uni- form, the test may be of a much shorter duration than where the feed water is measured to the boiler. The longer the test with a given set of conditions the more accurate the work, and no test should be so short that it cannot be divided into several intervals which will give results agreeing substantially with each other. The commercial test of a complete plant, embracing boilers as well as engine, should continue at least one full day of twenty-four hours, whether the engine is in motion during the entire time or not. A continuous coal test of a boiler and engine should be of at least ten hours' duration, or the nearest multiple of the interval between times of cleaning fires. VIII. Starting and Stopping a Test. — (a) Standard Heat Test and Feed-Water Test of Engine : The engine having been brought to the normal condition of running, and operated a sufficient length of time to be thoroughly heated in all its parts, and the measuring apparatus having been adjusted and set to work, the height of water in the gauge glasses of the boilers is observed, the depth of water in the reservoir from which the feed water is supplied is noted, the exact time of day is observed, and the test held to commence. Thereafter the measurements determined upon for the test are begun and carried forward until its close. If practicable, the test may be commenced at some even hour or minute, but it is of the first importance to begin at such time as reliable observations of the water heights are obtained, whatever the exact time happens to be when these are satisfactorily determined. When the time for the close of the test arrives, the water should, if possible, be brought to the same height in the glasses and to the same depth in the feed- water reservoir as at the beginning, delaying the conclusion of the test if necessary to bring about this similarity of conditions. If differences occur, the proper corrections must be made. (6) Complete Engine and Boiler Test : For a continuous running test of combined engine or engines, and boiler or boilers, the same APPENDIX C 849 directions apply for beginning and ending the feed-water measurements as those just referred to under Section (a). The time of beginning and ending such a test should be the regular time of cleaning the fires, and the exact time of beginning and ending should be the time when the fires are fully cleaned, just preparatory to putting on fresh coal. In cases where there are a number of boilers, and it is inconvenient or undesirable to clean all fires at once, the time of beginning the test should be deferred until they are all cleaned and in a satisfactory state, all the fires being then burned down to a uniformly thin condition, the thickness and condition being estimated and the test begun just before firing the new coal previously weighed. The ending of the test is likewise deferred until the fires are all satisfactorily cleaned, being again burned down to the same uniformly thin condition as before, and the time of closing being taken just before replenishing the fires with new coal. For a commercial test of a combined engine and boiler, whether the engine runs continuously for the full twenty-four hours of the day or only a portion of the time, the fires in the boilers being banked during the time when the engine is not in motion, the beginning and ending of the test should occur at the regular time of cleaning the fires, the method followed being that already given. In cases where the engine is not in continuous motion, as, for example, in textile mills, where the working time is ten or eleven hours out of the twenty-four, and the fires are cleaned and banked at the close of the day's work, the best time for starting and stopping a test is the time just before banking, when the fires are well burned down and the thickness and condition can be most satisfactorily judged. In these, as in all other cases noted, the test should be begun by observing the exact time, the thickness and condi- tion of the fires on the grates, the height of water in the gauge glasses of the boilers, the depth of the water in the reservoir from which the feed water is supplied, and other conditions relating to the trial, the same observations being again taken at the end of the test, and the conditions in all respects being made as nearly as possible the same as at the beginning. IX. Measurement of Heat Units consumed by the Engine. — The measurement of the heat consumption requires the measurement of each supply of feed water to the boiler — that is, the water supplied by the main feed pump, that supplied by auxiliary pumps, such as jacket water, water from separators, drips, etc., and water supplied by gravity or other means; also the determination of the temperature of the water supplied from each source, together with the pressure and quality of the steam. The temperatures at the various points should be those applying to 850 STEAM POWER PLANT ENGINEERING the working conditions. The temperature of the feed water should be taken near the boiler. This causes the engine to suffer a disadvantage from the heat lost by radiation from the pipes which carry the water to the boiler, but it is, nevertheless, advisable on the score of simplicity. Such pipes would, therefore, be considered a portion of the engine plant. This conforms with the rule already recommended for the tests of pumping engines where the duty per million heat units is computed from the temperature of the feed water taken near the boiler. It frequently happens that the measurement of the water requires a change in the usual temperature of supply. For example, where the main supply is ordinarily drawn from a hot well in which the temperature is say 100 degrees F., it may be necessary, owing to the low level of the well, to take the supply from some source under a pressure or head sufficient to fill the weighing tanks used, and this supply may have a temperature much below that of the hot well; possibly as low as 40 degrees F. The temperature to be used is not the temperature of the water as weighed in this case, but that of the working temperature of the hot well. The working temperature in cases like this must be determined by a special test, and included in the log sheets. The heat to be determined is that used by the entire engine equip- ment, embracing the main cylinders and all auxiliary cylinders and mechanism concerned in the operation of the engine, including the air pump, circulating pump, and feed pumps, also the jacket and reheater when these are used. No deduction is to be made for steam used by auxiliaries unless these are shown by test to be unduly wasteful. In this matter an exception should be made in cases of guarantee tests where the engine contractor furnishes all the auxiliaries referred to. He should, in that case, be responsible for the whole, and no allowance should be made for inferior economy, if such exists. Should a deduction be made on account of the auxiliaries being unduly wasteful, the method of waste and its extent, as compared with the wastes of the main engine or other standard of known value, shall be reported definitely. The steam pressure and the quality of the steam are to be taken at some point conveniently near the throttle, valve. The quantity of steam used by the calorimeter must be determined and properly allowed for. (See Article XVI, on " Quality of Steam.") X. Measurement of Feed Water or Steam Consumption of Engine, etc. — The method of determining the steam consumption applicable to all plants is to measure all the feed water supplied to the boilers, and deduct therefrom the water discharged by separators and drips, as also the water and steam which escape on account of leakage of the boiler and its pipe connections and leakage of the steam main and branches APPENDIX C 851 connecting the boiler and the engine. In plants where the engine exhausts into a surface condenser the steam consumption can be measured by determining the quantity of water discharged by the air pump, corrected for any leakage of the condenser, and adding thereto the steam used by jackets, reheaters, and auxiliaries as determined independently. If the leakage of the condenser is too large to satis- factorily allow for it, the condenser should, of course, be repaired and the leakage again determined before making the test. In measuring the water it is best to carry it through a tank or tanks resting on platform weighing scales suitably arranged for the purpose, the water being afterwards emptied into a reservoir beneath, from which the pump is supplied. Where extremely large quantities of water must be measured, or in some places relatively small quantities, the orifice method of measuring is one that can be applied with satisfactory results. In this case the average head of water on the orifice must be determined, and, further- more, it is important that means should be at hand for calibrating the discharge of the orifice under the conditions of use. The corrections or deductions to be made for leakage above referred to should be applied only to the standard heat-unit test and tests for determining simply the steam or feed-water consumption, and not to coal tests of combined engine and boiler equipment. In the latter, no correction should be made except for leakage of valves connecting to other engines and boilers, or for steam used for purposes other than the operation of the plant under test. Losses of heat due to imperfections of the plant should be charged to the plant, and only such losses as are concerned in the working of the engine alone should be charged to the engine. In measuring jacket water or any supply under pressure which has a temperature exceeding 212 degrees F., the water should first be cooled, as may be done by discharging it into a tank of cold water previously weighed, or by passing it through a coil of pipe submerged in running and colder water, preventing thereby the loss of evapo- ration which occurs when such hot water is discharged into the open air. XI. Measurement of Steam used by Auxiliaries. — Although the steam used by the auxiliaries — embracing the air pump, circulating pump, feed pump, and any other apparatus of this nature, supposing them to be steam-driven, also the steam jackets, reheaters, etc., which consume steam required for the operation of the engine — is all included in the measurement of the steam consumption, as pointed out in Article X, yet it is highly desirable that the quantity of steam 852 STEAM POWER PLANT ENGINEERING used by the auxiliaries, and in many cases that used by each auxiliary, should be determined exactly, so that the net consumption of the main engine cylinders may be ascertained and a complete analysis made of the entire work of the engine plant. Where the auxiliary cylinders are non-condensing, the steam consumption can often be measured by carrying the exhaust for the purpose into a tank of cold water resting on scales or through a coil of pipe surrounded by cold running water. Another method is to run the auxiliaries as a whole, or one by one, from a spare boiler (preferably a small vertical one), and measure the feed water supplied to this boiler. The steam used by the air and circulating pumps may be measured by running them under, as near as possible, the working conditions and speed, the main engine and other auxiliaries being stopped, and testing the con- sumption by the measuring apparatus used on the main trial. For a short trial, to obtain approximate results, measurement can be made by the water-gauge glass method, the feed supply being shut off. When the engine has a surface condenser, the quantity of steam used by the auxiliaries may be ascertained by allowing the engine alone to exhaust into the condenser, measuring the feed water supplied to the boiler and the water discharged by the air pump, and subtracting one from the other, after allowing for losses by leakage. XII. Coal Measurement. — (a) Commercial Tests : In commercial tests of the combined engine and boiler equipment, or those made under ordinary conditions of commercial service, the test should, as pointed out in Article VII, extend over the entire period of the day; that is, twenty-four hours, or a number of days of that duration. Consequently, the coal consumption should be determined for the entire time. If the engine runs but a part of the time, and during the remaining portion the fires are banked, the measurement of coal should include that used for banking. It is well, however, in such cases, to determine separately the amount consumed during the time the engine is in operation and that consumed during the period while the fires are banked, so as to have complete data for purposes of analysis and comparison, using suitable precautions to obtain reliable measure- ments. The measurement of coal begins with the first firing, after cleaning the furnaces and burning down at the beginning of the test, as pointed out in Article VIII, and ends with the last firing, at the expiration of the allotted time. (b) Continuous Running Tests : In continuous running tests which, as pointed out in Article VII, cover one or more periods which elapse between the cleaning of the fires, the same principle applies as that mentioned under the above heading (a); viz., the coal measurement APPENDIX C 853 begins with the first firing, after cleaning and burning down, and the measurement ends with the last firing, before cleaning and burning down at the close of the trial. (c) Coal Tests in General : When not otherwise specially under- stood, a coal test of a combined engine and boiler plant is held to refer to the commercial test above noted, and the measurement of coal should conform thereto. In connection with coal measurements, whatever the class of tests, it is important to ascertain the percentage of moisture in the coal, the weight of ashes and refuse, and, where possible, the approximate and ultimate analysis of the coal, following all the methods and details advocated in the latest report of the Boiler Test Committee of the Society. (See Vol. XXI, p. 34.) (d) Other Fuels than Coal : For all other solid fuels than coal the same directions in regard to measurement should be followed as those given for coal. If the boilers are run with oil or gas, the measure- ments relating to stopping and starting are much simplified, because the fuel is burned as fast as supplied, and there is no body of fuel con- stantly in the furnace, as in the case of using solid fuel. When oil is used, it should be weighed, and when gas is used, it should be measured in a calibrated gas meter or a gasometer. XIII. Indicated Horse Power. — The indicated horse power should be determined from the average mean effective pressure of diagrams taken at intervals of twenty minutes, and at more frequent intervals if the nature of the test makes this necessary, for each end of each cylinder. With variable loads, such as those of engines driving gener- ators for electric railroad work, and of rubber-grinding and rolling-mill engines, the diagrams cannot be taken too often. In cases like the latter, one method of obtaining suitable averages is to take a series of diagrams on the same blank card without unhooking the driving cord, and apply the pencil at successive intervals of ten seconds until two minutes' time or more has elapsed, thereby obtaining a dozen or more indications in the time covered. This tends to insure the determina- tion of a fair average for that period. In taking diagrams for variable loads, as indeed for any load, the pencil should be applied long enough to cover several successive revolutions, so that the variations produced by the action of the governor may be properly recorded. To determine whether the governor is subject to what is called " racing " or " hunting," a " variation diagram " should be obtained; that is, one in which the pencil is applied a sufficient time to cover a complete cycle of variations. When the governor is found to be working in this manner, the defect should be remedied before proceeding with the test. 854 STEAM POWER PLANT ENGINEERING It is seldom necessary, as far as average power measurements are concerned, to obtain diagrams at precisely the same instant at the two ends of the cylinder, or at the same instant on all the cylinders, when there are more than one. All that is required is to take the diagrams at regular intervals. Should the diagrams vary so much among themselves that the average may not be a fair one, it signifies that they should be taken more frequently, and not that special care should be employed to obtain the diagrams of each set at precisely "the same time. When diagrams are taken during the time when the engine is working up to speed at the start, or when a study of valve setting and steam distribution is being made, they should be taken at as nearly the same time as practicable. In cases where the diagrams are to be taken simultaneously, the best plan is to have an operator stationed at each indicator. This is desirable, even where an electric or other device is employed to operate all the instruments at once ; for unless there are enough operators, it is necessary to open the indi- cator cocks some time before taking the diagrams and run the risk of clogging the pistons and heating the high-pressure springs above the ordinary working temperature. The most satisfactory driving rig for indicating seems to be some form of well-made pantagraph, with driving cord of fine annealed wire leading to the indicator. The reducing motion, whatever it may be, and the connections to the indicator, should be so perfect as to pro- duce diagrams of equal lengths when the same indicator is attached to either end of the cylinder, and produce a proportionate reduction of the motion of the piston at every point of the stroke, as proved by test. The use of a three-way cock and a single indicator connected to the two ends of the cylinder is not advised, except in cases where it is impracticable to use an indicator close to each end. If a three-way cock is used, the error produced should be determined and allowed for. To determine the average power developed in cases where the engine starts from rest during the progress of the trial, as in a commercial test of a plant where the engine runs only a portion of the twenty- four hours, a number of diagrams should be taken during the period of getting up speed and applying the working load, the corresponding speed for each set of diagrams being counted. The power shown by these diagrams for the proportionate time should be included in the average for the whole run, and the duration should be the time the throttle valve is open. XIV. Testing Indicator Springs. — To make a perfectly satisfactory comparison of indicator springs with standards, the calibration should be made, if this were practical, under the same conditions as those APPENDIX C 855 pertaining to their ordinary use. Owing to the fact that the pressure of the steam in the indicator cylinder and the corresponding temperature are undergoing continual changes, it becomes almost impossible to compare the springs with any standard under such conditions. There must be a constant pressure during the time that the comparison is being made. Although the best that can be done is not altogether satisfactory, it seems that we must be content with it. To bring the conditions as nearly as possible to those of the working indicator, the steam should be admitted to the indicator as short a time as practicable for each of the pressures tried, and then the indicator cock should be closed and the steam exhausted therefrom before another pressure is tried. By this means the parts are heated and cooled somewhat the same as under the working conditions. We recommend, therefore, that for each required pressure the first step be to open and close the indicator cock a number of times in quick succession, then to quickly draw the line on the paper for the desired record, observing the gauge or other standard at the instant when the line is drawn. A corresponding atmospheric line is taken immediately after obtaining the line at the given pressure, so as to eliminate any difference in the temperature of the parts of the indicator. This appears to be a better method (although less readily carried on and requiring more care) than the one heretofore more commonly used, where the indicator cock is kept continually open and the pressure is gradually rising or falling through the range of comparison. The calibration .should be made for at least five points, two of these being for the pressure corresponding as near as may be to the initial and back pressures, and three for intermediate points equally distant. For pressures above the atmosphere, the proper standard recom- mended is the dead-weight testing apparatus, or a reliable mercury column, or an accurate steam gauge proved correct, or of known error, by either of these standards. For pressures below the atmosphere the best standard to use is a mercury column. The correct scale of spring to be used for working out the mean effective pressure of the diagrams should be the average based on the calibration, and this may be ascertained in the manner pointed out below. XV. Brake Horse Power. — This term applies to the power delivered from the fly-wheel shaft of the engine. It is the power absorbed by a friction brake applied to the rim of the wheel or to the shaft. A form of brake is preferred that is self-adjusting to a certain extent, so that it will of itself tend to maintain a constant resistance at the rim of the wheel. One of the simplest brakes for comparatively small engines, 856 STEAM POWER PLANT ENGINEERING which may be made to embody this principle, consists of a cotton or hemp rope, or a number of ropes, encircling the wheel, arranged with weighing scales or other means for showing the strain. An ordinary band brake may also be constructed so as to embody the principle. The wheel should be provided with interior flanges for holding water used for keeping the rim cool. -^^^^^^^^^^^^^^^^^^^^^^^^^^^^^^j^^^^^^^^- Fig. 454. Rope Brakes. Fig. 455. A self-adjusting rope brake is illustrated in Fig. 454, where it will be seen that, if the friction at the rim of the wheel increases, it will lift the weight A, which action will diminish the tension in the end B of the rope and thus prevent a further increase in the friction. The same device can be used for a band brake of the ordinary construction. Where space below the wheel is limited, a cross bar, C, supported by a chain tackle exactly at its center point may be used as shown in Fig. 455, thereby causing the action of the weight on the brake to be up- ward. A safety stop should be used with either form, to prevent the weights being accidentally raised more than a certain amount. The water-friction brake is specially adapted for high speeds and has the advantage of being self-cooling. The Alden brake is also self- cooling and is capable of fine adjustment. A water-friction brake is shown in Fig. 456. It consists of two cir- cular disks, A and B, attached to the shaft C, and revolving in a case, E, between fixed planes. The space between the disks and planes is APPENDIX C 857 supplied with running water, which enters at D and escapes at the cocks F, G, and H. The friction of the water against the surfaces con- stitutes a resistance which absorbs the desired power, and the heat generated within is carried away by the water itself. The water is thrown outward by centrifugal action and fills the outer portion of the case. The greater the depth of the ring of water the greater the Fig. 456. Alden Absorption Dynamometer. amount of power absorbed. By suitably adjusting the amount of water entering and leaving any desired power can be obtained. Water-friction brakes have been used successfully at speeds of over 20,000 revolutions per minute. For description of the Alden brake, see Transactions, Vol. XI, p. 958. XVI. Quality of Steam. — When ordinary saturated steam is used, its quality should be obtained by the use of a throttling calorimeter attached to the main steam pipe near the throttle valve. When the steam is superheated, the amount of superheating should be found by the use of a thermometer placed in a thermometer-well filled with mer- cury, inserted in the pipe. The sampling pipe for the calorimeter should, if possible, be attached to a section of the main pipe having a vertical direction, with the steam preferably passing upward, and the sampling nozzle should be made of a half -inch pipe, having at least 20 one-eighth- inch holes in its perforated surface. The readings of the calorimeter should be corrected for radiation of the instrument, or they should be referred to a normal reading, as pointed out below. If the steam is 858 STEAM POWER PLANT ENGINEERING superheated, the amount of superheating should be obtained by refer- ring the reading of the thermometer to that of the same thermometer when the steam within the pipe is saturated, and not by taking the difference between the reading of the thermometer and the temper- ature of saturated steam at the observed pressure as given in a steam table. XVII. Speed. — There are several reliable methods of ascertaining speed, or the number of revolutions of the engine crank-shaft per minute. The simplest is the familiar method of counting the number of turns for a period of one minute with the eye fixed on the second hand of a timepiece. Another is the use of a counter held for a minute or a number of minutes against the end of the main shaft. Another is the use of a reliable calibrated tachometer held likewise against the end of the shaft. The most reliable method, and the one we recommend, is the use of a continuous recording engine register or counter, taking the total reading each time that the general test data are recorded, and computing the revolutions per minute corresponding to the difference in the readings of the instrument. When the speed is above 250 revolutions per minute, it is almost impossible to make a satisfactory counting of the revolutions without the use of some form of mechanical counter. The determination of variation of speed during a single revolution, or the effect of the fluctuation due to sudden changes of the load, is also desirable, especially in engines driving electric generators used for lighting purposes. There is at present no recognized standard method of making such determinations, and if such are desired, the method employed may be devised by the person making the test and described in detail in the report. XVIII. Recording the Data. — Take note of every event connected with the progress of the trial whether it seems at the time to be important or unimportant. Record the time of every event and time of taking every weight and every observation. Observe the pressures, temperatures, water heights, speeds, etc., every twenty or thirty minutes when the conditions are practically uniform, and at much more frequent intervals if the conditions vary. Observations which concern the feed-water measurements should be made with special care at the expiration of each hour of the trial, so as to divide the tests into hourly periods and show the uniformity of the conditions and results as the test goes forward. Where the water discharged from a surface condenser is weighed it may be advisable to divide the test by this means into periods of less than one hour. . APPENDIX C 859 The data and observations of the test should be kept on properly prepared blanks or in notebooks containing columns suitably arranged for a clear record. As different observers have their own individual ideas as to how such records should be kept, no special form of log sheet is given as a necessary part of the code. XIX. Uniformity of Conditions. — In a test having for an object the determination of the maximum economy obtainable from an engine, or where it is desired to ascertain with special accuracy the effect of predetermined conditions of operation, it is important that all the conditions under which the engine is operated should be main- tained uniformly constant. This requirement applies especially to the pressure, the speed, the load, the rate of feeding the various supplies of water, the height of water in the gauge glasses, and the depth of water in the feed-water reservoir. XX. Analysis of Indicator Diagrams. — (a) Steam accounted for by the Indicator : The simplest method of computing the steam accounted for by the indicator is the use of the formula M = 4 3 I 5 ^ [«? + E ) X Wc-(H + E) X Wh], which gives the weight in pounds per indicated horse power per hour. In this formula the symbol " M.E.P." refers to the mean effective pressure. In multiple-expansion engines this is the combined mean effective pressure referred to the cylinder in question. The symbol C refers to the proportion of the stroke completed at points on the expansion line of the diagram near the actual cut-off or release; the symbol H to the proportion of compression; and the symbol E to the proportion of clearance; all of which are determined from the indi- cator diagram. The symbol Wc refers to the weight of one cubic foot of steam at the cut-off or release pressure; and the symbol Wh to the weight of one cubic foot of steam at the compression pressure; these weights being taken from steam tables of recognized accuracy. The points near the cut-off and release on the expansion line and the point on the compression line are located as shown on the sample diagram, Fig. 457. They are the points in the case of the expansion and com- pression lines of the diagram which mark the complete closure of the valve. The point near the cut-off, for example, lies where the curve of expansion begins after the rounding of the diagram due to the wire- drawing which occurs while the valve is closing. This cut-off may be located by finding the point where the curve is tangent to a hyper- bolic curve. 860 STEAM POWER PLANT ENGINEERING Should the point in the compression curve be at the same height as the point in the expansion curve, then Wc = Wh, and the formula becomes in which (C — H) represents the distance between the two points divided by the length of the diagram. When the load and all other conditions are substantially uniform, it is unnecessary to work up the steam accounted for by the indicator Compression Atmospheric Line Fig. 457. Showing Points where " Steam Accounted for by Indicator "is Computed. from all the diagrams taken. Five or more sample diagrams may be selected and the computations based on the samples instead of on the whole. (b) Sample Indicator Diagrams : In order that the report of a test may afford complete information regarding the conditions of the test, sample indicator diagrams should be selected from those taken and copies appended to the tables of results. In cases where the engine is of the multiple-expansion type these sample diagrams may also be arranged in the form of a " combined " diagram. (c) The Point of Cut-off: The term " cut-off" as applied to steam engines, although somewhat indefinite, is usually considered to be at an earlier point in the stroke than the beginning of the real expansion line. That the cut-off point may be defined in exact terms for commercial purposes, as used in steam-engine specifications and contracts, the Committee recommends that, unless otherwise specified, the commercial cut-off, which seems to be an appropriate expression for this term, be ascertained as follows: Through a point showing the maximum pressure APPENDIX C 861 during admission draw a line parallel to the atmospheric line. Through the point on the expansion line near the actual cut-off, referred to in Section XX (a), draw a hyperbolic curve. The point where these two lines intersect is to be considered the commercial cut-off point. The percentage is then found by dividing the length of the diagram measured to this point by the total length of the diagram and multiplying the result by 100. E C B A Four Valve Engine, Slow Speed, Commercial Cut-off= I?. The principle involved in locating the commercial cut-off is shown in Figs. 458 and 459, the first of which represents a diagram from a slow- speed Corliss engine and the second a diagram from a single-valve high-speed engine. In the latter case where, owing to the fling of the Fig. 459. Single Valve Engine, High-Speed, Com- BC mercial Cut-off = AC" pencil, the steam line vibrates, the maximum pressure is found by taking a mean of the vibrations at the highest point. The commercial cut-off as thus determined is situated at an earlier 862 STEAM POWER PLANT ENGINEERING point of the stroke than the actual cut-off referred to in computing the " steam accounted for " by the indicator in Section XX (a). (d) Ratio of Expansion : The ratio of expansion for a simple engine is determined by dividing the volume corresponding to the piston dis- placement, including clearance, by the volume of the steam at the com- mercial cut-off, including clearance. In a multiple-expansion engine it is determined by dividing the net volume of the steam indicated by the low-pressure diagram at the end of the expansion line, assumed to be continued to the end of the stroke, by the net volume of the steam at the maximum pressure during admis- sion to the high-pressure cylinder. (e) Diagram Factor : The diagram factor is the proportion borne by the actual mean effective pressure measured from the indicator diagram to that of a diagram in which the various operations of admission, expansion, release, and compression are carried on under assumed con- ditions. The factor recommended refers to an ideal diagram which represents the maximum power obtainable from the steam accounted for by the indicator diagrams at the point of cut-off, assuming first that the engine has no clearance; second, that there are no losses through wire-drawing the steam either during the admission or the release; third, that the expansion line is a hyperbolic curve; and fourth, that the initial pressure is that of the boiler and the back pressure that of the atmosphere for a non-condensing engine and of the condenser for a condensing engine. The diagram factor is useful for comparing the steam distribution losses in different engines, and is of special use to the engine designer, for by multiplying the mean effective pressure obtained from the assumed theoretical diagrams by it he will obtain the actual mean effective pressure that should be developed in an engine of the type considered. The expansion and compression curves are taken as hyperbolas, because such curves are ordinarily used by engine builders in their work, and a diagram based on such curves will be more useful to them than one where the curves are constructed according to a more exact law. In cases where there is a considerable loss of pressure between the boiler and the engine, as where steam is transmitted from a central plant to a number of consumers, the pressure of the steam in the supply main should be used in place of the boiler pressure in constructing the diagrams. XXI. Standards of Economy and Efficiency. — The hourly consump- tion of heat, determined by employing the actual temperature of the feed water to the boiler, as pointed out in Article IX of the Code, divided APPENDIX C 863 by the indicated and brake horse power, that is, the number of heat units consumed per indicated and per brake horse power per hour, is the standard of engine efficiency recommended by the Committee. The consumption per hour is chosen rather than the consumption per minute, so as to conform with the designation of time applied to the more familiar units of coal and water measurement which have hereto- fore been used. The British standard, where the temperature of the feed water is taken as that corresponding to the temperature of the back-pressure steam, allowance being made for any drips from jackets or reheaters, is also included in the tables. It is useful in this connection to express the efficiency in its more scientific form, or what is called the " thermal efficiency ratio." The thermal efficiency ratio is the proportion which the heat equivalent of the power developed bears to the total amount of heat actually con- sumed, as determined by test. The heat converted into work repre- sented by one horse power is 1,980,000 foot-pounds per hour, and this divided by 778 equals 2545 British thermal units. Consequently the thermal efficiency ratio is expressed by the fraction 2545 2545 B.T.U. per H.P. per hour' XXII. Heat Analysis. — For certain scientific investigations it is useful to make a heat analysis of the diagram to show the interchange of heat from steam to cylinder walls, etc., which is going on within the cylinder. This is unnecessary for commercial tests. XXIII. Temperature- Entropy Diagram. — The study of the heat analysis is facilitated by the use of the temperature-entropy diagram in which areas represent quantities of heat, the coordinates being the absolute temperature and entropy. Such a diagram is shown in Fig. 460. When the quantities given in the steam tables are plotted, two curves, AA and BE, are obtained which may be termed the water line and the steam line, AA being the logarithmic curve if the specific heat of the water is taken as constant. The diagram refers to a unit weight of the agent, and the heat necessary to raise a pound of water from the temperature ma to the temperature pa' and evaporate it at that temperature is represented by the area aa'b'qm. If the steam be now expanded adiabatically the temperature will fall to qs and as x per cent = - — will remain as steam, the rest being liquefied. If the ab steam is now rejected, it carries away with it the heat sqma, the work 864 STEAM POWER PLANT ENGINEERING area being a'b'sa, from which must be deducted the work w (ex- pressed in heat units) to pump a pound of water into the boiler. The efficiency of this cycle is evidently in which X = + L l — xL 2 — w h + L t ' , T i Li log c — + — ar + p'y 6c T 2 T. x ab L„ By the action of the walls a portion of the steam is liquefied prior to the expansion, which therefore begins at e, and since the cooling Fig. 460. Fig. 461. Temperature-Entropy Diagrams. action of the walls continues, the expansion line falls off to ef, from which point a reverse action takes place and the expansion line bends APPENDIX C 865 over to g. Finally, since the release takes place before the condenser temperature is reached, the heat rejection starts at g, following a line of equal volume until the exhaust port temperature is reached at /. If heat is added during expansion enough to keep the steam theo- retically saturated, as, for example, by a water jacket, such additional heat is represented by the area b'bnq, and the additional work obtained by the triangle b'bs. If the steam is superheated sufficiently to give by expansion theoretically dry steam at the end, such additional heat is represented by the area b'vnq and the additional work by b'vbs. Neither of these extra amounts of work is realized in practice, and it is evident from the diagram that the heat thus applied is in both cases less efficient than in the principal cycle. Nevertheless the action in each case is to bring the point e nearer the point &' and to effect a notable net economy. The Carnot cycle would be obtained if in the Rankine cycle the rejection of heat were stopped at r and the temperature of the mix- ture raised to a' by compression. This cannot be practically accom- plished, but a system of feed-water heaters has been suggested and exemplified in the Nordberg engine, which is theoretically a close equivalent to it. Where steam is expanded in say three cylinders, the feed water may be successively heated from the receiver intermediate between each pair, the effect of which is illustrated in Fig. 461. The expansion line follows the heavy line,, being carried over to y by the first feed-water heater and to y' by the second feed-water heater. With an infinite number of such feed-water heaters, the line yy f would be parallel to aa', and the cycle equivalent to that of Carnot. XXIV. Ratio of Economy of an Engine to that of an Ideal Engine. — The ideal engine recommended for obtaining this ratio is that which was adopted by the Committee appointed by the Civil Engineers of London to consider and report a standard thermal efficiency for steam engines. This engine is one which follows the Rankine cycle, where steam at a constant pressure is admitted into the cylinder with no clearance, and after the point of cut-off is expanded adiabat- ically to the back pressure. In obtaining the economy of this engine the feed water is assumed to be returned to the boiler at the exhaust temperature. Such a cycle is preferable to the Carnot for the purpose at hand, because the Carnot cycle is theoretically impossible for an engine using superheated steam produced at a constant pressure, and the gain in efficiency for superheated steam corresponding to the Carnot efficiency will be much greater than that possible for the actual cycle. The ratio of the economy of an engine to that of the ideal engine is obtained by dividing the heat consumption per indicated 866 STEAM POWER PLANT ENGINEERING horse power per minute for the ideal engine by that of the actual engine. XXV. Miscellaneous. — In the case of tests of combined engines and boiler plants, where the full data of the boiler performance is to be determined, reference should be made to the directions given by the Boiler Test Committee of the Society, Code of 1899. (See Vol. XXI, p. 34.) In tests made for scientific research, and in those made on special forms of engines, the line of procedure must be varied according to the special objects in view, and it has been deemed unnecessary to go into particulars applying to such tests. In testing steam pumping engines and locomotives in accordance with the standard methods of conducting such tests, recommended by the committees of the Society, reference should be made to the reports of those committees in the Transactions, Vol. XII, p. 530, and in Vol. XIV, p. 1312. XXVI. Report of Test. — The data and results of the test should be reported in the manner and in the order outlined in one of the fol- lowing tables, the first of which gives, it is hoped, a complete sum- mary of all the data and results as applied not only to the standard heat-unit test but also to tests of combined engine and boiler for deter- mining all questions of performance, whatever the class of service; the second refers to a short form of report giving the necessary data and results for the standard heat test; and the third to a short form of report for a feed-water test. It is the intention that the tables should be full enough to apply to any type of engine, but where not so, or where special data and results are determined, additional results may be inserted under the appropriate headings. Although these forms are arranged so as to be used for expressing the principal data and results of tests of pumping engines and locomotives, as well as for all other classes of steam engines, it is not the intention that they shall supplant the forms recommended by the committees on Duty Trials and Locomotives in cases where the full report of a test of such engines is desired. _ It is recommended that any report be supplemented by a chart in which the data of the test are graphically presented. (As an example of such a chart as applied to a boiler test, see Vol. XXI, p. 104.) TABLE NO. 1. Not reprinted here. See Trans. A.S.M.E. 24-702. APPENDIX C 867 TABLE NO. 2. DATA AND RESULTS OF STANDARD HEAT TEST OF STEAM ENGINE. Arranged according to the Short Form advised by the Engine Test Committee of the American Society of Mechanical Engineers. Code of 1902. 1. Made by of on engine located at to determine 2. Date of trial 3. Type and class of engine ; also of condenser 4. Dimensions of main engine (a) Diameter of cylinder in. (b) Stroke of piston ft. (c) Diameter of piston rod in. (d) Average clearance p. c. (e) Ratio of volume of cylinder to high- pressure cylinder (/) Horse-power constant for one pound mean effective pressure and one revolution per minute 5. Dimensions and type of auxiliaries 1st Cyl. 2d Cyl. 3d Cyl. Total Quantities, Time, etc. 6. Duration of test hours. 7. Total water fed to boilers from main source of supply pounds. 8. Total water fed from auxiliary supplies : (a) (6) (0 9. Total water fed to boilers from all sources pounds. 10. Moisture in steam or superheating near throttle p. c. or deg. 11. Factor of correction for quality of steam 12. Total dry steam consumed for all purposes pounds Hourly Quantities. 13. Water fed from main source of supply " 14. Water fed from auxiliary supplies: (a) (b) (c) 15. Total water fed to boilers per hour 16. Total dry steam consumed per hour 17. Loss of steam and water per hour due to drips from main steam pipes and to leakage of plant 18. Net dry steam consumed per hour by engine and auxiliaries 868 STEAM POWER PLANT ENGINEERING Pressures and Temperatures (Corrected). 19. Pressure in steam pipe near throttle by gauge lb. per sq. in. 20. Barometric pressure of atmosphere in inches of mercury.. . inches. 21. Pressure in receivers by gauge lb. per sq. in. 22. Vacuum in condenser in inches of mercury inches. 23. Pressure in jackets and reheaters by gauge lb. per sq. in. 24. Temperature of main supply of feed water degrees F. 25. Temperature of auxiliary supplies of feed water: («) (&) (c) 26. Ideal feed-water temperature corresponding to pressure of steam in the exhaust pipe, allowance being made for heat derived from jacket or reheater drips " Data Relating to Heat Measurement. 27. Heat units per pound of feed water, main supply 28. Heat units per pound of feed water, auxiliary supplies : («) (b) (c) 29. Heat units consumed per hour, main supply 30. Heat units consumed per hour, auxiliary supplies: B.T.U. 31. 32. 33. 34. (a) (6) (c). Total heat units consumed per hour for all purposes Loss of heat per hour due to leakage of plant, drips, etc Net heat units consumed per hour: (a) By engine alone (b) By auxiliaries Heat units consumed per hour by engine alone, reckoned from temperature given in line 26 35. 36. 38. 39. Indicator Diagrams. Commercial cut-off in per cent of stroke Initial pressure in pounds per square inch above atmosphere Back pressure at mid-stroke above or below at- mosphere in pounds per square inch Mean effective pressure in pounds per square inch Equivalent mean effective pressure in pounds per square inch : (a) Referred to first cylinder (6) Referred to second cylinder (c) Referred to third cylinder 1st Cyl. 2d Cyl. 3d Cyl. APPENDIX C 869 IstCyl. 2dCyl. 3d Cyl. 40. Pressures and percentages used in computing the steam accounted for by the indicator diagrams, measured to points on the expan- sion and compression curves Pressure above zero in pounds per square inch: (a) Near cut-off (6) Near release (c) Near beginning of compression Percentage of stroke at points where pressures are measured: (a) Near cut-off (6) Near release (c) Near beginning of compression 41. Steam accounted for by indicator in pounds per I.H.P. per hour: (a) Near cut-off (6) Near release 42. Ratio of expansion Speed 43. Revolutions per minute revolutions. Power. 44:. Indicated horse power developed by main-engine cylinders: First cylinder horse power. Second cylinder " Third cylinder " Total 45. Brake horse power developed by engine " Standard Efficiency and other Results * 46. Heat units consumed by engine and auxiliaries per hour: (a) per indicated horse power B.T.TJ. (6) per brake horse power " 47. Equivalent standard coal in pounds per hour: (a) per indicated horse power pounds. (6) per brake horse power " 48. Heat units consumed by main engine per hour corresponding to ideal maximum temperature of feed water given in line 26, British standard: (a) per indicated horse power B.T.U. (b) per brake horse power " 49. Dry steam consumed per indicated horse power per hour: (a) Main cylinders, including jackets pounds. (6) Auxiliary cylinders " (c) Engine and auxiliaries " * The horse power referred to above (items 46-50) is that of the main engine, exclusive of auxiliaries. 870 STEAM POWER PLANT ENGINEERING 50. Dry steam consumed per brake horse power per hour: (a) Main cylinders, including jackets pounds. (6) Auxiliary cylinders " (c) Engine and auxiliaries " 51. Percentage of steam used by main-engine cylinders accounted for by indicator diagrams, near cut-off of high-pressure cylinder per cent. Additional Data. Add any additional data bearing on the particular objects of the test or relating to the special class of service for which the engine is used. Also give copies of indicator diagrams nearest the mean, and the corresponding scales. TABLE NO. 3. DATA AND RESULTS OF FEED-WATER TEST OF STEAM ENGINE. Arranged according to the Short Form advised by the Engine Test Committee of the American Society of Mechanical Engineers. Code of 1902. 1. Made by of , on engine located at to determine 2. Date of trial 3. Type of engine (simple, compound, or other multiple expansion; condensing or non-condensing) 4. Class of engine (mill, marine, locomotive, pumping, electric, or other) 5. Rated power of engine 6. Name of builders 7. Number and arrangement of cylinders of engine; how lagged; type of valves and of condensers 8. Dimensions of engine lst CyL 2d CyL 3d CyL (a) Single or double acting (6) Cylinder dimensions: Bore in. Stroke , ft. Diameter of piston rod in. Diameter of tail rod in. (c) Clearance in per cent of volume displaced by piston per stroke: Head end Crank end Average (d) Ratio of volume of each cylinder to volume of high-pressure cylinder. . . . (e) Horse-power constant for one pound mean effective pressure and one revolution per minute APPENDIX C 871 Total Quantities, Time, etc. 9. Duration of test hours. 10. Water fed to boilers from main source of supply pounds. 11. Water fed from auxiliary supplies: (a) (b) (c) 12. Total water fed from all sources 13. Moisture in steam or superheating near throttle * p. c. or deg. 14. Factor of correction for quality of steam 15. Total dry steam consumed for all purposes pounds. Hourly Quantities. 16. Water fed from main source of supply " 17. Water fed from auxiliary supplies: (a) (6) (c) 18. Total water fed to boilers per hour " 19. Total dry steam consumed per hour " 20. Loss of steam and water per hour due to leakage of plant, drips, etc ' 21. Net dry steam consumed per hour by engine and auxiliaries " 22. Dry steam consumed per hour: (a) Main cylinders " (b) Jackets and reheaters " Pressures and Temperatures (Corrected). 23. Steam pipe pressure near throttle, by gauge lb. per sq. in. 24. Barometric pressure of atmosphere in inches of mercury inches. 25. Pressure in first receiver by gauge lb. per sq. in. 26. Pressure in second receiver by gauge 27. Vacuum in condenser: (a) In inches of mercury inches. (b) Corresponding total pressure lb. per sq. in. 28. Pressure in steam jackets by gauge lb. per sq. in. 29. Pressure in reheater by gauge 30. Superheating of steam in first receiver degrees F. 31. Superheating of steam in second receiver Indicator Diagrams. 1st Cyl. 2d Cyl. 3d Cyl. 32. Commercial cut-off in per cent of stroke 33. Initial pressure in pounds per square inch above atmosphere * In case of superheated steam engines, determine, if practicable, the temperature of the steam in each cvlinder. 872 STEAM POWER PLANT ENGINEERING IstCyl. 2dCyl. 3dCyl. 34. Back pressure at mid-stroke above or below atmosphere in pounds per square inch 35. Mean effective pressure in pounds per square inch 36. Equivalent mean effective pressure in pounds per square inch per indicated horse power (a) Referred to first cylinder. (b) Referred to second cylinder. (c) Referred to third cylinder. 37. Pressures and percentages used in computing the steam accounted for by the indicator diagrams, measured to points on the expan- sion and compression curves . Pressures above zero in pounds per square inch: (a) Near cut-off (6) Near release (c) Near beginning of compression Percentage of stroke at points where pressures are measured: (a) Near cut-off (6) Near release (c) Near beginning of compression 38. Aggregate M.E.P. in pounds per square inch referred to each cylinder given in heading 39. Mean back pressure above zero, pounds per square inch 40. Steam accounted for in pounds per indicated horse power per hour: (a) Near cut-off (6) Near release 41. Ratio of expansion: (a) Commercial (6) Ideal Speed. 42. Revolutions per minute revolutions. 43. Piston speed per minute feet. Power. 44. Indicated horse power developed by main-engine cylinders: First cylinder horse power. Second cylinder Third cylinder Total " APPENDIX C 873 Efficiency Results. 45. Dry steam consumed per indicated horse power per hour: (a) Main cylinder, including jackets pounds. (6) Auxiliary cylinders, etc " (c) Engine and auxiliaries 46. Percentage of steam used by main-engine cylinders accounted for by indicator diagrams : IstCyl. 2dCyl. 3d Cyl. (a) Near cut-off (6) Near release Sample Diagrams. Copies of indicator diagrams, nearest the mean, with corresponding scales, should be given in connection with table. 874 STEAM POWER PLANT ENGINEERING s X S3 * g CO h O CO W ►— 1 H W O PC A Density Weight per Cubic Foot, Pounds. ?- 000430 000679 000961 001259 001555 001850 002143 002431 002719 00300 00576 00845 01107 01364 01616 01867 02115 02361 02606 02849 03090 © © © o o o © © © © © © © © © © © © © © © 2 I a o OQ O O O o o o © © n © iO IO CO © »o CO 00 ©NO »0 CM CO OO © CO co t-h co 2938 1540 1041 © ■* -* N © »0 NNffi © r- 1 CO ■*** ^ CO CO CO OO co n t-h CO T-H T-H © CO i-H © N N CO N CM *0 "4< "* 00 lO CM CO CO CO o is H c W + 1728 1127 0775 © oo -* CO CO oo o o o CO CM CO io ^ ->*i © © 00 © © © "tfi © OO IOOOTH 05 © 00 "* CM lO t-h CO 00 © ^ CM OO OO 00 8161 8053 7958 t- © CM OONN N N N INMN CM CM CM Entropy of the Vapor. s-l^. 1666 0704 0135 9730 9415 9155 CONOO co -* n 05 N IO OO 00 00 I- T-H O CM CO Tt< ■* -tf OO OONCO CO "* ^i ^H OO t-H "* © 00 CO © lO CM © CM OO 00 o lO CO CM *o IO »o CM lO © Tt< 05 © OCCN IO ■* Tt< MINN Entropy of the Liquid. «> CO CM -* O "* CO o o o © CO © O CM CM OO © © © O i-H 1117 1195 1265 N- © 00 CM ^" © CO N © T-H T-H CM OO 00 i-H © "<*l N T-H CO UO CM CM CM 05 CO © n n »o lOtON CM CM CM CM CM N CO © CO OO © © o o o © © © © © © © © © © © © © © © © © © Heat Equivalent of External Work. © CM © m co os -* © "tf N- CO 00 © © 00 ■* © © t-h iO 00 ^CON oo © © O-H-H © © CO THT) n O O lO lO -* •* OS co co OOiO-H t-h ■<* CO OO © ■«*' © t-H 05 N OO CM thNOS CO t-h CM cm co ONCO -HH -HH lO OO CM © iO © CO 05 TP © CO © © © © N- CM CO CO T* © © Tji iO IO t-H IO OS © co co Tempera- ture, Degrees F. - O iO © O T-l •«* HQOW 05 © CO OO © CO rH"*M CO lO CM OO t-h IO t-h OO CO © CM © IO © N QO 00 CM CM iO © INN05 lO CO ■* CO »o CO 0(0 t- 00 © t-h CM CO •* »0 © N00 05 o o o © © © © © © © t-h CM . APPENDIX D 875 OON CO O MtOM 000-* OOHM lO lO lO ^M(M H O) N -^< ,-m oO »0 CM N- MhN (N N CO MiON OOOl-H CM ■<*< lO (Ot^OO O) O H N IN CO Tt< lO lO CO N N- OO O O OO^H co co co co -* co nooo o *-h cm co io co nqoq o >— i cm co tt 10 co n oo o <— < cm OOO OOO OOO r- 1.— It-h ^ht-It— I i— I .— i ,— i CM CM CM N cs n- oo -* COION NlOO t-h CO CM <-h n- O OO -* y-4 o ffliON oo co "* o n- o co -*r oo CM O OO CO CO CM CM OO lO N- lO -* 1-1 O CM CO —I ^H COINH -<*< •* lO O O CO t-HOOt-H ICrHlO (NOON NOH lOr- IO O - n- n n N- cococo cococo cococo cococo cococo io«oio >o 10 10 10 10 irj O CO N- COiCtJ* t-hOi-h "«cH 00 O O^CO OOlOh ^-(t^t--. t-hoOO ^00! (N»C^ COM^ i— I CO O r-KO-* ^(00 COCOO) NCOCO O CO N 050CO »0 OO O -* N- .— I COlO^f ^* OO COOCO CO-^CO l— <000 N CD lO -"CHCOCM i— I i— I O 050000 N CO CO "*-*!"* ■* CO CO COCOCM CMCMCM CM CM i-l i-Ht-It-h ,-h ,-h ,_i ,_,,_, ,_( OOO OOO IOhCO CO lO CM O00O HCOCO CM ■"*<>— I -*»0O ■"*! TH CO 05-^l>- 050500 "O (N N CMO0i-H COIOCO OC CO CM CM >— I 05 t^-^^H N- CO O -* O -* 00 CO N H lO O CO N O O O i— I rHCOlO CO00O5 (Z> t-* r— I 0 O lO O W h N COOCO COOCO COON- »0 CM O N lO — I Tt< OOCJH *HCO t— IN-CO *OCO IOIOCO O»00 CO lO CO N- CO •* i-iCDOO OO N CO COOCO CM lO CM N-CMCO O "* 00 CMlOO (N lO 00 th >* N OOOO i— i CM CO -*iO»0 COCON NOOOO 050505 OOO ,— i ,— i ,— i CMCMCM rtriM CMCMCM CMCMCM CMCMCM CMCMCM CMCMCM COCOCO COCOCO COCOCO «OC5CM COOOO 005N- •* r-i N (N OO (N N-CMCO 0"*N- i— I ■* 00 r— I •* N. OCOlO O O i— I ^ (N ■* lOiOCO N OO CO OOO O^Hi— I CMCMCM COCOCO "* ■* 'SP lO lO lO CMCMCM CMCMCM CMCMCM CMCMCM CMCMCO COCOCO COCOCO COCOCO COCOCO COCOCO 876 STEAM POWER PLANT ENGINEERING Density Weight per Cubic Foot, Pounds. «►> CN CN CN CO ■* LO CO CO CO CO OO CO COW* CO N OO CO CO CO OONN * LO LO OOrH CN CONN CNMTtl N 00 00 LO CO N "* -* -* O) OS O 00 OS CN ■*">*' >o t-H CO LO ** OS ■* LO to co o o o o o o o o o o o o O O O O O O o o o 11 II =0 CN O TtH O OS 00 CO LO CN lONO N CO CO CO OO CO CO CD O LO * * CO O N -* os co CO CN CN N O0 t— I 00 CO OS i-l r-H O CO ■* * ■* CD CN O O OS O CO T-H LO 00 LO OO CO LO MM(N CN CN CN CN CN CN CN CN CN CN CN CN CN CN i-H T-, ,-H T-H o ^ H + CN -* OS OS CD CO CO CD CO LO O N H05CO CO LO LO LO LO lO CO O OO -tfl CN OS LO LO TtH LO LO LO CO CD CO N LO CO lo to to CO 00 OS t-h OS N -* co co LO LO LO i-H t-H OS CO ■* o CO CO CO »0 LO LO CO OS OS N OS CN CN T-H T-H LO LO LO Entropy of the Vapor. s.|h O OS T-H lOOOM LO •* * o o o CO i-H 00 N CN CO CO CO CN o o o 0215 0164 0114 0066 0019 9973 O0 LO .— '. MOC* OS 00 OO Os os Os OS OO CO OJiON n n co OS OS OS 9600 9419 9251 o o o o o o o o o Entropy of the Liquid. - H05N CO LO LO N- t^ !>. to rji CN LO LO LO N N N O 00 LO LO TjH ■* N N N CN -* CO ■* CO CN N- t^ N- T II << rfOW O -* OS "* OO CO NHlfl 00 CN CO OS CN OS LO O0 T-H CO •* rji 05 0)05 LO LO LO OS OS OS CO CO N os os os N O0 OO OS OS OS 00 OS OS os os os OS O O OS o o i-H CN CN t-h CN ■* O O O CN CN CN Heat of Vapori- zation. i. CN O 00 CO N N OO O0 OS ON'* CO OS CN •* 00 LO CO CO CO CO i— I 00 CO CD LO 00 00 OO CO -* CN LO LO LO 00 OO 00 O OO CO LO * * O0 00 OO LO CO r-H OO 00 OO OS N CO CO CO CO OO OO 00 ■* CN OS CO CO CN 00 00 00 CO 00 i-H CN T-H T-H OO OO OO Heat of the Liquid. O CN OS CO CN N CN (DC* NOSrH CN •* Tt< tO LO T* CN CN N O CN to CO CO CO CO CO CO OO O CO CO CO CO LO OO o -* TJH LO CO CO CO CN * N LO LO LO CO CO CO OS t-H CO LO CO CO CO CO CO lO N t-H CO CO N CO CO CO lO * CN N 00 OS CO CO CO Tempera- ture, Degrees F. ~ LO O CO O lO OO i-H •* CO O0 OS O O O OS OS 00 "* T-H LO LO OO i— I CO LO CO CO CO CO CO CD OO O CO CO N CO CO CO CO tO N N N N CO CO CO OS t-H Tfl N O0 OO CO CO CO CO OO OS O0 O0 OO CO CO CO H«N OS OS OS CO CO CO HC5N O O t-H Absolute Pressure, Pounds per Square Inch. a, O to o LO lO CO LO O to (ONN o to o OO OO OS LO CD LO OS O CD t-h CN CN O to o T-H T-H CN CN CN CN LO o o CN CO ■* CN CN CN CD to O tO N O CN CN CO APPENDIX E. EQUIVALENT VALUES OF ELECTRICAL AND MECHANICAL UNITS. 1 Kilowatt Hour = 1,000 watt hours 1.34 horse-power hours 2,654,200 foot-pounds 3,600,000 joules 367,000 kilogram meters 3.53 pounds of water evaporated from and at 212° F. 1 Kilowatt = 1,000 watts 1.34 horse power 1.358 cheval-vapeur 2,654,200 foot-pounds per hour 44,240 foot-pounds per minute 737.3 foot-pounds per second 3,412 B.T.U. per hour 56.9 B.T.U. per minute 0.948 B.T.U. per second 3.53 pounds of water evaporated from and at 212° F. 1 Joule = 1 watt second 0.000000278 kilowatt hour 0.102 kilogram meter 0.0009477 B.T.U. 0.7373 foot-pound 1 Watt = 1 joule per second 0.00134 horse power 3.412 B.T.U. per hour 0.7373 foot-pound per second 0.0035 pound water evaporated from and at 212° F. 1 Kilogram-Meter = 7.233 foot-pounds 0.00000365 horse-power hour 0.00000272 kilowatt hour 0.0093 B.T.U. 1 Horse-Power Hour = 0.746 kilowatt hours 1,980,000 foot-pounds 2,545 B.T.U. 273,740 kilogram meters 2.64 pounds of water evaporated from and at 212° F. 1 Horse Power = 746 watts 0.746 kilowatts 1.0136 cheval-vapeur 33,000 foot-pounds per minute 550 foot-pounds per second 2,545 B.T.U. per hour 42.4 B.T.U. per minute 0.707 B.T.U. per second 2.64 pounds of water evaporated from and at 212° F. 1 Foot-Pound = 1.356 joules 0.1383 kilogram meter 0.000000377 kilowatt hour 0.001285 B.T.U. 0.0000005 horse-power hour 1 B.T.U. = 1,055 watt seconds 778 foot-pounds 107.6 kilogram meters 0.000293 kilowatt hour 0.000393 horse-power hour 0.001036 pound water evaporated from and at 212° F. 1 Cheval-Vapeur = 75 kilogrammeters per second 0.9863 horse power 0.7357 kilowatt b77 APPENDIX F. MISCELLANEOUS CONVERSION FACTORS. 1 Pound per Square Inch = 2.0355 inches of mercury at 32° F. 2.0416 inches of mercury at 62° F. 2.309 feet of water at 62° F. 0.07031 kilogram per square centi- meter 0.06804 atmosphere 51.7 millimeters of mercury at 32° F. 1 Foot of Water at 62° F. = 0.433 pound per square inch 62.355 pounds per square foot 0.883 inch of mercury at 62° F. 821.2 feet of air at 62° F. and barometer 29.92 1 Inch of Water 62° F. = 0.0361 pound per square inch 5.196 pounds per square foot 0.5776 ounce per square inch 0.0736 inch of mercury at 62° F. 68.44 feet of air at 62° F. and barometer 29.92 1 Foot of Air at 32° F. and Barometer 29.92 = 0.0761 pound per square foot 0.0146 inch of water at 62° F. 1 Inch of Mercury at 62° F. = 0.4912 pound per square inch 1.132 feet of water at 62° F. 13.58 inches of water at 62° F. 1 Atmosphere = 760.0 millimeters of mercury at 32° F. 14.7 pounds per square inch 29.921 inches of mercury at 32° F. 2,116.0 pounds per square foot 1.033 kilograms per square centi- meter 1 Millimeter = 0.03937 inch 1 Centimeter = 0.3937 inch 1 Meter = 39.37 inches 1 Meter = 32808 feet 1 Square Meter = 10.764 square feet 1 Liter = 61.023 cubic inches 0.264 U. S. gallons 1 Gram = 1 cubic centimeter of distilled water 15.43 grains troy 0.0353 ounce 1 Kilogram = 2.20462 pounds avoirdupois APPENDIX G. RULES FOR FIREMEN USING ILLINOIS AND INDIANA COAL IN HAND- FIRED FURNACES. (Formulated by the Coal Stoking and Anti-Smoke Committee of the Illinois Coal Operator's Association.) 1. Break all lumps and do not throw any in furnace any larger than one's fist. The reason for this is, that large lumps do not ignite promptly and their presence also causes holes to form in the fire, which allow the passage of too much air. 2. Keep the ash pits bright at all times. If they become dark it is evident that the fire is getting dirty and needs cleaning, which, if not done, will cause imperfect combustion and smoke. If the furnace is equipped with a shaking grate, it should be operated often enough to prevent any accumulation of ashes in the fire. Do not allow ashes to collect in the ash pits, as they not only shut off the air supply, but may cause the grate to be burned. 3. In firing do not land the coal all in one heap, but spread it over as wide a space as possible as it leaves the shovel. A little practice will enable one to catch the proper motion to give the shovel to make the coal spread properly. 4. Place the fresh coal from the bridge wall forward to the dead plate and do not add more than 3 or 4 shovels at a charge. If this amount makes smoke it should be reduced till smoke ceases, which means, of course, that firing will be at more frequent intervals than formerly to keep up steam. This rule applies in cases where the boiler is worked at a large capacity. In such instances, however, where a small capacity only is required, firing by the coking method is the best, wherein the fresh coal is placed at the front of the fire and pushed back and leveled when it has become coked. 5. Fire one side of the furnace at a time so that the other side contain- ing a bright fire will ignite the volatile gases from the fresh charge. 6. Do not allow the fire to burn down dull before charging. If this is done, it will not only result in a smoky chimney, but an irregular steam pressure. 7. Do not allow holes to form in the fire. Should one form, fill it by leveling and not by a scoop full of coal. Keep the fire even and level at all times. As far as possible level the fire after the coal has become coked. 879 880 STEAM POWER PLANT ENGINEERING 8. Carry as thick a fire as the draft will allow, but in deciding on the proper thickness, judgment must be exercised. If the draft is poor a thin fire will be in order, but if strong, a thicker fire should be carried. 9. Regulate the draft by the bottom or ash pit doors and not by the stack dampers, because when the stack damper is used it tends to pro- duce a smoky chimney, as it reduces the draft, while the closing of the ash pit door diminishes the capacity to burn coal. If strict attention is given to firing, and accounting to demand, for steam, there will be no occasion to have recourse to dampers, except when there is a sudden interruption in the amount of steam being used. 10. A good general rule is to fire little and often, according to steam demands, rather than heavy and seldom. The former means economy in fuel and a clean chimney, while the latter signifies extravagance in fuel and a smoky chimney. APPENDIX H. MOLLIER'S DIAGRAM. The steam tables give values of the simultaneous physical properties of steam, such as pressure, entropy, temperature, etc. When certain of these properties are known the remainder can be obtained from the tables. The simultaneous properties can also be shown by means of a diagram each point on which represents steam in a perfectly definite condition. Fig. 462 gives a skeleton outline of such a diagram and Fig. 463 a H G. 462. B.T.X7.per Pound of Steam ' ' 2 reduced reproduction of the complete chart as ordinarily constructed. Referring to Fig. 462, abscissas represent the heat contents or B.T.U. per pound of steam and ordinates represent the total entropy. Vertical lines then represent lines of constant heat content, and horizontal lines constant entropy. P 1 P 1 and P 2 P 2 represent lines of constant pressure and X X X 1 and X 2 X 2 lines of constant quality. Evidently any point in the chart represents a fixed condition of heat content, pressure, quality and entropy as determined by its location with respect to the different lines. Thus point 1 represents a pressure P t as determined by the numerical value of line P 1 P l , quality x x by its location on line X X X U 881 882 STEAM POWER PLANT ENGINEERING entropy n t by its projection N t on the Y axis, and heat content H l by its projection on the X axis. The principal advantages of a total heat-entropy diagram over the tables are that they give the properties of wet and superheated steam and offer a simple means of solving many problems without calculations. For example, the chart offers a ready solution of problems involving (a) Adiabatic expansion. (b) Throttling. (c) Expansion with frictional resistances. (a) Adiabatic Expansion: From thermodynamics we know that during an adiabatic change the entropy is constant; thus, in expanding from pressure P 1 and condition represented in point 1 to a lower pressure P 2 it is only necessary to find the intersection 2 of a horizontal line from point 1 with line P 2 P 2 . The various properties corresponding to point 2 can be read directly from the diagram. The line 1-2 = H l H 2 represents the difference in heat content following adiabatic expansion from pressure P t and condition 1 to pressure P, or line #i# 2 =H t — H 2 = x x r x + q t - x 2 r 2 - q 2 . The quality x 2 is read directly from the intersection of line 1-2 with the constant quality line X 2 X 2 . The entropy n 2 , of course, remains the same. From equation (73), p. we find that the velocity due to adiabatic expansion is V = 223.9 V H t - H Mollier has added along the margin of the diagram (Fig. 463) a scale of velocity so that V may be ascertained by laying off the length H 1 H 2 on the scale. Example: Steam at 120 pounds absolute, quality 0.98, expands adia- batically to a back pressure of 2 pounds absolute. Find the quality and heat content at the lower pressure. From Fig. 462 we locate P x at the intersection of pressure curve 120 and quality curve 0.98. The corresponding values of H i and n x are found by interpolation to be 1174.7 and 1.564 repectively. Follow horizontal line 1.564 until it intersects pressure line P 2 . The corre- sponding values of H 2 and x 2 are found to be 910 and 0.797 respectively. The horizontal intercept between the two pressure lines laid off on the velocity diagram gives V = 3640 feet per second. Supposing the steam to be superheated 200 degrees instead of being wet, find the quality and heat content at the end of expansion. — I :— I E— i —I 884 STEAM POWER PLANT ENGINEERING Locate P x at the intersection of pressure curve 120 and superheat curve 200. The corresponding values of H 1 and n l are found to be 1295 and 1.703 respectively. Follow horizontal line 1.703 until it inter- sects pressure line P 2 . The corresponding values of H 2 and x 2 are found to be 990 and 877 respectively. (b) Throttling: If steam expands through a small orifice without the addition or abstraction of heat and is brought finally to its initial con- dition its total heat will be unchanged. This process is called throttling and occurs when steam passes through a reducing valve. Vertical lines in Figs. 462 and 463 are lines of constant total heat and consequently show the changes in the condition of steam which result from throttling. Thus in throttling steam from pressure P 1} Fig. 462, to P 2 it is only necessary to find the intersection, 4, of a vertical line from point 1 with line P 2 P 2 . Example: Steam at 200 pounds pressure and quality 0.96 passes through a reducing valve and its pressure is lowered to 15 pounds. Find its quality at the lower pressure. The intersection of pressure line 190 with quality line 0.96 gives H x = 1165. Follow vertical line 1165 until it intersects pressure line 15. The corresponding value for x 2 is found to be 30, that is, the steam is superheated 30 degrees. To what pressure must the steam be reduced in order that it may be dry and saturated? Follow vertical line 1165 until it intersects the saturation curve. The corresponding pressure is found to be 30 pounds. (c) Expansion Involving Frictional Resistances: As steam expands in the nozzle of a turbine or passes between the vanes it experiences fric- tional resistances which cause it to give up less energy than it would under ideal conditions. The work of friction causes the entropy of the steam at its lowest temperature to be greater than it would be if adia- batic expansion occurred and serves to increase its dryness fraction. If y one hundredths of the heat H 1 — H 2 (given up in adiabatic expansion) is lost due to friction, the heat available for useful work is {l-y)(H x -H 2 ), the resulting velocity of the jet is V = 223.9 V(l -y) (H x - H 2 ), and the increase in quality of the exhaust steam is y(H t -H 2 ) m These equations may be readily solved by means of the diagram. Referring to Fig. 462 line 1 3 represents an expansion from pressure P x to pressure P 2 with frictional resistances. APPENDIX H 885 From the diagram y = N 2 Hfl 2 12 HM, (1 - y) (H x - H 2 ) = line 1 N = H X H Z . Increase in quality = — = distance 2 3 between X 2 X 2 and x s x 3 . Increase in entropy = NS = N t N 3 . Example: Steam at 160 pounds absolute initial pressure, quality 0.97, expands through a nozzle to a back pressure of 2 pounds absolute. If 15 per cent of the heat energy is lost in friction, find the quality of the steam at the lower pressure and the velocity of the jet. From Fig. 462 we locate P t at the intersection of pressure curve 160 and quality curve 0.97. The corresponding value of H t is 1170. Follow line 1170 horizontally until it intersects pressure line P 2 . From the diagram we find for adiabatic expansion # 2 = 910 x 2 = 0.797. But the friction increases the heat content at the end of expansion an amount 0.15 xH t - H 2 = 0.15 (1170 - 910) = 39, so that the final heat con- tent = 910 + 39 - 949. Follow pressure line P 2 until it intersects heat line 949. The quality x 2 is found to be 0.836 and the entropy n 2 = 1.632. From the velocity scale we find V = 3320 feet per second for H x - H 2 = (1170 - 949). INDEX Absorption dynamometer, 857. Acetylene, properties of, 26. Acidity, tests for, in oils, 674. Acme bucket trap, 590. Acton atmospheric relief valve, 666. Ados C0 2 recorder, 743. Aero-pulverizer powdered coal burner, 50. Air chambers, 528. Air-cooled surface condensers, 428. Air lift, 572. Air, properties of, 26. Air pumps, 552-560. size of dry, 558. size of wet, 553. Air required for operating air lift, 573. Air required for combustion, 28. Air spaces, grate bars, 114. Air supply above grate, 151. Air thermometers, recording, Air vs. steam as an oil atomizer, 60. Alarm, high and low water, 120. Alberger barometric condenser, 412. cooling tower, 459. rotative dry air pumps, 557. Alden absorption dynamometer, 857. Allis-Chalmers steam turbine, 372. Alternate method of starting and stopping boiler tests, 826. American underfeed stoker, 139. Analyses of boiler scales, 473. of flue gases, of fuel oils, 52. of typical American coals, 23. of waters for boiler feeding, 473. Anchors, pipe, 622. Anderson automatic non-return valve, 658. feed-water pumping system, 482. triple-duty emergency valve, 659. Animal fats and oils, 669. Anthracite coals, 15. Aqueous vapor, pressure of, 399. effect of, on degree of vacuum, 405. Armour Glue Works, vacuum ash system at, 198. Armour Institute, brick chimney at, 227. Arndt's econometer, 742. Ash bins, 182. Ash conveyor, vacuum system, 196. Ash, influence of, on fuel value of dry coal, 40. Ash-handling systems, 181-204. Ash, treatment of, in boiler tests, A.S.M.E. rules for conducting boiler trials, 822-845. A.S.M.E. rules for conducting engine tests, 846-872. Atmospheric heaters, 486. Atmospheric surface lubrication, 675. Atmospheric relief valves, 666. Augmenter, Parsons vacuum, 444. Aurora and Elgin Interurban Ry., coal- handling system, 194. Austin steam separator, 579. Automatic cut-off vs. throttling engines, 310. Automatic injectors, 547. Automatic non-return valves, 658. Automatic temperature control, 646. Auxiliaries, power consumption of con- denser, 449. Auxiliaries, measurement of steam used by, 851. Babcock & Wilcox boilers, 81. chain grate, 128. superheater, 163. Back connection, return tubular boileF, 79. Back pressure on engines, 283. Back-pressure valves, 665. Baffle-plate steam separator, 579. 887 888 INDEX Bagasse as fuel, 20. Balanced-draft system, 264. Baragwanath feed-water heater, 495. siphon condenser, 408. surface condenser, 416. Barnard- Wheeler cooling tower, 457. Basement plan, West Albany station, N. Y. C. R. R., 720. Bearings, lubrication of, 675. Belliss engines, tests of, with superheated steam, 314. Belt conveyors, 192. Bends, pipe, 619. Bibliography : Cost of electric power, 724. Cost of gas power, 726. Cost of steam power, 727. Cost of water power, 728. Description of gas-driven power plants, 798. Description of central stations, steam engines, 802-818. Description of central stations, steam turbines, 808, 819. Description of hydraulic power plants, 798, 818. Description of isolated station, 809. apartment buildings, 809. manufacturing plants, 810. office buildings, 812. stores, 814. Design of power plants, Binary- vapor engines, 321. Bituminous coals, 16. Blades, arrangement of, in steam tur- bines, 352, 367. Blake jet condenser, 403. Blast furnace gas, properties of, 67. Bloomsburg steam jet, 246. Blowers, fan, 249. tests of, 257. steam jet, 245. Blow-off piping, South Side Elevated R.R., 662. Blow-offs, 116. Blow-off tank, 117. Blow-off valve, 661. Boiler compounds, 476. Boiler-feed pumps {see Pumps). Boiler room area, 86. Boiler tests, A.S.M.E. code, 822-845. Boiler tests, discrepancy between com- mercial and experimental results, 842. Boilers, 66-123. Babcock & Wilcox, chain-grate, 128. Babcock & Wilcox, hand-fired,- 80. capacity of, 104. classification of, 68. cost of, 112. efficiency of, 98, 762. fire-box, 71. furnaces for, 68-88, 124-151. grates for, 114, 126. heating surface of, 92. Heine, 83. horizontal return tubular, 74. horse power of, 93. inspector's report (1907), Hartford Boiler Insurance Company, 474. Manning vertical, 70. Parker boilers, 85. performances of, 99. Robb-Mumford, 73. Scotch-marine, 72. selection of type, 112. settings for, 68-90, 126-140. specifications for, 754. Stirling, 87. vertical tubular, 69. Wickes, 84. Booth fuel oil burner, 56. Boston Elevated, cost of operation, 710. Brake horse power, 855. Brake, rope, 856. Branch fuel oil burner, 57. Brass pipes, 608. Breeching, 240. Brick chimneys, 224. Bristol recording air thermometers, 737. thermo-electric pyrometer, 737. Bucket conveyor, 184. Bucket traps, 590. Buckeye skimmer, 118. Bundy steam separator, 579. Bunkers, coal, 182. Burgeon, specific heat of superheated steam, 157. Burke's smokeless furnace, 151. Burners, oil, 53-61. powdered coal, 44-51. Burnham steam meter, 734. . Burning point, oils, 674. INDEX 889 Bursting strength of pipes, 607. By-pass system of piping, 625. Calorific value of coals, 31. Calorimeters, fuel, 747. Calorimeters, steam, 745. Cannel-coal gas, properties of, 67. Capacity, effect of, on boiler efficiency, 104. Carbon dioxide, properties of, 26. percentage of, in flue gases, 30. Carbon monoxide, heat losses due to formation of, 36. properties of, 26. Carbu retted water gas, properties of, 67. Carnot cycle, 267. Carpenter separating calorimeter, 745. Cast-iron pipes, 607. Central condensing systems, 439. Central hydrostatic cylinder lubricator, 684. Centrifugal oilers, 678. Centrifugal pumps, 560-567. characteristics of, 565. performance of, 563. tables of sizes, 566, 567, types of, 560. Centrifugal steam separators, 578. Chain grates, 126. Chattanooga Electric Company, test of spray fountain, 455. Check valves, 660. Chemical purification of feed water, 476. Chicago setting, hand-fired furnace, 143. Chimney at Armour Institute, 227. Chimney draft, 207. Chimney draft, table of, for various temperatures, 210. Chimney, efficiency of, 241. height of, for burning fuel oil, 218. test of 100-foot steel, 213. Chimney vs. mechanical draft, 261. Chimneys, 207-244. brick, 224. classification of, 218. core and lining for, 230. cost of, 243. Custodis radial brick, 225, 234. dimensions of, 216, 244. formulas for, 212. foundations for, 240. guyed steel, 219. Chimneys, materials for brick, 230. self-sustaining steel, 219. stability of brick, 231. stability of steel, 223. steel, 219. strain sheet for reenforced concrete, 234. thickness of walls, brick, 226. thickness of shell, steel, 220. Weber reenforced concrete, 235. Cincinnati Traction Company, coal con- veyors, 195. Circulating pumps, 572. - Classification of boilers, 68. chimneys, 218. condensers, 400. feed-water heaters, 485. fuel-oil burners, 53. fuels, 14. lubricating oils, 669. powdered-coal burners, 44. pumps, 522. steam separators, 576. steam traps, 588. steam turbines, 327. stokers, 126. testing instruments, 730-749. Clearance volume, influence of r on engine economy, 281. Coal, 15. analysis of, for boiler tests, 829. anthracite, 15. bituminous, 16. calorific value of, 31. composition of, 23. measurements of, boiler tests, 852. powdered, 44-51. proximate analysis, 31. purchasing, 42. sampling, 828. specifications for purchasing, 769. storage of, 181. ultimate analysis, 31. washed, 40. Coal and ash handling, 181-204. Coal bunkers, 182. Coal gas, properties of, 67. Coal fields of the United States (ref.), 16. Coal hoppers, 202. Coal valves, 205. Cochrane heater, 487. 890 INDEX Coefficient of expansion, pipe materials, 620. Coke-oven gas, properties of, 67. Cold test, oils, 674. Columbia expansion trap, 592. Combustion, 24. Commercial National Bank Building, Chicago, ash system, 192. Commonwealth Edison Company, Fisk Street Station, 774-787. Compressed-air oiling system, 681. Compressed air, power required, air lift, 572. Compression, effect of, on engine economy, 283. Compound engines, 300. Compounds, boiler, 476. Condensers, 397-469. air for cooling purposes, surface, 429. Alberger barometric, 412. Baragwanath siphon, 408. barometric, 411. Blake jet, 403. choice of, 450. classification of, 400. cooling water for, 404. cost of, 450. counter-current, 411. dry tube, 424. economical vacuum for, 451. ejector, 410. extent of cooling surface for, 421. function of, 398. high-vacuum, 441. independent, 434. injection orifice, 407. jet, 401. Korting multi-jet, 446. location of, 433. multi-flow surface, 419. Schutte ejector, 410. siphon, 408. sizes of siphon, 409. specifications for, surface, dry air-cooled, 428. surface, evaporative, 433. surface, water-cooled, 416. tests of surface, 434. Tomlinson barometric, 415. volume of condenser chamber for jet, 408. Condensers, Weighton multi-flow, 419. Weiss barometric, 411. Westinghouse-Leblanc, 445. Wheeler admiralty, 417. Worthington barometric, 414. Condensers, Worthington jet, 402. Concrete chimneys, 234-240. Condensation and leakage losses in en- gines, 279. Condensing, influence on engine econ- omy, 307. Condensing plant, elementary, 7. Condensing plant with full complement of heat-saving appliances, 10. Conoidal pump, test of, 568. Conversion tables, 878. Conveyors, 183-197. Cooling ponds, 454. Cooling towers, 456. Cooling towers, fan vs. natural draft, 460. Copper pipes, 607. Corliss engine, 269. Correction factors for steam turbines, 387. Cost of boilers and settings, 112. chimneys, 243. condensers, 450. engines, 326. evaporating water, 105. handling coal and ashes, 201. mechanical draft systems, 263. pipe flanges, 614. power (see power costs). stokers, 151. turbines, 392. Costs, operating, 693. Coverings for steam pipes, 636. Crusher and cross conveyor, 190. Curtis steam turbine, 350. Curve load factor, 691. Custodis radial brick chimney, 234. Cut-off, commercial, 860. Cut-off, point of, 861. Cylinder condensation, 279. Cylinder cups, 682. Cylinder lubrication, 682. Cylinder ratios, compound engines, 300. Damper regulators, 118. Davis back-pressure valve, 665. Dean air pump, 552. INDEX 891 De Laval centrifugal pump, 568. steam turbine, 331. Density of air and flue gas, 209. Depreciation of powdered-coal furnace,44. Depreciation, rate of, Depreciation percentages, Chicago Trac- tion Valuation Commission, 696. Desmond injector, test of, 549. Detroit Edison Co., coal-handling system, 196. Diagram factor, steam engine, 862. Diaphragm valve, 647. Differential traps, 594. "Direct" steam separator, 580. Disk water meter, 732. Divergent nozzle, design of, 334. Dodge, A. R., specific heat of super- heated steam, 156. Double-deck turbine installation, 381. Double-flow steam turbine, 369. Double stoker, 135. Down-draft furnace, 139. Draft, balanced, 264. chimney, 207. for powdered-coal burner, 47. forced, 251. gauges, 735. influence of, on boiler efficiency, 106. induced, 252. mechanical, 245-266. Drainage of jackets and receivers, 597. Drains, office building, 603. Drips, 586. high-pressure, 588. low-pressure, 586. removal of oil from, 587. under alternate pressure and vacuum, 599. under vacuum, 598. Dry-air pumps, 537. Dry-air surface condensers, 428. Dry docks, centrifugal pump character- istics for, 565. Dry tube surface condensers, 424. Dulong's formula, 32. Dunham steam trap, 593. Duplex coal valve, 204. Duplex steam pump, 524. Duplicate piping system, 624. Dutch oven, 141. Duty, pump, 536. Economizers, 508. factors for determining installation of, 532. Green, 510. heat transmission in, 512. tests of, 515. Edwards air pump, 555. Efficiencies of boilers and grates, 98. boilers with oil fuel, 54. Efficiencies of boiler-feed pumps, 534. centrifugal pumps, 565-568. compound engines, saturated steam, 306. compound engines, superheated steam, 317, 321. fans, 257. piston pumps, 533. simple engines, saturated steam, 296. simple engines, superheated steam, 316. steam turbines, 384. triple-expansion engines, superheated steam, 318. triplex pumps, 544, 545. Efficiency, air lift, 574. Carnot cycle, 267. condensing plants, 9, 13. furnace, 99. grate, 99. mechanical, 275. non-condensing plants, 4. Rankine cycle, 268. thermal, 273. Ejector condenser, 410, 446. Ejector, Shone, 603. Electrical power, cost of, 704-729. Elementary condensing plants, 2. Elementary non-condensing plant, 7. Elementary theory, Curtis turbines, 358. De Laval turbine, 333. Westinghouse-Parsons turbine, 373. Elevating tower, cable-car distribution, 193. Elevating tower, hand-car distribution, 196. Ellison's universal steam calorimeter, 746. Emergency valves, 658. Engines (steam), 267-326. A.S.M.E. code for testing, 846-872. automatic cut-off, 310. back pressure, effect of, on economy, 283. 892 INDEX Engines (steam), binary vapor, 321. clearance volume, 281. compound, 300. compression, effect of, on economy, 283. condensing, effect of, on economy, 307. cost of, 326. cylinder condensation, 279. economy of (see Tests). efficiencies of (see Efficiencies). friction of, 284. heat losses in, 278. high-speed, 290. ideal, 267. incomplete expansion, loss due to, 282. increasing initial pressure, effect of, 286. jackets, influence of, 170. leakage losses in, 279. low-speed, 299. mechanical efficiency, 275. non-condensing, test of, 296. receiver-reheaters, economy of, 287. simple, 291. single-acting, 290. specifications, 750. sulzer, 319. superheated steam, 313. tests of (see Tests). thermal efficiency of, 273. throttling vs. automatic cut-off, 310. triple and quadruple expansion, 305. wire drawing, effect of, 284. with low-pressure turbines, 376. Entropy diagram, 863, 881. Equation of pipes, 640. Exhaust heads, 585. Exhaust piping, 642. Expansion traps, 592. Expansion of pipe materials, 618. Expansion, ratio of, 862. Extended boiler front setting, Evaporation, cooling pond, 454. cost of, coal fuel, 105. cost of, oil vs. coal, 55. from and at 212° F. per square foot per hour, 95. rate of, boilers, 96. rates of, in still air, 454. unit of, 88. Evaporative surface condenser, 433. Factor of evaporation, 88. Fan draft, 249. Fans, capacity of induced-draft, 262. capacity of forced-draft, 261. performance of, 258. theory of, 252. Feed water, analyses of, 473. Feed-water heaters (see Heaters). Feed-water heating system, choice of, 516. Feed-water piping, 647. Feed- water purification, 471. Feed-water regulators, 542. Fery radiation pyrometer, 740. Filters, oil, 688. Fire-box, boilers, 71. Fire-tile "Economy," 131. Fire-tile combustion chamber, 128. Fire, thickness of, 110. Fire-tube boiler, 71. First National Bank Building, Chicago, power costs, 710. Fisher pump governor, 541. Fittings, pipe, 610. Fixed carbon in coal 18. Fixed charges, 693. Flanged fittings, 610. Flanges, table of extra heavy, 615. Flanges, table of standard, 614. Flap coal valve, 205. Flash point, oil testing, 673. Flemming four-valve engine, 320. Flinn trap, 595. Float trap, 589. Floor space, turbine vs. reciprocating engine, 383. Flow of steam in pipes, 632. steam through nozzles, 336. water through pipes, 650. Flue-gas analysis, 29, 741. Flue-gas apparatus, Ados recording, 743. Arndt's econometer, indicating, 742. Orsat, 741. Sarco recording, 744. Flush-front boiler setting, Fly-wheel pumps, 523. Foot valves, 668. Forced draft, 245-266. Forced-feed lubricator, 684. Forcing capacity of boilers, 109. Foster back-pressure valve, 665. pressure regulator, 667. INDEX 893 Foster superheater, 165. Foundations, chimney, 240. Fountain, spray, 455. Four- valve engines, tests of, 298, 306. Friction of engines, 284. Friction of water in pipes, 652. Friction tests of oil, 674. Friction through valves and fittings, 639, 653. Fuel, cost of, 700. Fuel calorimeters, 747. Fuel oil, 51-66. Fuel-oil burners (see Burners). Fuels and combustion, 14-16. Fuels, classification of, 14. Function of the condenser, 398. Furnace arch bars, 79. Furnace efficiency, 99. Furnace temperature, influence on boiler efficiency, 111. Furnace influence on gas composition, 36. Furnace for burning oil fuel, 59. Furnace for burning powdered coal, 47. Furnace, smokeless (see Smokeless fur- naces). Fusible plugs, 121. Gaseous fuels, 66. characteristics of, 67. Gauge cocks, 3, 119. Gauges, water, 119. Gate valves, 655. Geipel steam trap, 593. Globe valves, 655. Government specifications for purchasing coal, 769. Governor, steam pump, 541. Goubert feed-water heater, 492. Grate, loss of fuel through, 37. Grate surface, 95, 98. Grate bars, thickness of, 114. Grates, chain, 126. rocking, 116. stationary, 114. Gravity oil feed, 680. Gravity, Baume oils, 672. Gravity, specific oils, 672. Grease extractor, 584. Greases, 671. Green chain grate, 126. economizer, 510. Guyed steel chimneys, 219. Hamilton-Holzworth turbine, 362. Hamler-Eddy Smoke Recorder, Hammel fuel oil burner, 57. Hancock injector, 546. Hand shoveling, 183. Hangers, for pipes, 622. Hartford Boiler Insurance Company, annual report (1907), 474. Hartford boiler specifications, 754. Hawley down-draft furnace, 139. Headers, main steam, 563. Heat balance, boiler tests, 832. Heat distribution, condensing plants, 9, 13. Heat distribution, non-condensing plants, 4. Heat losses in burning coal, 32. in the chimney gases, 33. in steam engines, 278. Heat transmission, boilers, 90, closed heaters, 497. economizers, 512. influence of scale on, 472. superheaters, 170. Heating surface, boilers, 92, 95. Heaters, feed- water, 471-521. Baragwanath, 595. choice of, 516. classification of, 485. closed, 486. Cochrane, 487. counter-current, 485. flue-gas, 485. Goubert, 492. Harrisburg, 494. Hoppe's, 489. induced, 486, 506. live steam, 485-507. open, 486, 504. Otis, 494. parallel-current, 492. primary, 486. secondary, 486. single-flow, 492. steam tube, 494. through, 486, 505. vacuum, 485, 506. Wainwright, 493. Webster, 488. Heater and purifier combined, 489. Heine boiler, 83. 894 INDEX Heinrich smokeless furnace, 249. Heintz expansion trap, 594. Herringbone grate, 115. Hewes and Phillips air pump, 556. Heyworth Building, Chicago, plan of piping, 421. High and low speed engines, 290. High-pressure drips, 588. High-speed double- valve engine, 297. High-speed single- valve engine, 290. Hollow bridge wal, 151, 247. Holly loop, 602. Hoppe's feed-water heater, 489. steam separators, 577. Horizontal return tubular boilers, 74. Horse power of boilers, 93. Hot-well temperatures, surface con- densers, 426. Hot-well pumps, 560. Hunt coal conveyor, 189. Hydraulic packing, 530. Hydraulic Oil Storage Company's fuel oil system, 64. Hydraulic valve gear, Curtis turbine, 355. Hydrogen, properties of, 26. losses due to, 38. Hydrometer, Baume, oils, 672. Hydrostatic cylinder lubricator, 683 -Hygrometry, 468. Ideal engine, 267. Illinois Engineering Company's auto- matic vacuum valve, 643. Impulse turbine, 331. Incomplete combustion, loss due to, 36. Incomplete expansion, loss due to, 282. Increasing boiler pressure, economy of, 286. Increasing rotative speed, economy of, 290. Increasing degree of vacuum, cost of, 451 . Increasing degree of vacuum, economy of, 308, 395. Identification of oils, 672. Independent condensers, 434. Independently fired superheaters, 166. Indicated horse power, 853. Indicator cards, air pump, 560. Indicator cards, analysis of, 859. automatic cut-off engines, 311. four- valve engine, 861. Indicator cards, throttling engine, 311. Westinghouse-Parsons turbine, 365. Indicator springs, tests of, 854. Induced draft, 252. Induced heaters, 506. Initial condensation, 279. Injection orifice, 408. Injectors, 545-548. performance of, 548. range in working pressures, 350. vs. steam as boiler feeders, 550. Intermittent oiling, 675. International Gas Company's fuel oil system, 63. Interest charges, 693. Isolated stations, cost of power in, 71 6-723. Isolated stations, influence of load factor on economy, 720. Jackets, influence of, 170. Jackets, methods of draining, 597. Jet condensers, Jet, Bloomsburg, 246. Jet, ring steam, 246. Jets, steam consumption of, 248. Jones underfeed stoker, 72. Kent's wing-wall furnace, 149. Kerosene, use of, in boilers, 477. Kerr steam turbine, 346. Keystone separator, 578. Kieley reducing valve, 667. Kindling temperatures, 25. Kirkwood oil burner, 58. Kitts feed-water regulator, 542. Kitts hydraulic damper regulator, 53. Knoblauch and Linde, specific heat of superheated steam, 156. Knowles triplex pump, test of, 544. Korting fuel-oil burner, 56. Korting multi-jet condenser, 446. Labor, cost of, in power plants, 699. cost of, in street railway plants, 701. cost of, in tall office buildings, 702. Lea-Degan three-stage turbine pump, 562, 569. Leakage of steam in engines, 279. Leyland automatic cylinder cup, 683. Life of power plant appliances, 694. Lignite, 18. Limit of superheat, 154. INDEX 895 Link Belt Company coal-handling sys- tem, 184. Live-steam feed-water heaters, 507. Load factor, 691. influence of, on cost of power, 692. Location of condensers, 433. of separators, 580. of traps, 596. Loew grease extractor, 584. Loop header, 624. Loop, Holly, 602. Loop, steam, 600. Loss of heat from bare pipes, 615. Loss of heat from covered pipes, 616. Losses in burning fuel, 32-40. Losses in steam engines, 278. Low-pressure drips, 586. Low-pressure turbines, 376. Low-speed engines, 299. Lubricants, 669-680. Lubrication, atmospheric, 675. central hydrostatic system, 684. compressed-air feed, 680. cylinder lubrication, 682. forced system, 684. gravity systems, 680. Siegrist system, 685. Lubricating oils, classification of, 670. properties of, 676. specific gravity of, 672. Lubricators, hydrostatic, 683. Ludlow angle valve, 667. Lunkenheimer sight-feed lubricator, 684. Mahler bomb calorimeter, 747. Mains, steam, 622. Maintenance, 699. Manning vertical boilers, 70. Marks' and Davis' steam tables, Marsh gas, properties of, 26. Marsh steam pump, 527. Marsh steam pump, test of, 534. Materials for brick chimneys, 230. Materials for pipes and fittings, 606. Materials for superheaters, 170. McClaves argand blower, 246. McDaniel float trap, 589. Mean temperature difference, heaters, 497. Measurement of heat units consumed by engines, 849. Measurement of feed-water consumption, engine, 849. Measurement of steam used by auxiliaries, 850. Mechanical draft, 245-266. Mechanical efficiency of engines, 275. of pumps, 533. Mechanical boiler-tube cleaner, 121. Mechanical purification of feed water, 481. Mechanical stokers, 125. Mechanical valve gear, Curtis turbine, 354. Medium and low speed engines, 299. Mesh steam separators, 580., Meters, steam, 734. Meters, water, 732. Meyers Bagasse furnace, 21. Meyers tan bark furnace, 22. Mineral oils, 670. Mixed pressure turbines, 376. Moisture in air, loss due to, in combustion, 37. in fuel, loss due to, 37. in steam, determination of, 827. in steam, effect on engine economy, 278. in steam evaporated by throttling, 312. Mollier diagram, Mullan valveless air pump, 556. Murgue's theory, centrifugal fans, 254. Murphy furnace, 137. Napier's rule for the flow of steam, 336. Natural-draft cooling tower, 460. Natural gas, properties of, 67. Naval Liquid Fuel Board, report of United States, 65. Nitrogen, properties of, 26. Non-condensing engines, test of, 296, 315. Non-condensing plants, arrangement of, 2,5. exhaust piping in, Paul system, 643. exhaust piping in, Webster system, 542. feed- water piping in, 647. open heater in, 506. Non-return valves, 658. Norfolk Traction Co.'s ash-hananiig sys- tem, 200. Northwestern Elevated R. R. power house, condenser piping, 446. Nozzles, De Laval steam turbine, 332. 896 INDEX ' Nozzles, flow of steam through, 334. flow of water through, 650. Kerr steam turbine, 346. theoretical design of divergent, 334. Nugent telescopic oiler, 677. Office buildings, cost of power in, 703- 720. Oil bath, 675. Oil burners (see Burners). cups, 677. eliminators, 581. filters, 687. fuels, analysis of, 31. pressure in fuel oil systems, 63. separators, 581. storage, 64. Oil, waste, and supplies, cost of, in power plants, 703. Oiler, centrifugal, 678. gravity, 679. pendulum, 678. ring, 678. telescope, 677. Oiling systems (see Lubricating systems). Oils, animal, 669. distilled, 670. identification of, 672. mineral, 670. properties of, 671. specific gravity of, 672. specifications for, 672. tests for, 672. Olefiant gas, properties of, 26. Open heaters, 486. Open heater vs. closed heater, 504. Operating costs, Operating costs, reciprocating engine vs. steam turbine, 392. Optical pyrometers, 738. Orsat apparatus, 741. Orifice, size of injection, 408. Orifices, flow of steam through, 336. flow of water through, 584. Otis feed-water heater, 428. Overhead storage, bucket hoist, 195. Overload capacity, steam turbines, 364. Oxygen, properties of, 26. Pan surface, open heaters, 491. Parallel-current condenser, 401. Parallel-current feed-water heater, 492. Parker boiler, 85. Parr coal calorimeter, 748. Parsons smokeless furnace, 248. Parsons vacuum augmenter, 444. Paul exhauster, 645. Paul heating system, 643. Peat, 18. Penberthy injector, 546. Pendulum oiler, 678. Pennel saturated-air surface condenser, 431. Pinther powdered-coal burner, 49. Pipe anchors, 622. Pipe bends, 619. Pipe, brass, 608. cast-iron, 607. cast-steel, 607. copper, 608. mild steel, 607. sizes of standard, 611. Pipe flanges, 610. size of extra heavy, 615. sizes of standard, 614. Pipe hangers, 622. Pipe supports, 621. Pipe threads, United States standard, 615. Pipes, equation of, 636. flow of steam in, 632. flow of water in, 650. size of, for low-pressure drips, 588. strength of, 609. Piping, by-pass system of steam, 625. Commonwealth Edison Company, Fisk Street Station, 774-787. condenser, 642. Des Moines City Railway Company, 631. duplicate system of, 624. feed-water, 648. Heyworth Building, Chicago, 626. loop header system of, 624. Manhattan Elevated, New York, 627. Paul heating system, 643. Princeton University, 623. ' specifications for, 760. steam, 622. Webster heating system, 641. West Albany Station, New York Cen- tral Railway Company, 788. Pistons, water, 530. Pitot tubes, 253. INDEX 897 Plungers, pump, 530. Ponds, cooling, 454. Pop safety valves, 664. Positive injectors, 547. Powdered coal, 44. Powdered-coal burners (see Burners). Power consumption of condenser auxil- iaries, 449. Power cost, Boston Elevated, 710. British electric light and power plants, 709. compound engine plants, 711. depreciation, 694. First National Bank Building, Chi- cago, 710-724. fixed charges, 693. fuel, 700. insurance, 699. interest, 699. isolated stations, 723. labor, 699. maintenance, 699. operating charges, 699. simple engine plant, 711. street railway plants, C. C. Moore, 701-705. street railway plants, R. C. Carpenter, 715. street railway plants, typical U.S., 709. taxes, 699. Power measurement, 741. Powers thermostat, 646. Preheating feed water, economy of, 484. Pressure gauges, 735. Pressure of aqueous vapor for different temperatures, 399. Pressure regulation, 667. Producer gas, properties of, 67. Properties of air, 462. of fuel oil, 52. of gases, 57. of lubricating oils, 670, 676. of steam, 873. Proximate analysis, 31. Pulsometer, 572. Pump governors, 541. Pumping engines, surface condensing for, 436. Pumps, air, 552-560. air lift, 572. air, jet condensers, 552. Pumps, air, sizes of, 553. air, surface condensers, 558. air, theoretical work, 560. boiler-feed, 524. classification of, 522. centrifugal, 522, 523, 560-566. circulating, 572. direct-pressure, 528. duplex, 524. duty of, 536. effect of piston speed on economy, 473. fly-wheel, 532, 561. jet, 522, 523, 566. Marsh boiler-feed, 527. multi-stage centrifugal, 561. outside packed plunger, 531. performance of piston, 531. power, 543. rotary, 567. simplex, 537. size of boiler-feed, 539. tests of (see Tests). triplex, 543. turbine, 561. volute, 561. water pistons for, 530. Purchasing coal, 42. government specifications for, 769. Purification, chemical, feed water, 476. mechanical, 479. thermal, 479. Purifiers, live-steam, 507. Purifying plants, Anderson, 583. Scaife, 581. We-Fu-Go, 581. Pyrometers, air-recording, 737. Bristol thermo-electric, 737. Callendar resistance, 738. Fery radiation, 740. Wanner optical, 738. Quality of steam, 827. Qualifications for a good lubricant, 671. Radial brick chimneys, 234. Radiation and minor losses, boilers, 39. Radiation pyrometer, 740. Rankine cycle, 268. Rate of combustion, powdered coal, 45. Rate of combustion, relation of air sup- ply on, 264. 898 INDEX Rate of depreciation, 694. Rate of driving, effect on economy of boilers, 106. Rateau regenerator accumulator, 380. six-stage turbine pump, 562. steam turbine, low-pressure, 376. Ratio of cooling water to condensed steam, 406. Ratio of heating to grate surface, 98. Ratio of expansion, 862. Reaction turbines, 365. Receiver-reheaters, 287. Reciprocating engines vs. steam turbines, 386. Records, power plant, 690. Reducing valves, 396, 462. Reenforced concrete chimneys, 234. Regulation of steam turbines, 353, 368. Regulators, damper, 118. feed- water, 541. Repairs, cost of, power plants, 703. Report of United States Naval Liquid Fuel Board, 65. Restricted feed, lubrication, 675. Returns tank, 603. Ring oiler, 678. Ring steam jet, 246. Ringleman smoke chart, 844. Riveted joints, steel chimneys, 223. Robb-Mumford boiler, 73. Robins belt conveyor, 192. Rochester forced-feed lubricator, 685. Roney stoker, 134. Rope brake, 856. Rotary pumps, 523, 567. Rowe feed-water regulator, 542. Safety valve, dead weight, 663. level, 563. pop, 564. Safety valves, capacity of, 565. rules for loading, 564. Sampling coal, 828. Sarco C0 2 recorder, 744. Saturated-air surface condenser, 430. Saturated-steam tables, 893. Sawdust as fuel, 18. Scaife system of feed-water purification, 481. Scale, analyses of boiler, 473. influence on heat transmission, 472. Schmidt, independently fired superheaters, 167. Schutte ejector condenser, 410. Schwartzkopff powdered-fuel burner, 50. Scotch marine boiler, 72. Screwed fittings, 610. Seaton coal valve, 205. Separating calorimeters, 745. Separators, 575-586. Austin, 579. baffle-plate, 579. Baum, 583. Bundy, 579. centrifugal, 578. classification of, 576. direct, 580. exhaust steam, 581. Hoppes, 577. Keystone, 578. live steam, 576. location of, 580. mesh, 579. oil, tests of, 582. reverse current, 577. Stratton, 578. Settings, smokeless, 124- standard, 68- Shone ejector, 604. Siegrist oiling system, 686. Simple engines, 291. Simplex coal valve, 204. Single vs. double-acting engines, 291. Siphon condensers, 408. Siphon traps, 595. Skimmer, Buckeye, 118. Slip expansion point, 621. Smoke observation, 844. Smoke prevention, 124- Smoke chart, 844. Smoke recorders, 749. Smoke, visible, loss due to, 32. Smokeless furnaces, 124- balanced draft, 264. Burke's, 151. Chicago Smoke Inspection Depart- ments, 143. Dutch oven, 141. fire tile, 128. Heinrich's, 249. Kent's wing wall, 149. Parsons, 248. INDEX 899 Smokeless steam jets, 245-250. stokers, 125-140. *" twin fire-arch, 143. Wooley, 148. Solid lubricants, 671. South Side Elevated Ry. Co., Chicago, chimney at, 221. coal crusher and cross conveyor at, 190. Special furnaces, 141. Specific heat of superheated steam, 155- 162. A. R. Dodge, 156. C. C. Thomas, 161. C. E. Burgoon, 157. Knoblauch and Jakob, 156, 159. value of c p at atmospheric pressure, 157. Specific volume of superheated steam, 160. Specific gravity of lubricating oils, 672. Specifications, boiler, 754. condenser, 758. engine, 750. government, for purchasing coal, 769. piping, 760. Speed, influence on piston pump econ- omy, 2. influence on engine economy, 290. Split bridge wall, 151. Spray fountain, 455. Sprinkling stokers, 141. Stability of brick chimneys, 231. Stability of steel chimneys, 223. Standard method of starting and stop- ping boiler tests, 825. Station load factor, 691. Steam boilers, 68. Steam engines, 267- Steam, flow of, in pipes, 632. Steam jets, 245-250. Steam loop, 600. Steam mains, 629. Steam piping, 606, 668. Steam, properties of, 878. Steam pumps, 522-541. Steam separators, 575-605. Steam, specific heat of superheated, 130- 155. Steam traps, 575-605 Steam turbines, 327. Steel concrete chimneys, 234. Steel chimneys, 219. Step bearing, 356. Stirling boiler, 87. Stirling superheater, 164. St. John's steam meter, 734. Stokers, 125-140. American underfeed, 139. Babcock & Wilcox, 128, 132. chain grates, 126. cost of, 151. Green chain grate, 127. Jones underfeed, 138. Murphy, 137. Roney, 134. Wilkinson, 136. Stop valves, 655. Storage, coal, 181. oil, 64. powdered coal, 45. Stratton separator, 578. Sulphur in fuel, 25, 40. Sulphur dioxide, properties of, 26. Sulzer engine for superheated steam, 319. Superheat, limit of, 154. Superheated steam, 153-180. economy of, steam engine, 313. economy of, steam turbines, 386. properties of, 180. specific heat of, 155. specific volume of, 153, 160. Superheating moisture in air, loss due to, 37. Superheating surface, extent of, 170. Superheaters, 163- Babcock & Wilcox, 163. Foster, 165, 168. independently fired, 166. Schmidt, 167. Schwoerer, 166. Stirling, 164. tests of independently fired vs. flue fired, 176. tests, miscellaneous, 176-179. yearly expense for repairs, cast-iron, 171. Supports, pipe, 621. Surface blow, 117. Surface condensers, 416-430. Surface, cooling, condensers, 421. Surface, heating, feed-water heaters, 496. Surface, heating, superheaters, 170. 900 INDEX Tan bark as fuel, 18. Tank, blow-off, 117. Tank, returns, 602. Taxes, power cost, 693. Telescopic oiler, 677. Temperature-entropy diagram, 864, 881. Temperature of combustion, 27. Temperature measurements, 736. Temperature regulators, 646. Terry steam turbine, 345. Tests of blowers, 257. Tests of boilers: A.S.M.E. code, 822. coal burning, 100. evaporation, Armour Glue Works, 105. influence of draft on efficiency, 106. influence of rate of combustion on air supply, 264. influence of rate of driving on capac- ity, 107. influence of size of coal on capacity, 40. influence of thickness of fire on effi- ciency, 109. oil fuel, 62. powdered coal, 46. Tests of burners, oil fuel, 62. Tests of chimney, 100-foot steel, 213. Tests of condenser auxiliaries, 449. Tests of condensers: dry air, 429. evaporative, 434. Pennel saturated air, 432. Weighton surface, 421. Tests of cooling towers, 468. Tests of economizers, 515. Tests of engines, A.S.M.E. code, 846. binary vapor, 324. compound condensing vs. non-con- densing, 304. compound vs. simple high-speed, 303. condensing, increase in power due to, 307. Corliss, 369. 5500-h.p. engine at Waterside Station, 309. four-valve vs. single-valve high-speed, 298. friction, 284. increasing back pressure, 283. increasing initial pressure, 286. Reeves simple engine, 297. Tests of engines, simple high-speed engines, saturated steam, 292. tables of, 296, 306, superheated steam, compound, 317. superheated steam, influence of degree of superheat, 322. superheated steam, influence of size of engine, 332. superheated steam, record perform- ance of, 321. superheated steam, triple, 318. Tests of furnace, relation of gas compo- sition to temperature, 36. Tests of injectors, 549. Tests of jets, steam, 248. Tests of oil burners, 62. Tests of oil separators, 572. Tests of oils, fuel, 52. Tests of oils, lubricating, 676. Tests of pipe coverings, 617. Tests of pumps: air lift, 572. boiler feed, 534. centrifugal, De Laval, 568. centrifugal, Lea-Degan, 569. centrifugal, Morris, 564. centrifugal, Worthington, 568. direct-connected triplex, 544. duplex fire pump, 533. geared triplex, 545. rotary, 570. Tests of separators, 576. Tests of spray fountain, 455. Tests of superheaters, 176-181. Tests of turbines, 390. Thermal efficiency of engines, 273. Thermal purification of feed water, 479. Thermo-electric pyrometers, 736. Thermometers, classification of, 731. Thermostat, Powers, 646. Thickness of fire, 109. Thickness of walls, brick chimneys, 226. Thomas, specific heat of superheated steam, 161. Throttling, calorimeter, 746. Throttling, moisture evaporated by, 312. Throttling vs. automatic cut-off, 310. Tile, "Economy," 131. Tile-roof furnaces, 63. Tilden damper regulator, 119. Tomlinson condenser, 415. INDEX 901 Towers, cooling, 456-470. Alberger, 459. Barnard, 457. test of, 468. theory of, 460. Worthington, 458. Traps (steam), 588-599. Acme, 590. bowl, 591. bucket, 590. Bundy, 591. classification of, 588. Columbia, 592. differential, 594. Dunham, 593. expansion, 592. Flinn, 595. float, 589. Geipel, 593. Heintz, 594. location of, 596. McDaniel, 589. return, 596. Traveling coal hoppers, 203. Traveling grates, 126. Triumph powdered-coal furnace, 50. Try cocks, 3, 120. Tube cleaners, 121. Tupper grate bar, 115. Turbines (steam), 327. advantages of, 382. Allis-Chalmers, 372. correction factors, 387. cost of, 392. Curtis, 350. De Laval, 331. Double-flow, 369. economy of space, 382. efficiency of, 384. elementary theory, 328. Hamilton-Holzworth, 362. impulse, 331. influence of superheat, 393. influence of vacuum, 395. Kerr, 346. low-pressure, 376. overload capacity, 384. reaction, 365. regulation, 384. simplicity of, 382. Terry, 345. Turbines (Steam), tests of, 390. Westinghouse-Parsons, 365. Turner oil filter, 688. Twin fire-arch furnace, 142. Ultimate analysis, 31. Underfeed stokers, 138. Unit of evaporation, 88. Units, conversion, 877. Universal calorimeter, 746. Useful life of power-plant appliances, 694. Vacua, increase of power due to increas- ing, 307. Vacua, influence of, on economy of engines, 307. Vacua, influence of high, on steam tur- bines, 395. Vacuum ash conveyor, 196. Vacuum augmenter, Parsons, 444. Vacuum chambers, 529. Vacuum, degree of, as affected by aqueous vapor, 405. drips under, 597. most economical, 451. Vacuum pumps, 552. Vacuum systems, high, 441. Valves, Act on atmospheric relief, 666. Anderson automatic non-return, 658. Anderson triple duty emergency, 659. atmospheric relief, 665. back pressure, 665. blow-off, 661. by-pass, Westinghouse-Parsons, 369. check, 660. coal, 205. consolidated pop safety, 668. Crane atmospheric, 666. Crane hydraulic emergency, 659. Davis back pressure, 665. diaphragm, 647. disk, 527. emergency, 558. foot, 668. Foster back pressure, 665. gate and globe, 655. Illinois Eng. Company's vacuum, 643. Kieley reducing, 667. Ludlow angle, gate pattern, 657. non-return, 658. nozzle, Curtis turbine, 352. OCT 5 191C 902 INDEX Valves, nozzle, De Laval turbine, 332. nozzle, theoretical, 334. Paul vacuum, 645. reducing, 666. stop, 655. Webster vacuum, 643. Vanes of Curtis turbine, 352. Hamilton-Holzworth turbine, 362. Westinghouse-Parsons turbine, 365. Van Stone joint, 613. Vegetable oils, 669. Velocity of steam through nozzles, 328. through pipes, 633. Velocity of water through nozzles, 650. through pipes, 650. Venturi meter, Vertical blow-off connections, 116. Vertical tubular boilers, 69. Volatile matter in coal, 23. Visible smoke, loss due to, 38. Viscosity of oils, 673. Volute centrifugal pump, 561. Volume of jet condenser chamber, 408. Wainwright feed-water heater, 493. Wall brackets, piping, 622. Wanner optical pyrometer, 738. Warren fuel oil burner, 59. Washed coals, 40. Water and boiler scale, analyses of, 407. Water columns, 119. Water-cooling systems, 453. Water, flow of, in pipes, 650. Water, friction coefficient in clean iron pipes, 652. Water, height of lift by suction, 538. Water pistons, 530. Water-softening plants, 480. Water temperatures, feed-water heaters, 490. Water, weight of, for condensing, 406. Waterworks, centrifugal pump charac- teristics for, 565. Weber concrete-steel chimneys, 235. Webster feed-water heaters, 488. Webster heating system, 406. Webster vacuum valve, 643. We-Fu-Go purifying system, 481. Weighing fuel, 852. Weighing water, 850. Weight of air as indicated by flue-gas analysis, 30. Weight of, Weight of boiler compound necessary, 478. Weight of guyed steel chimneys, 219. Weight of water evaporated per square foot of heating surface, 95. Weighton multi-flow surface condenser, 419. Weiss barometric condenser, 411. West Albany power station, N. Y. C. R. R., 788-797. Western Electric Co.'s Power plant, 272. Westinghouse-Leblanc condenser, 445. Westinghouse-Parsons steam turbines, 365. Wet air pumps, 553. Wheeler admiralty surface condenser, 417. Wheeler, C. H., multi-flow surface con- denser, 443. White-Star oil filter, 688. Wickes, boiler, 84. Wilkinson stoker, 136. Williams oil burner, 58. Wire drawing, 254. Wood as fuel, 18. Wooley smokeless furnace, 149. Worthington barometric condenser, 414. conoidal pump, test of, 568. cooling tower, 458. jet condenser, 402. Wrought-iron pipe, 611. Yonkers power house, N.Y.C.R., piping for, Zinc, use of, in boilers, 477.