Class T J Z 15 Boole !._i Copyright^?- /?££. COPYRIGHT DEPOSIT. THE WILEY TECHNICAL SERIES FOR VOCATIONAL AND INDUSTRIAL SCHOOLS EDITED BY JOSEPH M. JAMESON GIRARD COLLEGE THE WILEY TECHNICAL SERIES EDITED BY JOSEPH M. JAMESON TEXT BOOKS IN MECHANICS, HEAT AND POWER NOW READY Steam Power. By C. F. Hirshfeld, Formerly Pro- fessor of Power Engineering, Sibley College, Cornell University, and T. C. Ulbricht. xi-f 474 pages. 5§ by 8. 252 figures. Cloth, $3.25 net. Gas Power. By C. F. Hirshfeld, Formerly Professor of Power Engineering, Sibley College, Cornell Uni- versity, and T. C. Ulbricht, Formerly Instructor, Department of Power Engineering, Cornell Uni- versity, x + 209 pages. 5£by8. 60 figures. Cloth $1.75 net. Heat: A Text Book for Technical and Industrial Students. By J. A. Randall, formerly Pratt Institute, xiv+331 pages. 5j by 8. 80 figures. Cloth, $2.00 net. Elementary Practical Mechanics. By J. M. Jameson, Girard College. xii + 321 pages. 5£ by 8. 212 figures. Cloth, $1.75 net. 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Exercises in Heat and Light. By J. A. Randall, Formerlv Pratt Institute. 8 by lOf. Loose Leaf. 17 exercises. Comnlete, paper cover, 34 cents net. Jameson, Girard 8 by 10^. Loose paper cover, 85 For full announcement see 1'st following index. STEAM POWER BY C. F. HIRSHFELD, M.M.E. FOKMEKLY PROFESSOR OF POWER ENGINEERING, SIBLEY COLLEGE, CORNELL UNIVERSITY T. C. ULBRICHT, M.E., M.M.E. FORMERLY INSTRUCTOR, DEPARTMENT OF POWER ENGINEERING, SIBLEY COLLEGE, CORNELL UNIVERSITY, MEMBER AMERICAN SOCIETY OF MECHANICAL ENGINEERS SECOND EDITION Revised and Enlarged TOTAL ISSUE FOURTEEN THOUSAND NEW YORK JOHN WILEY & SONS, Inc London: CHAPMAN & HALL, Limited 1922 Copyright, 191G, 1922, by C. F. HIRSHFELD and T. C. ULBRICHT $ y . .. , - ^ 8/22 DEC -SB22 PRESS OF BRAUNWORTH & CO. BOOK MANUFACTURERS BROOKLYN, N. Y„ C1A690514 PREFACE TO SECOND EDITION The comments and criticisms which have come to the authors during the five years since this book appeared do not seem to call for any radical change in arrangement or treatment. There does appear to be a demand for the addi- tion of a small amount of material. This has been met by the inclusion of a chapter on " Performance of Steam Power Equipment " in which are included treatments of those sub- jects which users of the book seem to feel are necessary for completeness. Other additions of minor character have been made to care for the development of the art since the date of the first edition. The entire text has been carefully reviewed and certain parts have been rewritten to make them clearer. Certain minor errors and misprints discovered by users have also been corrected. The authors desire to express their sincere thanks to all the users of this book who have assisted in this revision by indicating errors and omissions and points at which the text was not readily intelligible. C. F. H. T. C. U. A' PREFACE TO FIRST EDITION The following pages represent the results of an attempt to collect in a comparatively small book such parts of the field of steam power as should be familiar to engineers whose work does not require that they be conversant with the more complicated thermodynamic principles considered in advanced treatments. The experience of the authors has led them to believe that a book of this sort should give a correct view-point with regard to the use of heat in the power plant even though it does not enter deeply into the theoretical considerations leading up to that view-point; that it should supply the tools required for the solution of power plant problems of the common sort ; and that it should give sufficient description of power plant apparatus to make the reader fairly familiar with the more common types. Mathematical treatment of the subject has been elim- inated to the greatest possible extent, and anyone familiar with elementary algebra should be able to understand readily such equations as it has been deemed necessary to include. Brief explanations of physical and chemical concepts are given in every case in which the text required their use, so that those who have not studied these subjects, and those who have but have failed to crystallize and hold the neces- v i PREFACE sary ideas, should have little difficulty in reading the text understandingly. It is hoped that the book may prove serviceable as a text for steam power courses given to civil engineers in the various colleges and that it may also meet the needs of those instructing power plant operators in industrial schools. C. F. H. T. C. U. June, 1916. CONTENTS CHAPTER I PAGE Physical Conceptions and Units 1 1. Matter. 2. Energy. 3. Units of Matter and Energy. 4. Work. 5. Mechanical Energy. 6. Heat. 7. Temper- ature. 8. Measurement of Temperature. 9. The Unit of Heat Energy. 10. Specific Heat. 11. Quantity of Heat. 12. Work and Power. CHAPTER II The Heat-power Plant 20 13. The Simple Steam-power Plant. 14. Cycle of Events. 15. Action of Steam in the Cylinder. 16. Hydraulic Analogy. CHAPTER III Steam 27 17. Vapors and Gases. 18. Properties of Steam. 19. Generation of Steam or Water Vapor. 20. Heat of Liquid, q or h. 21. Latent Heat of Vaporization, r or L. 22. Total Heat of Dry Saturated Steam, X or H. 23. Total Heat of Wet Steam. 24. Heat of Superheat. 25. Total Heat of Super- heated Steam. 26. Specific Volume of Dry Saturated Steam, V or S. 27. Specific Density of Dry Saturated Steam, — or 8. 28. Reversal of the Phenomena Just De- scribed. 29. Generation of Steam in Real Steam Boiler. 30. Gauge Pressure. CHAPTER IV The Ideal Steam Engine 43 31. The Engine. 32. Operation of the Engine. 33. Work Done by the Engine. 34. Heat Quantities Involved. 35. Efficiency. 36. Effect of Wet Steam. 37. Application to viii CONTENTS PAGE a Real Engine. 38. Desirability of Other Cycles. 39. The Complete Expansion Cycle. 40. The Incomplete Expan- sion Cycle. CHAPTER V Entropy Diagram 61 41. Definitions. 42. Temperature-Entropy Chart for Steam. 43. Quality from T^-Chart. 44. Volume from T- chart. 45. Heat from T0-chart. 46. The Complete T<£- chart for Steam. CHAPTER VI Temperature Entropy Diagrams of Steam Cycles 72 47. Complete-expansion Cycle. 48. Area of Cycle Repre- sentative of Work. 49. Modifications for Wet and Super- heated Steam. 50. Incomplete Expansion Cycle. 51. Effect of Temperature Range on Efficiency. CHAPTER VII The Real Steam Engine 77 52. Operations of Real Engine. 53. Losses in Real In- stallations. 54. Clearance. 55. Cushion Steam and Cylinder Feed. 56. Determination of Initial Condensation. 57. Methods of Decreasing Cylinder Condensation. 58. Classi- fication of Steam Engines. 59. Rotative Speeds and Piston Speeds. 60. The Simple D-Slide Valve Engine. 61. Engine Nomenclature. 62. Principal Parts of Engines. CHAPTER VIII The Indicator Diagram and Derived Values 115 63. The Indicator. 64. Determination of I.h.p. 65. Conventional Diagram and Card Factors. 66. Ratio of Expansion. 67. Determination of Clearance Volume from Diagram. 68. Diagram Water Rate. 69. T^-diagram for a Real Engine. 70. Mechanical and Thermal Efficiencies. CHAPTER IX Compounding 141 71. Gain by Expansion. 72. Compounding. 73. The Compound Engine. 74. Cylinder Ratios. 75. Indicator Diagrams and Mean Pressures. 76. Combined Indicator Diagrams. CONTENTS ix CHAPTER X PAGE The D-Slide Valve 159 77. Description and Method of Operation. 78. Steam Lap. 79. Lead. 80. Angle of Advance. 81. Exhaust Lap. 82. The Bilgram Diagram. 83. Exhaust and Compression. 84. Diagram for Both Cylinder Ends. 85. Piston Positions. 86. Indicator Diagram from Bilgram Diagram. 87. Limita- tions of the D-slide Valve. 88. Reversing Engines. 89. Valve . Setting. CHAPTER XI Corliss and Other High-efficiency Engines 196 90. The Trip-cut-off Corliss Engine. 91. Non-detaching Corliss Gears. 92. Poppet Valves. 93. The Una-flow En- gine. 94. The Locomobile Type. CHAPTER XII Regulation 216 95. Kinds of Regulation. 96. Governor Regulation. 97. Methods of Varying Mean Effective Pressure. 98. Con- stant Speed Governing. 99. Governors. CHAPTER XIII The Steam Turbine 224 100. The Impulse Turbine. 101. Theoretical Cycle of Steam Turbine. 102. Nozzle Design. 103. Action of Steam on Impulse Blades. 104. De Laval Impulse Turbine. 105. Gearing and Staging. 106. The Reaction Type. 107. Com- bined Types. 108. Steam Consumption of Steam Turbines. 109. Low Pressure Turbines. 110. Steam Turbo-generators. CHAPTER XIV Condensers and Related Apparatus 261 111. The Advantage of Condensing. 112. Measurement of Vacuum. 113. Conversion of Readings from Inches of Mercury to Pounds per Square Inch. 114. Principle of the Condenser. 115. Types of Condensers. 116. The Jet Con- denser. 117. Non-contact Condensers. 118. Vacuum Pumps or Air Pumps. 119. Water Required by Contact Condensers. 120. Weight of Water Eequired by Non-contact CONTENTS PAGE Condensers. 121. Relative Advantages of Contact and Sur- face Condensers. 122. Cooling Towers. CHAPTER XV Combustion 296 123. Definitions. 124. Combustion of Carbon. 125. Combustion to CO. 126. Combustion to C0 2 . 127. Com- bustion of CO to CO2. 128. Conditions Determining Forma- tion of CO and C0 2 . 129. Flue Gases from Combustion of Carbon. 130. Combustion of Hydrogen. 131. Combustion of Hydrocarbons. 132. Combustion of Sulphur. 133. Com- bustion of Mixtures. 134. Composition of Flue Gases. 135. Temperature of Combustion, CHAPTER XVI Fuels 317 136. Commercial Fuels. 137. Coal. 138. Coal Analyses. 139. Calorific Value of Coals. 140. Purchase of Coals on Analysis. 141. Petroleum. CHAPTER XVII Steam Boilers 326 142. Definitions and Classification. 143. Functions of Parts. 144. Furnaces and Combustion. 145. Hand Firing. 146. Mechanical Grates. 147. Smoke and Its Prevention. 148. Mechanical Stokers. 149. Rate of Combustion. 150. Strength and Safety of Boiler. 151. Circulation in Boilers. 152. Types of Boilers. 153. Boiler Rating. 154. Boiler Efficiencies. 155. Effects of Soot and Scale. 156. Scale. 157. Scale Prevention. 158. Superheaters. 159. Draft Apparatus. CHAPTER XVIII Recovery of Waste Heat 39ft 160. Waste Heat in Steam Plant. 161. Utilization of Ex- haust for Heating Buildings. 162. Feed-water Heating. CONTENTS xi CHAPTER XIX PAGE Boiler-feed Pumps and Other Auxiliaries. 410 163. Boiler-feed Pumps. 164. The Steam Injector. 165. Separators. 166. Steam Traps. 167. Steam Piping. CHAPTER XX Performance of Steam Power Equipment 419 168. Meaning of Performance. 169. Determination of Boiler Performance. 170. Fuel Calorimeters. 171. Steam Calorimeters. 172. Determination of Engine Performance. 173. Determination of Delivered Horse-power. STEAM POWER CHAPTER I PHYSICAL CONCEPTIONS AND UNITS 1. Matter. The universe is generally pictured as com- posed of matter and energy. Matter is regarded as that which is possessed of mass, or as that which is possessed of inertia; i.e., which requires the action of force to put it in motion, to bring it to rest or to change its velocity. These definitions merely enumerate characteristics of matter; they do not tell what it really is. In the present state of knowledge it is, however, impossible to define matter in any other way. No experiment has yet shown that matter can be created or destroyed by man. It can be changed from one form to another, it can be given certain physical and certain chem- ical characteristics, more or less at will, but the actual quantity of matter concerned is always the same after and before such changes. It is customary to state this experi- ence in the form of a law known as the Law of the Con- servation of Matter, which states that the " total quantity of matter in the universe is constant." Matter is known to exist in several physical states or conditions of aggregation. The three most familiar are (1) solid, (2) liquid and (3) gaseous. In each of these states matter is conceived as made up of minute particles called molecules which in turn are apparently composed of still smaller parts known as atoms. These atoms can also be broken into parts, but for the purposes of this book it is not necessary to consider such further subdivision. 2 STEAM POWER Experiment and mathematical reasoning seem to indi- cate that the molecules of all materials are in constant motion and that there are neutralizing attractive and repul- sive forces acting between them. In solids the molecules are apparently bound together in such a way that, although they are in constant motion, the external form or shape of the body tends to remain constant; in fact it requires the expenditure of force to cause a change of form. In liquids the molecular attraction is so altered that practically all rigidity disappears and the shape assumed by the liquid is determined by that of the surrounding surfaces, as, for instance, the shape of the vessel containing the liquid. In gases the molecules are still more free and actually tend to move apart as far as possible, so that a gas will spread in all directions until it fills any closed containing vessel. 2. Energy. Nearly everyone has a conception of what is meant by the term energy, but no one yet knows what energy really is. It is defined as the capacity for doing work, or the ability to overcome resistance. A man is said to be very energetic or to be possessed of a great deal of energy when he has the ability to perform a great amount of work or to overcome great resistances. Matter is said to be pos- sessed of energy when it can perform work or overcome resistance. Actually, matter is not known in any form in which it is not possessed of energy. There are many different forms of energy. A body in motion can do work and is said to be possessed of mechani- cal energy. A body which we recognize as hot can do work at the expense of the heat associated with it and is said to be possessed of heat energy. Light, sound and electricity are all forms of energy. Experiment and experience have never shown that energy can be destroyed or created by man, but they have shown that one form of energy can be converted into another form under proper conditions. The first part of this experience is stated as a law known as the Law of the Conservation of PHYSICAL CONCEPTIONS AND UNITS Energy. This law states that " the total quantity of energy in the universe is constant" 3. Units of Matter and of Energy. When attempts are made to measure the amount of anything, some unit of measurement is adopted. Matter is measured in numerous ways and many units are used. The common methods of measuring matter are by volume and by weight. Engineers in English-speaking countries use the cubic yard, the cubic foot or the cubic inch as units in measuring matter by vol- ume and they use the pound, the ounce, the grain, etc. as units in measuring matter by weight. Energy is measured in many units and, in general, there is a characteristic unit or set of units for each form in which it occurs. Thus the foot-pound is very commonly used for measuring mechanical energy; the British thermal unit for measuring heat energy; and the joule for measuring electrical energy. Some of these units will be defined and considered in greater detail in subsequent paragraphs. 4. Work. Work is defined as the overcoming of a resistance through a distance. Thus, work is done when a weight is raised against the resistance offered by gravity; work is done when a spring is compressed against the resistance which the metal offers to change of shape; work is done when a body is moved over another against the resistance offered by friction. The unit of work is the quantity of work which must be done in raising a weight of one pound through a vertical distance of one foot. It is called the foot-pound. Thus, one foot-pound of work must be done in raising one pound one foot; two foot-pounds of work must be done in raising two pounds one foot or in raising one pound two feet. fig. i If a weight of one pound were suspended from a spring balance as shown in Fig. 1, the balance would in- dicate a pull or force of one pound. No work would be 4 STEAM POWER done by this force as long as the weight remained stationary, because no resistance would be overcome through a distance. If, however, the same weight were slowly or rapidly raised a vertical distance of a foot, one foot-pound of work would be done. A force or pull of one pound would then have overcome a resistance of one pound through a distance of one foot. In general: Work in ft.-lbs. = Resistance overcome in lbs. X distance. = Force in Ibs.Xdistance in ft. so that if a force of 10 lbs. pushes or pulls anything which offers a resistance of 10 lbs. while that something travels a distance of, say, 5 ft., the work done will be given by the expression, Work =10X5, = 50 ft.-lbs. A body in falling a certain distance can do work equal to its weight multiplied by the distance it falls because it could theoretically raise an equal weight an equal distance against the action of gravity, and the work done upon this second body would be equal to its weight multiplied by the distance through which it was raised. It is very important to note that no work is done by a force if there is no motion; resistance must be overcome through a distance in order that work may be done. Thus, a force of 1000 lbs. might be required to hold something in position, that is to balance a resistance, but no work would be done if the body upon which the 1000-pound force acted did not move. Again, a weight of 50 lbs. held at a distance ol 10 ft. above the surface of the earth would exert a downward push or pull equal to 50 lbs. on whatever held it in that position; it would, however, do no work if held in that position. If allowed to fall through the distance of 10 ft. it could do 50X10 = 500 ft.-lbs. of work. It is very convenient to represent graphically the action PHYSICAL CONCEPTIONS AND UNITS 5 of forces overcoming resistances, that is, doing work. This is done by plotting points showing the magnitude of the force at the time that the body on which it is acting has traveled different distances. Thus, suppose a constant force of 10 lbs. pushes a body a distance of 15 ft. against a constant resistance of 10 lbs. The force acting on the body will have a value of 10 lbs. just as the body starts to move, a value of 10 lbs. when the body has moved 1 ft., a value of 10 lbs. when the body has moved 2 ft., and so on. This might be represented by points on squared paper as shown o 4 3 1 6 7 8 9 10 11 Distance traveled in Feet Fig. 2. 12 13 14 15 in Fig. 2 or by a horizontal line joining those points as shown in the same figure. The work done by this force would be 10X15 = 150 ft.-lbs. according to our previous definition. But 10X15 is also the number of small squares under the line represent- ing the action of this force in Fig. 2. The number of these small squares then must be a measure of the work done, but it is also a measure of the area under the line represent- ing the action of the force, so that this area must be a measure of the work done. Each small square represents 1 lb. by its vertical dimension and 1 ft. by its horizontal dimension, 6 STEAM POWER so that its area must represent 1 lb. XI ft. = 1 ft. -lb. The total number of squares below the line equals 10 X 15 = 150, and since the area of each one represents 1 ft.-lb. the total area under the line represents 150X1 = 150 ft.-lbs. It is not always convenient to choose such simple scales as those just used. Thus it might be more convenient to plot the action of this force as is done in Fig. 3. Here the height of a square represents 2 lbs. and the width represents 1 ft.; the area then represents 2X1=2 ft.-lbs. There are 5X15 = 75 squares under the line and as each „ 12 lio I 8 .5 6 o 4 L 5 > j 5 L > ( Dis tanc 1 3 tra i i veled ) l inF 1 eet 1 12 13 . 14 15 Fig. 3. represents 2 ft.-lbs. the total area under the line represents 2X75 = 150 ft.-lbs. as before. This is a very useful property of these diagrams and the area under the line representing the action of the force always represents the work done, no matter what the shape of that line. Thus, assume a force which compresses a spring a distance of 6 ins. Suppose that a force of 10 lbs. is required to com- press the spring 1 in., a force of 20 lbs. to compress it 2 ins., and so on up to a force of 60 lbs. to compress it 6 ins. Starting with a force of zero, the force will have to gradually increase as the spring is compressed, as shown by the line in Fig. 4. The area of each of the small squares will represent wx h=T2 ft - lbs - Under the line there is an area equal PHYSICAL CONCEPTIONS AND UNITS to 5^5 - is sma ,n squares, and the work done in compressing the spring must then be 18X?o = 15 ft.-lbs. 5. Mechanical Energy. Any body which exists in such a position or location that it could do work by dropping or falling is said to be possessed of potential mechanical energy, or of mechanical energy due to position. As long as it remains in this position, it cannot do work at the expense of this energy, but, if allowed to fall, it could do so. The Fig. 4. — Graph Showing Action of Spring. work it could do would be equal to the product of its weight by the distance it could fall and the potential energy it possesses before starting to fall is measured by this work. Thus, a body weighing 40 lbs. located 10 ft. above the surface of the earth could do 40X10 = 400 ft.-lbs. of work in falling, and, therefore, it is said to be possessed of 400 ft.-lbs. of potential energy before it starts to fall. If in falling it raises a weight equal to its own (theo- retically) through a distance equal to that through which it falls (theoretically), it will have used up 400 ft.-lbs. of energy in doing 400 ft.-lbs. of work upon the body raised 8 STEAM POWER and will no longer be possessed of that amount of potential energy. The body which has been raised will, however, have an equal amount of energy stored in it and will in turn be able to do 400 ft. -lbs. of work if allowed to fall a distance of 10 ft. If the body assumed above falls through a distance of 10 ft. without raising another body or doing an equivalent amount of work in some other way, it acquires a high velocity. When it arrives at the bottom of the fall of 10 ft., it certainly does not possess the 400 ft. -lbs. of potential energy which it had before dropping nor has it done work at the expense of that energy. Moreover, the energy could not have been destroyed because it is indestructible. The only conclusion is that it must still be possessed of this energy in some way. At the end of the fall it has lost its advantageous position, but it has acquired a high velocity, and experience shows that if brought to rest it can do work upon that which brings it to rest equal to what it could have done in raising a weight as previously described. At. the end of its fall and before being brought to rest, the body is therefore said to be possessed of energy by virtue of its velocity, and this form of energy is called kinetic mechanical energy. The kinetic energy will be exactly equal to the 'potential energy which disappeared as the body fell. Any body which is moving is possessed of kinetic energy because it can do work on anything which brings it to rest. This energy is expressed by the equation, 1 W Kinetic Energy in ft.-lbs.=-Xw^X V 2 , in which W = the weight of the moving body in pounds. F = the velocity in ft. per second, and 32.2 = a gravitational constant commonly represented by g. PHYSICAL CONCEPTIONS AND UNITS 9 6. Heat. One of the most familiar forms of energy is heat, which manifests itself to man through the sense of touch. In reality every body with which man is familiar possesses an unknown amount of heat energy and it is assumed that this heat energy is in some way associated with the motions and relative positions of the molecules and their constituents. For this reason heat is often described as molecular activity and is regarded as energy stored up in a substance by virtue of its molecular condition. Heat energy can be made to perform work in ways which will be discussed later and this is proof that it is a form of energy and not a material substance, as was once supposed. Heat is observed and recorded by its effects on matter, producing changes in the dimensions or volumes of objects; changes of internal stress; changes of state, as ice to water and water to steam; changes of temperature; and electrical and chemical effects. Neglecting certain atomic phenomena not yet well under- stood, the probable source of all heat energy appearing on the earth is the sun. Heat, however, may be obtained from mechanical and electrical energy; from chemical changes; from changes of physical state; from the internal heat of the earth. 7. Temperature. Man early realized that under certain conditions bodies felt " hotter " than under other conditions and gradually came to speak of the " degree of hotness " as the temperature of the body. It was later realized that what was really measured as the " hotness " or intensity of heat or temperature of a body was the ability of that body to trans- mit heat to others and that it had no connection with quantity of heat. Thus, if the temperature of two adjacent bodies happened to be the same, one of them could not lose heat by trans- mitting it to the other, but if the temperature of one happened to be higher than that of another, the body at 10 STEAM POWER higher temperature would always lose heat to the one at lower temperature. As a means of measuring temperature certain arbitrary scales have been chosen. The centigrade scale of tempera- ture, for instance, is based upon the temperatures of melting ice and boiling water under atmospheric pressure. The tem- perature difference between boiling water at atmospheric pressure and melting ice at atmospheric pressure is arbi- trarily called one hundred degrees of temperature, and the temperature of the melting ice is called zero, making that of the boiling water 100 degrees. Any body which has such a temperature that it will not give heat to, or take heat from, melting ice is said to be at a temperature of zero degrees centigrade, represented as 0° C. Similarly, any body in such a condition that it will not give heat to or take heat from water boiling under atmospheric pressure is said to have a temperature of 100° C. A body with a temperature exactly half way between these two limits would then be said to have a temperature of 50° C. 8. Measurement of Temperature. The temperatures of bodies could be determined by bringing them in contact with such things as melting ice and boiling water and determining whether or not a transfer of heat occurred, but this would be a very cumbersome and unsatisfactory method. As a consequence many other means have been devised for the measurement of temperature. One of the most common and convenient methods in- volves the use of what are known as mercury thermometers. These depend upon the fact that the expansion of mercury with changing temperature is very uniform over a wide temperature range. Thus, if mercury expands a certain amount when its temperature is raised from that of melting ice to that of boiling water, i.e., 100° C, it will expand just half as much when its temperature is raised half as high, and one-quarter as much when its temperature is raised one- quarter of the range from 0° to 100° C. PHYSICAL CONCEPTIONS AND UNITS 11 n The thermometer is made by enclosing a small quan- tity of mercury in a glass tube fitted with a bulb at one end, as shown in Fig. 5. The lower end of the thermometer is immersed in melting ice and the point on the stem which is reached by the top of the mercury column is marked and labelled 0° C. The thermometer is then immersed in the steam from water boil- ing under atmospheric pressure and the point reached by the top of the mercury column is marked and labelled 100° C. The distance be- tween the two marks is then divided into one hundred parts and each represents the distance which the end of the column of mercury will move when its temperature changes one centi- grade degree. It is customary to extend this same scale below 0° and above 100°, carrying it, on ex- pensive thermometers, as far in each direction as the approximation to a constant expansion on the part of the mercury and to constant properties of the glass justifies. The temperature of a body can then be found by placing the thermometer in or in contact with that body and noting the point reached by the end of the mercury column. The division reached gives the temperature directly. The centigrade scale just described is the one commonly used by scientists the world over, but engineers in this country more often use what is known as the Fahrenheit scale, This is so chosen that the temperature of melting ice is called 32° F. and the temperature of water boil- ing under atmospheric pressure is called 212° F. There are i Fig. 5.— Mer- cury Ther- mometer. c.° i o -1778 c « Fig. 6. — Comparison of Centigrade and Fahrenheit Scales, 12 STEAM POWER thus 180° on this scale for the same temperature difference as is represented by 100° on the centigrade scale. The relation between the two scales is shown diagrammatically in Fig. 6. It is apparent that the temperature of a body at 0° C. will be 32° F. and that of a body at 0° F. will be -17.8° C. Since 100 centigrade degrees are equal to 180 Fahren- heit degrees, it follows that P 180 9 C = 100 = 5 K and that 1 F -T80~9 °- (1) (2) Therefore, if t F and t c represent temperatures on the Fahren- heit and centigrade scales respectively, tF — -=tc ~T~o2i . and 9 (fc-32) (3) (4) Cent . ^ Abs. — ^Fahr. 100 073 273" 491.4 459.4' There is still another temperature scale of great impor- tance. It is known as the absolute scale and temperatures measured on it are spoken of as absolute tem- peratures. The zero on this scale is located at -273° C. or 273 centi- grade degrees below centigrade zero, or, what is the same thing, at -459.4° F., or 459.4 Fahrenheit degrees below Fahrenheit zero. The degrees used are either centigrade or Fahrenheit, as convenient, so that there are absolute tempera- tures expressed in centigrade de- grees above absolute zero and there are absolute tempera- -273 Fig. 7. — Comparison of Ab- solute and Ordinary Temperature Scales. PHYSICAL CONCEPTIONS AND UNITS 13 tures expressed in Fahrenheit degrees above absolute zero. The relations between the various scales are shown dia- grammatically in Fig. 7. It is apparent from this diagram that, and that 7V = £f+460 (approximately) . . . (5) T c = t c +27S (6) if TV and T c represent absolute temperatures and if the number 459.4 is rounded out to 460, as is commonly done. 9. The Unit of Heat Energy. The unit used in the measurement of heat energy in the United States is the British Thermal Unit (abbreviated B.t.u). It is defined as the quantity of heat required to raise the temperature of one pound of pure water one degree Fahrenheit. In order to make the definition very exact it is necessary to state the temperature of the water before the temperature rise occurs, because it requires different amounts of heat to raise the temperature of a pound of water one degree from differ- ent initial temperatures. For ordinary engineering pur- poses, however, such refinements generally may be omitted. Many experimenters have shown that heat energy and mechanical energy are mutually convertible, that is, the one can be changed into the other. When such a change occurs no energy can be lost since energy is indestructible, and it follows that, if one form is changed into the other, there must be just as much energy present after the change as there was before. As the units used in measuring the two forms of energy are very different and as it is often necessary to express quantities of energy taking part in such conversions, it is desirable to determine the relations between these units. This was first accurately done by Joule, who showed that one British thermal unit of heat energy resulted from the con- 14 STEAM POWER ' version of 772 ft.-lbs. of mechanical- energy. Later experi- menters have shown that the number 778 more nearly expresses the truth than does the number 772 and the larger value is now known as Joule's Equivalent. Expressed mathematically, the relation between the units is 1 B.t.u.= 778 ft.-lbs (7) lft.-lb.=^B.t.u (8) 10. Specific Heat. The specific heat of a substance is defined as that quantity of heat which is used up or recovered when the temperature of one pound of the material in question is raised or lowered one degree. Its numerical value depends upon the specific heat of water since the quantity of heat is measured in units dependent upon the amount required to raise the temperature of water. The specific heat of water is, however, very variable, as shown by the values given in Table L, and it is therefore evident that exact numerical values of specific heats can only be given when the definition of the B.t.u. is exactly expressed. The specific heats of all real substances vary with tem- perature and the values commonly used are either rough averages or are those determined by experiments at one temperature. For most engineering purposes errors arising from this source may, however, be neglected. From the definition of specific heat it follows that : C ~W{t2-hY (9) in which C = a mean or average specific heat over a range of tem- perature from t\ to t2, and Q = the heat supplied to raise the temperature of W pounds of material from t\ to fo. PHYSICAL CONCEPTIONS AND UNITS 15 TABLE I Specific Heats of Water. (Value at 55° F. taken as unity) Temp. F°. Spec. Ht. Temp. F°. Spec. Ht. 20 1.0168 250 1.045 30 " 1.0098 400 1.064 40 1.0045 450 1.086 50 1.0012 500 1.112 60 0.9990 510 1.117 70 . 0.9977 520 1.123 80 0.9970 530 1.128 90 0.9967 540 1.134 100 0.9967 550 1.140 120 0.9974 560 1.146 140 0.9986 570 1.152 160 1.0002 580 1.158 180 1.0019 590 1.165 200 1.0039 600 1.172 220 1.007 240 1.012 260 1.018 280 1.023 300 1.029 * Values taken from Marks and Davis, " Steam Tables and Diagrams," p. 68. ILLUSTRATIVE PROBLEMS 1. Given: Sp. ht. of iron=0.113, of aluminum =0.211; Initial temp. =150° F. Temp, range (k-U) =100° F. If 1 lb. of iron and 1 lb. of aluminum are cooled through this temperature range, how much more heat is lost in one case than in the other? Q a i = IFCai(*2-* 1 )==lX.21lX100=21.1 B.t.u. Q lT = WCir(t 2 -t 1 )=lX. 113X100 = 11.3 B.t.u. Difference 9.8 B.t.u. 2. If the difference obtained in Prob. 1 were used to heat up 5 lbs. of silver, with a specific heat equal to 0.057, what would be the temperature range through which it would be raised? Q =9.8 =5X0.057(^2 -^) =0.285(£ 2 -h) /. k-*i=34.4° F. 16 STEAM POWER 3. If the initial temperature of the silver in Prob. 2 were 150° F e what would be the final absolute temperature Fahr.? k =fc+34.4° = 150+34.4 = 184° (approximately). T 2 =460+184 =644° F. Abs. 4. 100 lbs. of water in a 20-lb. tank of iron, both at 60° F., are placed in salt brine at 0° F. The water becomes ice at' 32° F. and the temperature of the ice is lowered to 26° F., the brine being raised to 26° F. Sp. ht. water = 1.0; Sp. ht. ice =0.5; Sp. ht. iron =0.113; Sp. ht. brine =0.8; and 143 B.t.u. per pound of water must be removed to convert liquid water at 32° F. to ice at the same temperature. What weight of brine is required? 100[l(60-32) + 143+.5(32 -26)]+20X0.113(60 -26) = 17X0.8(26-0) W = 840 lbs. of brine. 11. Quantity of Heat. It is impossible to determine the total quantity of heat in or " associated with " a substance, because no means of removing and measuring all the heat contained in any real material have ever been devised. Since, however, the engineer is concerned with changes of heat content ratber than with the total amount of heat contained, this fact causes him no difficulty. For convenience in figuring changes of heat content, it is customary to assume some arbitrary starting 'point or datum and to call the heat in the material in question zero at that point. Thus, for example, if it were necessary to figure heat changes experienced by a piece of iron weighing 5 lbs. and having a specific heat of 0.1138, and the temperature of this iron never dropped below 40° F. under the conditions exist- ing, this temperature might be taken as an arbitrary starting point above which to figure heat contents. If the iron were later found at a temperature of 75° F., " the heat content above 40° F." would be said to be Q = CW (t 2 -h) =0.1138X5(75-40) =2.27 B.t.u. PHYSICAL CONCEPTIONS AND UNITS 17 This type of formula can only be used when the sub- stance does not change its state between the limits of tem- perature concerned. In the case of water which might change to steam during such a rise of temperature, it might be necessary to include other heat quantities in the cal- culations, as shown in a later chapter. 12. Work and Power. Since steam engines are designed for the purpose of converting the heat energy contained in fuel into mechanical energy which may be used to perform work, it will be necessary to consider the units used in measuring work and power. Work was defined in a previous paragraph as the over- coining of a resistance through a distance, by the application of a force; that is, a force expressed in pounds, multiplied by the distance in feet through which the force acts, gives a product expressed in foot-pounds. The amount of work performed in a unit of time is termed power, which may be defined as the rate of doing work, Therefore, .p. Force X Distance ,.,_,. Power = —. j—. -. . . . (10) lime (mm. or sec.) The unit of power used by steam engineers is the horse- power, which is equivalent to the performance of 33,000 ft.-lbs. of work per minute, or 550 ft. -lbs. of work per second, or 1,980,000 ft.-lbs. per hour. Therefore, the horse-power developed by any mechanism is , ft.-lbs. of work per min. ,.,.,. h - p - = 3^000 • • • • (U) Since 33,000 ft.-lbs. of work can be accomplished only by the expenditure of 33,000 ft.-lbs. of energy and since one B.t.u. of energy is equal to 778 ft.-lbs., it follows that 33,000 33 000 ft.-lbs. of work must be the equivalent of ' =42.41 B.t.u. 77o It is customary to speak of power in terms of horse- 18 STEAM POWER power-hours. One horse-power-hour means the doing of work equivalent to one horse-power for the period of one hour, or the doing of work at the rate of 33,000 ft.-lbs. per minute for an hour. A horse-power-hour is therefore equiva- lent to 33,000X60 = 1,980,000 ft.-lbs. As 33,000 ft.-lbs. are equivalent to 42.41 B.t.u., it follows that 42.41X60 = 2544.6 or about 2545 B.t.u. are the equivalent of one horse- power-hour. The number 2545 should be memorized as it is very often used in steam-power calculations. If an engine could deliver one horse-power-hour for every 2545 B.t.u. it received, it would be working without losses of any kind; that is, all the heat energy entering it would leave it in the form of useful mechanical energy. It will be shown later that this is impossible even in the most perfect or ideal engine. REVIEW PROBLEMS 1. Express 32° F. in degrees centigrade. 2. Express 150° F. in degrees centigrade. 3. Express 250° C. in degrees Fahrenheit. 4. Express the results of problems 1, 2 and 3 in absolute values. 5. What is the heat equivalent of 233,400 ft.-lbs. of work? 6. Find the heat supplied 10 lbs. of water when its temperature is raised from 20° F. to 160° F., assuming the mean specific heat over this range to be 0.997. 7. Find the temperature change of 2 lbs. of lead (sp. ht. 0.0314) when 20 B.t.u. are added. 8. How many B.t.u. must be abstracted to lower the tem- perature of 15 lbs. of water from 212° F. to 32° F., assuming the specific heat of water to be unity? 9. Find the weight of water which will have its temperature tripled in value by the addition of 250 B.t.u., the final temperature being 150° F. Assume specific heat unity. 10. The specific heat of a piece of wrought iron is 0.113 and of a given weight of water is 1.015. 1 cu. ft. of water weighs approxi- mately 62.5 lbs. Find the increase in temperature of 4 cu. ft. of water when a common temperature of 65° F. results from placing in the water a piece of iron weighing 15 lbs. at a temperature of 900° F. PHYSICAL CONCEPTIONS AND UNITS 19 11. Find the final temperature of the mixture, when 100 lbs. of iron (sp. ht. =0.113), at a temperature of 1200° F. are immersed in 300 lbs. of water (sp. ht. 1.001) at a temperature of 50° F. 12. Five pounds of silver (sp. ht. =0.057) at 800° F. are im- mersed in water at 60° F., resulting in a final temperature of 85° F. Assume Sp. ht. water = 1. What weight of water is necessary? 13. An engine is developing 10 horse-power. Express this in ft .-lbs. of work done per minute and find the amount of heat energy equivalent to this quantity of mechanical energy. 14. A pump raises 1000 lbs. of water 50 ft. every minute. How much work is done? Find the equivalent horse-power. 15. An engine develops 1,980,000 ft. -lbs. of work at the fly- wheel per minute. (a) Find the horse-power developed. (b) If this engine operated in this way for an hour, how many horse-power hours would it make available? (c) What would be the equivalent of this number of horse- power hours in British thermal units? CHAPTER II THE HEAT-POWER PLANT 13. The Simple Steam-Power Plant. The various pieces of apparatus necessary for the proper conversion of heat energy into mechanical power constitute what may be called a " Heat-Power Plant," just as the apparatus used in obtaining mechanical energy from moving water is called an hydraulic or water-power plant. Heat-power plants are distinguished as " Steam-Power Plants"; " Gas-Power Plants"; etc., according to the way in which the heat of the fuel happens to be utilized. The apparatus around which the plant as a whole centers, that is, the apparatus in which heat energy is received and from which mechanical energy is delivered, is termed the engine or prime-mover. This heat engine may use steam generated in boilers and may require certain apparatus, such as condensers, pumps, etc., for proper operation; or it may use gas, generated in gas-producers requiring coolers, scrubbers, tar extractors and holders, depending upon the class of fuel used and upon certain commercial considera- tions. Again, the power-plant may simply contain an in- ternal-combustion engine using natural gas, gasoline or oil, a type of plant which is now very common. But whatever type of plant is used, a general method of operation is common to all. Heat energy in fuel is constantly fed in at one end of the system and mechanical energy is delivered at the other end. The steam-power plant will be briefly described in the following paragraphs, showing the cycle of events with the attendant losses through the system. 20 THE HEAT-POWER PLANT 21 22 STEAM POWEE In Fig. 8 is shown a simple steam-power plant which con- verts into mechanical energy part of the heat energy, origi- nally stored in coal, by means of a prime-mover called a steam-engine. The main pieces of apparatus used in this type of plant are the steam-boiler; the steam-engine; the condenser; the vacuum pump; and the feed-pump. The energy stream shows the various losses occurring through- out the plant. These losses cause the " delivered energy " stream to be only a small fraction of the total heat sent into the system. 14. Cycle of Events. 1. Fuel is charged on the grate under the boiler, where it is burned with the liberation of a large amount of energy. Air is drawn or forced through the grates in proper proportions to support this combustion. The hot gases resulting pass over the tubes, in a definite path set by the baffle plates, so that the largest possible amount of heating surface may be presented to the products of combustion. There are certain losses accompanying this operation, such as radiation, loss of volatile fuel passing off unburned, loss of fuel through the grate, and loss of heat through the excess air which must always be supplied to insure com- bustion. 2. That part of the heat in the gases which is not lost by radiation from the boiler and in the hot gases flowing up the stack passes through the heating surfaces of the boiler to the water within. From 50 to 80 per cent of the total heat energy in the fuel passes through the heating surfaces and serves to raise the temperature of the water to the boiling point at the pressure maintained, and to con- vert this water into steam according to the requirements. 3. Having obtained steam within the boiler, it is led through a system of pipes to a steam engine, where some of the heat stored in the steam is converted into mechanical energy by the action of that steam against a piston. The steam is then discharged, or exhausted, from the engine THE HEAT-POWER PLANT 23 at a much lower temperature and pressure than when it entered. From 5 to 22 per cent of the available heat in the steam is converted into mechanical energy in the engine cylinder, and because of frictional and other losses occurring in the mechanism, only from 85 to 95 per cent of this energy is turned into useful work at the fly-wheel. 4. In some plants, known as non-condensing plants, the exhaust steam, which still contains the greater part of all the heat received in the boiler, is discharged to the atmos- phere and represents a complete loss. In others, known as condensing plants, the exhaust steam is led to a condenser, where^ it is condensed by cold water, which absorbs and carries away the greater quantity of the heat not utilized in the engine. The • condensed steam or " condensate " is then either discharged to the sewer or transferred by means of a vacuum-pump to the hot-well, from which it is drawn by means of the feed-water-pump, raised to the original pressure of the steam, and returned to the boiler. Here it is again turned into steam and the cycle of operations outlined above is repeated. Naturally there is some loss due to evaporation and leaks throughout the system, so that " make-up " water must be supplied. The series of events just described constitutes a complete, closed cycle of operations, wherein the water is heated, vaporized, condensed and returned to the boiler, having served only as a medium for the transfer of heat energy from fuel to engine and the conversion of part of that energy within the cylinder. The water in such a case is known as the working substance. It is often more convenient to discard the working sub- stance after it leaves the cylinder, as suggested above in the case of a non-condensing plant; or, as in the case of a gas engine, where a new supply of working substance must be supplied for each cycle, because the burned gases of the previous cycle cannot be used again. 24 STEAM POWER 15. Action of Steam in the Cylinder. In order to pre- pare for the more detailed discussion of the action of the steam in the engine cylinder, to be taken up in a later chapter, a brief outline of the events occurring within the prime-mover will be considered at this point. Steam enters the cylinder through some kind of an admission valve, and acts upon the piston, just as the latter has approximately reached one end of its stroke and is ready to return. The heat-energy stored up in the steam causes it to expand behind the piston, thereby driving the latter out and performing work at the fly-wheel. At about 90 or 95 per cent of the stroke, the exhaust valve opens, and the steam begins to exhaust, the pressure within the cylinder dropping almost to atmospheric or to that main- tamed in the condenser by the time the piston has reached the end of its stroke. . On the next or return stroke the remaining steam is forced out through the exhaust port, until, at some point before the end of the piston travel, the exhaust valve closes, and the low-pressure steam trapped in the cylinder is compressed into the clearance space so that its pressure rises. Admission then occurs, and the cycle is repeated. The diagram given in Fig. 9 illustrates the operation of steam within the cylinder. *^\ ear all c<3 ^ ^ _ ^ rpj^g diagram jg plotted between pressures of steam within the cylinder as ordinates and correspond- ~~*~ c . ' Cut-off (Closing of Admission Valve) Release (Opening of Exhaust Valve) %- Ex n f^5^. in S P iston positions as ■fj~ ^Atmospheric Pressure^ - absdSSaS 7 piston positions Th e method of obtain- Closing of Exhaust Valve ™ A a- -r, • t j- i ing such a diagram, known Fig. 9. — -Steam Engine Indicator & ° ' Diagram. as an indicator-diagram, will be fully described in a later chapter. Since vertical ordinates represent pressure in pounds per square inch, and horizontal abscissas renre- THE HEAT-POWER PLANT 25 sent feet moved through by the piston, the product of these two must be work. But the product of vertical by horizontal distances must also give area. Therefore, by Source of Water at High Head Energy Supplied Water Useful Energy- Motor f f Made Available <- Energy Discharged eceiver of Discharged Water at Low Head (a) Source of Heat at High Temp. IVvP & -Energy Supplied" High|Temp. tjabovej datum Heat Engine Useful Energy Made Available Low|Temp. 1 2 above datum -Energy Discharged [Receiver of Discharged Heat at Low Temperature (6) Fig. 10. — Hydraulic Analogy. finding the area enclosed within the bounding lines of the cycle and multiplying this by a proper factor, the foot- pounds of work developed within the cylinder can be determined. 16. Hydraulic Analogy. The operation of heat-engines is analogous to that of water-wheels. A water-wheel de- 26 STEAM POWER velops mechanical energy by receiving water under a high head, absorbing some of its energy, and then rejecting the fluid under a low head. Similarly, the heat-engine receives heat energy at a high temperature (head), absorbs some of it by conversion into mechanical energy, and then rejects the rest at a low temperature (head) . The analogy can be carried still further. The water- wheel cannot remove all the energy from the water, nor can the heat-engine remove all the heat-energy from the working substance. There is a certain loss in the material discharged in both cases and this cannot be avoided. This analogy is illustrated diagrammatically in Fig. 10 (a) and (6) in which the widths of the streams represent quantity of energy. CHAPTER III STEAM 17. Vapors and Gases. When a solid is heated, under the proper pressure conditions, it ultimately melts or fuses and becomes a liquid. The temperature at which this occurs depends upon the particular material in question and upon the pressure under which it exists. Ice, which is merely solid water, melts at 32° F. under atmospheric pressure, while iron melts at about 2000° F. under atmos- pheric pressure. When a liquid is heated, it ultimately becomes a gas, similar to the air and other familiar gases. If this gas is heated to a very high temperature and if the pressure under which it is held is not too great, it very nearly obeys certain laws which are simple and which are called the laws of ideal gases. When the material is in a state between that of a liquid and that in which it very nearly obeys the laws of ideal gases, it is generally spoken of as a vapor. This term is used in several different ways and with several different modifying adjectives which will be explained in greater detail in later sections. 18. Properties of Steam. Of the many vapors used by the engineer, steam or water vapor is probably the most important, because of the ease with which it can be formed and also because of the tremendous field in which it can be used. It is generated in a vessel known as a steam boiler y which is constructed of metal in such a way that it can contain water, and that heat energy, liberated from burning fuel, can be passed into the water, converting part or all of ft. into water vapor, that is, into steam. 27 28 STEAM POWER The properties of water vapor must be thoroughly under- stood before the steam engine and steam boiler can be studied profitably. Probably the easiest way of becoming familiar with these properties is to study the use made of heat in the generation of steam from cold water. 19. Generation of Steam or Water Vapor. For the pur- poses of development, assume a vessel of cylindrical form, fitted with a frictionless piston of known weight, as shown in Fig. 11, (a) and (b), the whole apparatus being placed under a bell-jar in which a perfect vacuum is maintained. IIH I hlli l HjB ' — ~jf (a) m !&%£■ (b) — ] . .:..;,/ To Vacuum Pump Fig. 11. — Formation of Steam at Constant Pressure. Assume one pound of water in the cylinder, with the piston resting on the surface of the liquid. There will be some definite pressure exerted by the piston upon the surface of the liquid, and its value will be determined entirely by the weight of the piston. It is convenient in engineering practice to refer all vaporization phenomena to some datum temperature, and since the melting point of ice, 32° F., is a convenient refer- ence point, it is used as a standard datum temperature, in practically all steam-engineering work. Therefore, assuming the water in the jar to be at 32° F., if heat is applied the temperature of the liquid will rise approximately 1° F. STEAM 29 for every B.t.u. of heat added, since the specific heat of water is approximately unity. Experiment shows that for each pressure under which the water may exist some definite temperature will be attained at which further rise of temperature will cease and the liquid will ozu / / 480 / / 440 / / a j / o / 5 3b0 3 / / u 6 ^ / / a / 3 o / Pi 4iU 1 / 1.7 m u 2 c CU .5 Ct/I '£/ 1 w 1 >7 \ c/ \ H' 1 \ §1 \ \% V .<%,? \ \ \\ \ \ 7^%/ ^ ^ s\ X N& ^^ ^ ^^-»^. C^: «^__ 50 — ■ i^^ 15 50 100 150 200 250 Temperatures above Saturation*^. Fig. 14. — Variation of Mean Specific Heat, Water Vapor. and representing the degrees of superheat (fe — £») by D, as is customary, this becomes Total heat per pound = q-\-r-\-C pm D. (18) 26. Specific Volume of Dry Saturated Steam, V or S. The volume occupied by one pound of a substance is spoken of as the specific volume of that material. In the case of dry saturated steam there are as many specific volumes as there are pressures under which the steam can exist. These values are generally tabulated in steam tables and are represented by the letter V or the letter S. STEAM 37 The values of the specific volumes of steam at different pressures are given in Fig. 15. It is important to note the very gradual change of specific volume at high pressures and the very rapid change and enormous increase at low pressures. These facts have considerable influence on steam engineering practice. DZV M .480 440 £400 a |360 p ^320 .§280 g £240 2 3 200 to CO \ \ p: 160 \ \ v 120 \ \ \ \ .80 \ V. 40 2 4 (3 8 10 12 14 16 18 20 22 Specific Volume of Dry Saturated Steam (Cubic Feet) Fig. 15. — Pressure-Volume Relations, Saturated Water Vapor. A curve giving properties of saturated steam is called a saturation curve, so that this name may be, and often is, applied to the curve given in Fig. 15. The volume occupied at any pressure by half a pound of dry saturated steam will obviously be half that occupied by one pound of such material at the same pressure, and 38 STEAM POWER the same statement can be made for any other fraction of a pound. It follows that if the small volume occupied by liquid water in wet steam be neglected, the volume occupied by one pound of steam (mixture) of 50 per cent quality can be assumed equal to half that occupied by an equal weight of dry saturated steam at the same pressure. A similar statement could of course be made for any other quality and a corresponding fraction. Hence if one pound of " wet steam " at a given pressure is found to have such a volume that it would be indicated by point b in Fig. 16, the quality of this material must be given by the expression x = — -if the volume occupied by the ac liquid water in the mixture be neglected. 27. Specific Density of Dry Saturated Steam, - or 8. The weight per cubic foot of saturated steam is spoken of as its specific density. The specific density is obviously the reciprocal of the specific volume and is there- Fig. 16. — Determining Quality from Volume. f re — 28. Reversal of the Phenomena Just Described. If any process which has resulted in the absorption of a quantity of heat by a substance be carried through in the reverse direction, the same amount of heat will be liberated. It follows that a pound of dry saturated steam will give up the total latent heat of vaporization when condensed to liquid at the same temperature, and that the resultant pound of hot water will give up the total heat of the liquid if cooled to 32° F. 29. Generation of Steam in Real Steam Boiler. The steam boiler is equivalent to a vessel partly filled with water STEAM 39 and fitted with means for supplying heat to the water and for carrying off the vapor formed. This is shown diagram- matically in Fig. 17. At first glance this would not seem to be at all similar to the cylinder and piston already con- sidered, but it really is the exact equivalent so far as the generation of steam is concerned. The flow of steam out of the steam-pipe is restricted to the extent necessary to maintain a high and constant pressure within the boiler, and, when in regular operation, steam is formed within the Fig. 17. — ■Foni.ation of Steam in a Steam Boiler. boiler under this pressure just as fast as necessary to replace that flowing out. By picturing the steam as flowing out in layers or lamina these lamina can be imagined as taking the place of the piston in the apparatus of Fig. 11, and each pound of steam formed will then push a piston before it exactly as was assumed in the previous discussion. 30. Gauge Pressure. The steam pressure in a boiler is commonly determined by means of an instrument called a pressure gauge. These instruments are almost always con- structed about as shown in Fig. 18 (a) and (6) . The Bourdon spring is a tube of elliptical section bent approximately into 40 STEAM POWER the arc of a circle. One end of this tube is connected directly to the pressure connection of the gauge and the other end is closed and connected to a toothed sector as shown. When the pressure inside a tube of this character is increased, the tube has a tendency to unroll or straighten out, and in so doing it moves the toothed sector in such a way as to rotate the pointer or gauge hand and make its end move over the scale in the direction of increasing pressure. With diminishing pressure the tube again rolls up and rotates the hand in the opposite direction. Treasure Connection Pressure Connection (a) (6) Fig. 18. — Bourdon Pressure Gauge. Instruments of this kind are so made and adjusted that the hand points to zero when the gauge is left open to the atmosphere. Under such conditions the pressure inside the tube is equal to that of the atmosphere and is not zero. The gauge therefore only indicates pressures above atmos- pheric on its scale, and the total pressure inside the boiler is really that shown by the gauge plus that of the atmos- phere. Pressures as indicated by the gauge are called gauge pressures. Pressures obtained by adding the pressure of STEAM 41 the atmosphere to the reading of the gauge are known as absolute pressures. Then Absolute Pressure = Gauge Pressure -{-Atmospheric Pressure and Gauge Pressure = Absolute Pressure — Atmospheric Pressure. In accurate work the existing atmospheric pressure should be determined by means of the barometer, but for ordinary, approximate calculations and for cases in which no barometric data are available, it is customary to assume the pressure of the atmosphere to be equal to 14.7 lbs. per square inch. This is very nearly true, on the average, at sea level, but is generally far from true at higher elevations. PROBLEMS 1. Determine by means of the steam tables the temperatures, total heats, heats of liquid, internal and external latent heats, and the specific volumes of 1 lb. of dry, saturated steam under the fol- lowing absolute pressures (lbs. per sq. in.) : 15, 50, 95, 180 and 400. 2. Determine the heats of the liquid, latent heats of vapor- ization and total heats for 2 lbs. of dry saturated steam at the following temperatures in °F.: 101.83, 212 and 327.8. 3. Determine the volumes occupied by 2 lbs. of dry saturated steam under the conditions of problem 2. 4. Determine the heats of the liquid, latent heats of vapor- ization and total heats for 1 lb. of saturated steam with a quality of 90% at the following absolute pressures: 25, 50, 75, 125. 5. Determine the total heat above 32° F. in 12 lbs. of saturated steam with quality of 97% at a pressure of 125 lbs. per square inch absolute. 6. What space will be filled by 20 lbs. of dry saturated steam at a pressure of 150 lbs. per square inch absolute? 7. What space will be filled by 20 lbs. of saturated steam at a pressure of 150 lbs. per square inch absolute and with a quality of 95% if the volume occupied by the water present be neglected? 8. How many pounds of dry saturated steam at a pressure of 75 lbs. per square inch absolute will be required to fill a spaceof 10 cu. ft.? 42 STEAM POWER 9. How many pounds of saturated steam with quality 96% and at a pressure of 1 10 lbs. per square inch absolute will be required to fill a space of 8 cu. ft.? 10. How much external work, measured in B.t.u., is done when 1 lb. of water at the temperature of 212° F. is converted into dry saturated vapor at the same temperature? 11. How much external work, measured in foot-pounds, is done when 2 lbs. of water at a temperature of 212° F. are converted into 90% quality steam at the same temperature? 12. How much heat is required for doing internal work during the vaporization of 1 lb. of water under such conditions that the total latent heat of vaporization is 852.7 B.t.u. and the external latent heat is 83.3 B.t.u.? 13. What is the quality of steam containing 1000 B.t.u. above 32° F. per pound when under a pressure of 150 lbs. per square inch absolute? 14. Heat is added to 1 lb. of mixed steam and water while the pressure is maintained constant at 100 lbs. per square inch absolute. The percentage of steam in the mixture is increased thereby from 50% to 95%. (a) How much heat was added? (b) How much internal latent heat was added? (c) How much external latent heat was added? 15. How much heat is required to completely vaporize 1000 lbs. of water at a temperature of 92° F. when pumped into a boiler in which steam is generated at a pressure of 150 lbs. per square inch gauge? Note that heat above 32° F. in 92° F. water is given as q in steam tables for a temperature of 92° F. 16. Find the amount of heat necessary to produce in a boiler 200 lbs. of steam having a quality of 97% at a pressure of 100 lbs. gauge when the feed water has a temperature of 205° F. 17. What volume would be occupied by the material leaving the boiler in problem 16, neglecting volume occupied by water? CHAPTER IV THE IDEAL STEAM ENGINE 31. The Engine. If the cylinder and piston assumed in the discussion of the last chapter be imagined as turned into a horizontal position and fitted with a frame, piston -Fly wheel Fig. 19. — Simple Steam Engine. rod, crosshead, connecting rod, crank shaft and flywheel as in Fig. 19, a device results which might be used as a steam engine for the production of power. By adding heat to, and taking heat from, the water and steam in the cylin- der in the proper way and at the proper time, the water and steam, or working substance, can be made to do work upon the piston. The piston can transmit this work through the mechanism to the rim of the flywheel, and it can be taken from the rim by a belt connected to a pulley on a machine which is to be driven, 43 44 STEAM POWER To make the analysis easier, a simplified type of engine will be assumed. It is shown in Fig. 20 and consists of the same cylinder, piston and piston rod as just described. A wire is fastened to the end of the piston rod and run back over a pulley in such a way that a weight fastened to the free end of the wire will be raised if the piston moves out. The weight is made up of two parts, one large and one small. When both are on the wire the pull which they exert causes the piston to exert a high pressure upon whatever is con- Cylinder Volume Fig. 20. — Simplified Steam Engine. tained in the cylinder. When only the small weight hangs on the wire, the piston exerts a much lower pressure upon the material in the cylinder. Imagine that the piston and the walls of the cylinder are made of some ideal material which will not receive or conduct heat. Imagine also that the cylinder is fitted with a permanent head which is a perfect conductor of heat. These conditions are of course ideal but are assumed for the sake of simplicity. Assume further that, when one pound of water is con- THE IDEAL STEAM ENGINE 45 tained in the cylinder and the piston is driven into the cylinder by the two weights until the space between the piston and the cylinder head is just large enough to contain the pound of water, the piston exerts a high pressure equal to Pi pounds per square foot against the water. The volume of this water and the pressure upon it can be represented by the point a of the PV diagram, Fig. 20. 32. Operation of the Engine. With conditions as de- scribed in the preceding paragraphs, imagine a flame or other source of heat at high temperature to be brought into contact with the conducting cylinder head and to pass heat into the cylinder, raise the temperature of the water within to the temperature of vaporization and ultimately vaporize it. As the water vaporizes it will push the piston out of the cylinder just as described in the last chapter and a hori- zontal line such as ab in Fig. 20 will represent the increase of volume (vaporization) at constant pressure. The point b may be assumed to represent the volume of one pound of dry saturated vapor at a pressure P\. Obviously the steam, as it is formed, does work in driving out the piston against the resistance offered by the weights which must be raised. If a stop is provided which will prevent the movement of the piston beyond the position corresponding to the point b, it will be possible to remove the larger weight when that point is reached and the high pressure steam will hold the piston and rod hard against the stop. If now some cooling medium is applied, such as a large piece of ice held against the conducting head of the cylinder or water running over that head, heat will be abstracted and a partial con- densation of the steam within the cylinder will occur. As condensation progresses the pressure will drop because there will be less and less steam, by weight, in a given volume. Such a process, would be indicated by the line be which represents a drop of pressure, while the volume 46 STEAM POWER contained within the cylinder walls between head and piston remains constant. When some point c is reached, the steam pressure will have been reduced to a value equal to that exerted by the small weight, and the piston will be driven in toward the cylinder head while the heat absorbing medium continues to remove heat from the steam and to cause further conden- sation. The combination of piston motion and heat ab- sorption will be so regulated that the pressure remains con- stant at P 2 during this process, because the weight will move the piston inward just as fast as necessary to main- tain a constant pressure. If sufficient heat is absorbed, the pound of material within the cylinder will ultimately all be condensed or liquefied and will just fill the volume Va. The heat absorbing body may now be removed and an infinitesimal motion of the piston toward the head would serve to raise the pressure on the liquid water from P 2 to Pi so that the volume Va may be taken equal to the volume V a and the line da may be assumed to be vertical. It would then represent an increase of pressure at constant volume. This might be caused by hanging a weight of the larger size on the wire when condition d was reached. Having brought the material, or working substance, back to the conditions originally shown at a, the high temperature source of heat can again be brought in contact with the end of the cylinder and the entire cycle carried through once more. There is obviously no reason why it could not be repeated as often as desired. 33. Work Done by the Engine. If the device just described is to serve as a steam engine, it must actually make mechanical energy available, that is, it must convert into mechanical form some of the heat energy supplied it. It is now necessary to see whether it does so. Water vaporizing and increasing in volume as from V a to Vb was shown in the last chapter to do work upon the piston confining it. Work has been shown to be equal to THE IDEAL STEAM ENGINE 47 (total force X total distance) and in this case if L repre- sents the distance in feet traveled by the piston, the work done by the steam upon the piston while the latter moves from a' to b' must be Work done on piston = total force X distance = PiXarea of piston XL . . . ft.-lbs. But the product of area of piston in square feet by distance traveled in feet is equal to the piston displace- ment or volume swept through by the piston, that is (Vb— Va) cubic feet. Therefore Work done on piston = Pi (V b - V a ) ft.-lbs. . (19a) Pl(V b -V a ) 778 B.t.u. . (196) The first form of this expression Pi(Vb—V a ) is very obviously represented by the area under the line ab in Fig. 20 and this area therefore represents the work done by the steam upon the piston during the change of volume at constant pressure represented by that line. While the steam is supplying this amount of energy to the piston or doing this amount of work upon the piston, the latter does an equivalent amount of work upon the weights if friction- less mechanism be assumed. In such a case the total weight hung on the wire multiplied by the distance raised would therefore give the same result in foot-pounds as that just obtained. It should be noted that Eq. (19) is merely an expression of the external work done during vaporization, that is, an expression of the amount of heat which is used for the doing of external work. It is the exact equivalent of the external latent heat previously discussed. In fact, the group of symbols APu is really a condensation of Eq. (19) formed by putting A for ^^ and u for (V b — V a ). 48 STEAM POWER The line cd also represents a change of volume at con- stant pressure and the same type of formula as applied to ab will express the work done during this process. In this case, however, the piston is being pushed into the cylinder by the small weight against the pressure of the steam, and energy is being supplied to push the piston in. This energy is equal to the weight of the small weight (pounds) multiplied by the distance it falls (feet). The piston is therefore doing work upon the steam, and the amount is Work done on steam = P 2 (V e -V d ) ft. -lbs. . . (20) P2(Vc-V d ) 778 B.t.u. . . (21) The first form of expression also represents the area under the line cd and this area therefore represents the work done by the piston upon the steam mixture in the cylinder during the process represented by cd. No work can be done by steam on piston or by pistoi? on steam during the processes represented by be or da because both the weights and the piston are stationary during these changes and it has already been shown that work involves motion. The total work done upon the piston by the steam is therefore represented by the area abef and this amount of energy is used in raising the two weights through a vertical distance equal to the piston travel. Some of this energy, or its equivalent, will have to be returned an instant later, however, in order that the piston may do the work shown by the area cdfe upon the steam. It is returned by the small weight dropping through a distance equal to the travel of the piston. The net mechanical energy made available by carrying through the series of processes is therefore represented by the area (abef) — (cdfe) = (abed) or the area enclosed by the four lines representing the THE IDEAL STEAM ENGINE 49 pressure and volume changes experienced by the working substance during one cycle of events. It is equal to the work done in raising the larger weight a vertical distance equal to the travel of the piston. This net energy made available is obviously Energy made- available = Pi ( V b — V a ) — P2 ( V c — Va) = (Pi-P 2 )(n-K)ft.-lbs. (22) (Pi-P 2 )0V-F.) 778 B.t.u. (23) Since this amount of energy is made available while one cycle of events is being carried out and since the cycle can be repeated time after time if sufficient heating and cooling mediums are available, any quantity of mechanical energy can be produced from heat energy by repeating the cycle a sufficient number of times. This would correspond to picking up a number of the larger weights which were slid on to the wire at the lower elevation and slid off at the higher. This repetition of cycles would correspond, in a real engine, to running at such a speed that the required number of cycles would be produced in a given time to make avail- able the amount of mechanical energy required. Or, the power made available per cycle could be increased. This is easily seen by an inspection of Eq. (22). Increas- ing the value of either of the right-hand terms will obviously increase the amount of energy made available. The value of (P1 — P2) can be increased by raising the initial pressure Pi or by lowering the final pressure P2. The value of (Vb— Va) may be increased by using more than one pound of material, thus increasing both the volume Vb of the satu- rated steam formed and increasing the volume V a of the liquid water, but getting a greater numerical value for (Vb— Va). This would correspond in a real case to using a larger cylinder and therefore a larger engine. 50 STEAM POWER ILLUSTRATIVE PROBLEM An engine of the type described is to work with a maximum pressure of 100 lbs. per square inch absolute and a minimum pressure of 15 lbs. per square inch absolute. The cylinder is to be of such size that 1 lb. of water is used and the steam is to be dry and saturated at the point b of the cycle. Find: (a) the amount of mechanical energy made available per cycle; (b) the amount of energy made available per minute if 150 cycles are produced per minute; and (c) the horse power of the engine. It will first be necessary to find the piston displacement required and the space necessary between piston and cylinder head to accommodate the pound of water in liquid form. The steam tables give the volume of one pound of dry saturated steam at 100 lbs. per square inch as 4.429 cu.ft. and the volume of one pound of water may be taken as 0.017 cu.ft. The values of the various volumes and pressures will therefore be Va = V d = 0.017 cu.ft,; V b = Vc= 4.429 cuit.; p a =p b = 100X 144 = 14,400 lbs. per sq.ft.; p c =p d = 15X144 =2160 lbs. per sq.ft. (a) Using Eq. (22) the amount of mechanical energy made available per cycle will be (Pj -P 2 )(V b - V a ) = (14,400 -2160) (4.429 -0.017) = 12,240X4.412; = 54,002.88 ft.lbs. (6) If 150 cycles are produced per minute, the total amount of mechanical energy made available per minute must be 150X54,002.88 =8,100,300 ft.-lbs. (c) The horse power must then be , 8,100,300 h - p - = ^ooo- =245+ - 34. Heat Quantities Involved. It is a very simple matter to determine the quantity of heat which must be supplied to produce the process ab, and the quantities of THE IDEAL STEAM ENGINE 51 heat which must be removed to produce the processes be and cd, This can be done by making use of the known properties of water and steam as given in the steam tables. The water at d must be at the temperature of vaporiza- tion corresponding to pressure P2 since it has just been formed by condensation from steam under that pressure. It therefore contains the heat of the liquid corresponding to that pressure. If it is to be vaporized at pressure Pi, it must first be raised to the higher temperature corresponding to that pressure. The amount of heat required to do this will obviously be the difference between the heat of the liquid at the temperature corresponding to Pi and the heat of the liquid at the temperature corresponding to P2, These can be found in the steam tables. The latent heat of vaporization at Pi must then be added to cause the increase of volume shown by ab. This can also be found in the steam tables for any given case. The quantity of heat which must be removed to produce the processes represented by be and cd can be found sim- ilarly from steam table values, although the exact method of procedure is not quite as obvious as in the preceding cases. Assuming that it is possible to find the heat supplied, Qi, and the heat removed, Q2, it is obvious that the energy made available in mechanical form, per cycle, must be equal to (Q1 — Q2) B.t.u., since this is the amount of heat energy which has disappeared and since it cannot have been destroyed. This may be put in the form of an equa- tion, thus Energy made available = Qi — Q2. . . (24) If the proper substitutions are made in this formula and it is then simplified, it becomes Energy made available = (APu) Pl — x c (APu) P2 B.t.u., (25) 52 STEAM POWER in which (APa) Pl = the external latent heat at pressure Pi; (APu)p 2 = the external latent heat at pressure P2, and x c = quality at point c, which can be found from the ratio of dc to dc'\ Numerical substitution in this equation for any given case will show that it gives exactly the same values as would be obtained by the use of Eq. (23). It is to be noted particularly that the energy made available is actually less than the external latent heat at the higher pressure, while the heat supplied must be equal to the total latent heat plus some of the heat of the liquid. An inspection of the steam tables will show that the exter- nal latent heat for ordinary steam pressures forms a very small fraction of even the total latent heat, and therefore the mechanical energy made available for a given expendi- ture of heat energy is very small in the case under dis- cussion. 35. Efficiency. The term efficiency is used in engineer- ing as a measure of the return obtained for a given expendi- ture. It may be defined in any one of the following ways: ^^ . Useful result Efficiency = Expenditure made to obtain that result Result Effort Output Input (26) In the case of a heat engine, the useful result is the mechanical energy obtained by the operation of the engine, while the expenditure made is the heat which is supplied. For this case efficiency may therefore be denned by the expression THE IDEAL STEAM ENGINE 53 _ . ^ . Mechanical energy obtained per cycle Engine efficiency = ^— ^—z f Heat supplied per cycle =£ ™ =^ • • • • <*> in which E stands for mechanical energy obtained, Qi stands for heat supplied, and Q2 stands for heat rejected. In the case of the type of steam engine just considered, this efficiency would have a value between 6 and 8 per cent for ordinary pressures. That is, the engine would produce in mechanical form only 6 to 8 per cent of the energy supplied it in the form of high temperature heat. Moreover, these figures would hold only for a theoretically perfect engine; a real engine built to operate upon this cycle would probably give efficiencies of the order of 2 to 3 per cent. The reasons for this great discrepancy will be discussed in a later chapter. 36. Effect of Wet Steam. In what has preceded, it was assumed that the pound of steam was completely vapor- ized along the line ab so that dry, saturated steam existed in the cylinder at b. It might, however, be assumed that vaporization was incomplete at the upper right-hand corner of the cycle, so that this corner occurred at a point to the left of b in Fig. 20 and with a quality x less than unity. Under such conditions, the cylinder would not have to be so big, since the maximum volume attained by the steam would be smaller than in the preceding case. The work done per cycle would obviously be smaller in quantity, because the area enclosed within the lines of the cycle would be smaller. It can also be shown that the efficiency would be lowered by incomplete vaporization, 54 STEAM POWER 37. Application to a Real Engine, The engine which has been described in the preceding paragraphs could easily be converted into the counterpart of a real engine by sub- stituting connecting rod, crank shaft and flywheel for wire, pulley and weights as described in the first paragraph of this chapter. It could then be made to do work in just the same way as has been described; some of the energy made available during the outstroke would be used for overcoming resistance at the shaft, that is, doing useful work, and some of it would be stored in the flywheel which would speed up slightly. The energy which must be expended on the steam during the return stroke would be obtained by allowing the flywheel to slow down and thus deliver suf- ficient kinetic energy to drive the piston back against the low-pressure steam. The cycle and the efficiency would thus, theoretically, be exactly the same as those just in- vestigated. Great difficulty would, however, be met in a real engine if the steam had to be formed and condensed within the cylinder, and another method which gives the same results is therefore used. Steam is generated in a boiler and allowed to flow into the cylinder and push out the piston just as though it were actually being formed in the cylin- der as previously described. When the piston reaches the end of its outstroke the inlet valve is closed and the exhaust valve is opened, allowing some of the steam to blow out into a space in which a lower pressure exists. As the piston stands still at the end of its stroke while the pres- sure drops, the line be is produced as in the previous descrip- tion, but by a different method. The piston then returns and drives the remaining steam out of the cylinder at a constant pressure theoretically equal to that of the space into which the steam is being forced or exhausted. The line cd is thus produced and the closure of the exhaust valve and opening of the admission valve when d is reached will start the cycle over again. THE IDEAL STEAM ENGINE 55 In order to get more work out of a given size of cylinder and to obviate the necessity of giving back energy which has already been given out, engines are generally made to take steam on both sides of the piston. They are then known as double acting engines. In this case the steam admitted on one side of the piston would supply the energy necessary both for overcoming the resistance due to the load and for driving out the low-pressure steam on the other side of the piston. On the return stroke conditions would be just reversed. 38. Desirability of Other Cycles. The cycle of opera- tions described in preceding paragraphs is the most inef- ficient of all those actually used, that is, it gives the small- est return for a given amount of heat supplied. This is because only the external latent heat supplied is con- verted into mechanical energy and part of that energy must be returned to complete the cycle. All of the heat of the liquid as well as all the internal latent heat supplied along ab passes through the engine without conversion and is ex- hausted. Therefore, cycles which differ from that described in such a way as to make it possible to convert into mechani- cal energy some of the internal latent heat and possibly some of the heat of the liquid should be highly desirable as they ought to yield a larger return of mechanical energy for the same total amount of heat supplied. Two such cycles are commonly used; they may be described as the Complete-expansion cycle and the Incomplete-expansion cycle. The former is used in steam turbines, the latter in most reciprocating steam engines. The rectangular cycle which has just been described is used in duplex pumps and similar apparatus. 39. The Complete-expansion Cycle. This cycle, which is also known as the Clausius and as the Rankine cycle, starts just the same as that already described. This is shown in Fig. 21. The pressure on, say, a pound of water 56 STEAM POWER is raised from P2 to Pi and its temperature is raised from that of vaporization at P2 to that of vaporization at Pi. After this it is vaporized, giving the increase of volume Coiled Spring b \c[ t h K'h A \ A *_ A \ /' \ / \ / \ }, 1 2 1 }&i j J Volume p' | 6 ' ic £.- ■ s I p 2 « 1 1 6 /V/sS Saturation Curve /v/y/AS. \ Work done by Steam ///w/W)v S /. ( luriag expansion 6-G d WMMiMMk Volume Fig. 21. — Complete Expansion Cycle or Clausius Cycle. shown by ab. The supply of heat is then stopped. The cylinder of the engine is made larger than in the preceding type so that when the point b' is reached the piston can travel still further, and it is allowed to do so, that is, the THE IDEAL STEAM ENGINE 57 high-pressure steam is allowed to push it further out. This can be pictured by imagining the steam to act like the compressed spring shown in the figure and to push the piston in much the same way as does the spring. The line bici shows the decreasing pressure exerted on the piston by the spring as the latter expands so as to get longer and longer. Because of the properties of a spring this is a straight line. The line be shows the decreasing pressure exerted on the piston by the steam as the latter expands so as to occupy greater and greater volumes. Because of the properties of steam this line is curved instead of straight. Work will be done on the piston by the expanding steam during the process be and the amount of this work will be indicated by the area under the line be as shown in the figure. This work must have been done by the expenditure of energy on the part of the steam and since no energy was added after the point b was reached the work must have been done at the expense of heat energy contained in the steam at b. It has already been shown that the heat above 32° in the steam at b is equal to the sum of the heat of the liquid and the internal latent heat, and some of this heat must obviously be used for the doing of work along be instead of being entirely rejected to the cooling medium as in the preceding cycle without " expansion." The expansion of the steam continues until the " back pressure " P2 is reached. The cooling medium may then be imagined to be brought into use and to abstract such heat of vaporization as may remain in the steam besides absorbing the equivalent of the work done on the steam by the returning piston, thus giving the process shown by the line cd. If the expansion line be of the cycle just described could be carried out within walls constructed of such mate- rial that it would not give heat to nor take heat from the steam, it is obvious that any heat energy lost by the steam during the expansion could be lost only by conver- 58 STEAM POWER sion into mechanical energy. An expansion of this kind is called an adiabatic expansion. In the figure, the curve of adiabatic expansion is shown in its correct position with respect to the saturation curve and it is obvious that for an adiabatic expansion, starting with dry, saturated steam, the quality decreases as the expan- sion progresses. Comparison with- Cycle without Expansion. The heat supplied is the same in both of the cycles just considered when they operate between the same two pressures, but the mechanical energy obtained in the case of the complete expansion cycle is much greater. In Fig. 21, for instance, the mechanical energy obtainable with the cycle first described is represented by the area abd'd while that obtain- able with the complete expansion cycle with the same heat supply Qi is represented by the same area abd'd plus the additional area bed' . The efficiency of the complete expansion cycle is therefore very much higher than that of the cycle without expansion. For conditions similar to those giving a theoretical efficiency of about 6 per cent without expansion, the com- plete expansion cycle will give a theoretical efficiency of about 12 per cent and this figure can be doubled by expedients which will be considered later. The cylinder required for the production of the com- plete expansion cycle would be much larger than that re- quired for the other cycle if both used the same weight of steam per cycle. The proportion would be in the ratio of the volume shown at c in Fig. 21 to the volume shown at b. But the complete expansion cycle would make avail- able much more energy per pound of steam than would the other, so that the difference in the size of cylinders would not be so great if both were required to make avail- able the same amount of mechanical energy per cycle. 40. The Incomplete-expansion Cycle. The shape of this cycle is shown in Fig. 22. It is just like the complete THE IDEAL STEAM ENGINE 59 Fig. 22. — Incomplete Expansion Cycle. expansion cycle down to the point c. The cylinder in which it is produced has a smaller volume than that used for the complete expansion cycle so that the piston arrives at the end of its stroke before it has opened up volume enough to enable the steam to expand all the way down to the lowest pressure (terminal or back pressure) . When the point c is reached in the real engine, the exhaust valve is opened and enough steam then blows out to reduce the pressure to the back pressure P a . The piston then returns and drives out the remainder of the steam as shown by the line de. In the ideal method assumed in the preceding treat- ment, the heat absorbing medium would be brought into use at c, absorbing sufficient heat to reduce the pressure from P c to Pa while the piston remained stationary at the end of its stroke. The latent heat of vaporization remain- ing in the steam at d would then be absorbed as the piston was driven back from d to e. Comparison with Other Cycles. The incomplete expan- sion cycle is intermediate between the two previously dis- cussed. This can be appreciated readily by an inspection of Fig. 22. In this figure the area abd'e represents the mechanical energy obtainable with the cycle without expansion; the area abc'e represents the energy obtainable from the same quantity of steam with complete expansion; and the area abcde represents the energy obtainable from ihe same amount of steam with incomplete expansion. The later the point at which the exhaust valve is opened, point c, the more nearly do efficiency and energy obtain- able approach the values for the complete expansion cycle. 60 STEAM '.POWER The earlier the point at which the exhaust valve is opened, the more nearly do efficiency and energy obtainable approach the values for no expansion. Despite the lower efficiency of the incomplete expan- sion cycle as brought out in connection with Fig. 22 it is universally used on all reciprocating engines excepting those which make do pretense to economy and use no expansion. The less efficient cycle is used for the simple reason that complete expansion in a reciprocating engine does not pay commercially. For complete expansion the cylinder must be larger in the ratio of V c to V c > as shown in Fig. 22 and the work obtained by completing the expan- sion is a very small part of the total. In most cases it would not be great enough to overcome the friction of the engine, not to mention paying interest on the necessarily higher cost of the larger cylinder and accompanying parts. It will be shown in a later chapter that the steam tur- bine can economically expand the steam completely and the complete expansion cycle is therefore used with such prime movers. CHAPTER V ENTROPY DIAGRAM 41. Definitions. In Chapter III temperature, pressure and volume were discussed as criteria determining the con- dition of water and steam. Other things may be used in determining the condition of such materials. One which is particularly useful from an engineering standpoint is known as entropy and is designated by the Greek letter cj>. For every condition of water and steam, there is a char- acteristic value of entropy just as there is a characteristic value of temperature, pressure, volume, heat above 32° F., etc. These values of entropy are given in the steam tables in just the same way as the value of temperature, pressure, volume, heat above 32° F., and such, are given. The entropy of the liquid given for any particular pres- sure is the change of entropy experienced by one pound of the liquid when its temperature is raised from 32° F. to the temperature of vaporization corresponding to that particular pressure. It might be spoken of as the entropy of the liquid above 32° F., just as q is spoken of as the heat of the liquid above 32° F. It is represented by i. The entropy of vaporization given for any particular pressure is the change of entropy experienced by one pound of the material while changing from water at the tempera- ture of vaporization to dry saturated steam at constant pressure. It corresponds to the latent heat of vaporiza- tion and is designated by fa. The entropy of dry saturated steam at any pressure is the sum of fa and fa and therefore is the total change of entropy experienced by a pound of material in changing 61 62 STEAM POWER from water at 32° F. to dry saturated steam at the particu- lar pressure in question. The entropy of superheat at any pressure and tempera- ture is the change of entropy experienced by a pound of dry, saturated steam at that pressure when superheated to that particular temperature. It is designated by s . The entropy of superheated steam at any pressure and temperature is the total change of entropy experienced by one pound of material when changed from water at 32° F. •So/ I *-} i,/ i Wet Steam Region i o/ 2W- •£/ JRegion of Incomplete! Vy^ ' | Evaporation Entropy Scale («) Fig. 23.— Temperature-Entropy Diagrams. to superheated steam at the pressure and temperature in question. It is equal to 0 opposite the horizontal axis. The various areas hatched in Fig. 23 (b) indicate the various quantities of heat previously discussed. It should be understood that the areas represent the heat quantities only for the particular pressure which corresponds to the temperature indicated by T v . For a higher pressure, the line be would be higher and the areas proportionately larger; for a lower pressure the line be would be lower and the areas smaller. ENTROPY DIAGRAM 65 ILLUSTRATIVE PROBLEM Starting with liquid at a temperature Tx corresponding to the temperature of vaporization at a pressure of 50 lbs. per square inch absolute, assume the liquid raised to the temperature of vaporization at a pressure of 100 lbs. per square inch absolute and then completely vaporized. Determine the various changes of entropy and indicate them on a T^-chart. The steam tables give entropy of the liquid, 4>i, as equal to 0.4113 for water about to vaporize under 50 lbs. per sq. in. absolute, and 0.4743 for water about to vaporize under a pres- sure of 100 lbs. per sq. in. absolute. The difference, that is, 0.4743-0.4113=0.0630, must be the entropy change experienced by the liquid when its tempera- ture is raised from the lower to the higher value. These values hie shown in Fig. 24. The steam tables give entropy of vaporization, v . In general, if a fraction x of the latent heat has been added, the entropy change has been x„ during the process. Therefore, if the temperature entropy condition of a pound of material should plot at a Fig. 24. 66 STEAM POWER point such as c in Fig. 25, it follows that the material is a mixture of water and steam and that a fraction of the be pound equal to — is steam, the rest being water. But, be by definition, the fraction — is x, the quality of the material. The temperature-entropy chart is very useful when used in connection with this property of showing quality. Thus, in Fig. 25, the area under be, down to absolute zero tem- perature, represents the fraction of the latent heat of Entropy Fig. 25. — Quality from Temperature- Fig. 26. — Constant Quality Entropy Chart. Curves. vaporization per pound which must be added to give a pound the quality x. For convenience in use, constant quality lines are generally drawn on temperature-entropy charts. Such lines are shown in Fig. 26. Each line is obtained by plot- ting the temperature entropy conditions for a given quality at different pressures. For this purpose, „ and i are taken from the steam tables for a given pressure. The numerical value of <}>„ is then multiplied by the fraction re- presenting the chosen quality, say 0.9, and the product is added to 4>i, giving the total entropy above 32° F. for quality 0.9 at the particular pressure chosen. The same ENTROPY DIAGRAM 67 y ^ sv // yy / \ ^l*/ 3— . A-A-A \ 8-^ ^J\ A £\ ?A 8^ o. ^4}V5^^^ ^nH^^ ^^^X*^ < m 1 ^%\ooob£5: s 2 ^ 8|* X\ 1 X i X 1 & (•j ) aan^aadtnax a^nxosqv 68 STEAM POWER process is repeated with the same value of the quality, but with different pressures, until enough points have been secured to make it possible to draw a smooth line through them. 44. Volume from T^-chart. Since quality changes at any given temperature, or pressure, are accompanied by volume changes, it is possible to find a series of values for the quality of a pound of wet steam which will make that pound occupy the same volume at different temperatures. Having found the quality which will be necessary at a num- ber of different temperatures, the total entropy above 32° F. can be found for each case and these values can then be plotted on the TV-chart. Connecting the points so obtained would give what is known as a Constant Volume Line. Several of these constant volume lines are shown in their correct positions in Fig. 27. It will be observed that, for each volume, the quality must increase as temperature (and pressure) increases in order to maintain a constant value for the volume occupied by one pound of mixture. 45. Heat from T^-chart. Equations for obtaining the total heat above 32° F. for wet and for superheated steam were given in an earlier chapter. By means of these equations, it is possible to find a succession of values for quality and superheat which will give a pound of material any chosen heat content at different pressures. If the corresponding values of temperature and entropy are found and plotted, what is known as a Constant Heat Line results. Several of these lines are shown in Fig. 27. 46. The Complete T^-chart for Steam. A very com- plete, graphical representation of the properties of water and steam can be procured by combining in one diagram all of the lines discussed in preceding paragraphs. Such a diagram is generally spoken of as the T (^-diagram or the T4>-chart for steam. An example of such a diagram is given in Fig. 28. This chart is very useful, as it enables one to solve by '■"Ri -ajn;BJadtuax 1 ■5 f g 1 o s § m $5 1 1 3 1 O c - 5 3 o 8 ^ ■/ 3 -Sd a-/ pg o S © -Jj 8 o 1 ^H i-i r?"~~ -3m #' i ^'> Y 1 ^ ■$. ?§§ p 5 ML 7^7- ^L? T 3?" kfl #J >5 ^ 0p^ 4* ^w ?^> % ■Of y ^fe k _LL €- TEMPERATURE-ENTROPY DIAGRAM TO ACCOMPANY STEAM POWER C.F.HlRSHFELD AND T.C.ULBRICHT. (Published by John Wiley & Sons) Redrawn (with permission) from larger diagram in Peabody's Steam and Entropy Tables ( Wiley & Sons) 90, Jwr ^L ^k 80, 70 I - *£' V: ,-- Mr .6 ^ \ ^50 1 l*^ \ £)' TPnjv i2_ P RM 30_ I" 2 1 \ L^ • i&W "VXs\ et*\ c uPi — ^0 7Y j>\ -A A% ^ <7£ < ■ ,J> \ ' pF A-\-r- V \\ \\ \ 3 2trU te^ \ > V \S Vk-VU J-dyif ^ ^ s v ^ ^ 3ml v # -chart. 7. Assume a pound of mixed water and steam to have a qual- ity of 80% at a pressure of 200 lbs. per square inch absolute. Determine from the TV-chart the heat above 32° per pound of mixture and the volume occupied by the mixture. Determine also the quality attained if the pressure of the material drops to 20 lbs. per square inch absolute at constant entropy. How does the heat above 32° F. change during such a process? 8. Assume a pound of mixture as in Prob. 7, but with a quality of 30% at a pressure of 200 lbs. Find all quantities called for in that problem. 9. Assume a pound of material as in Probs. 7 and 8 above, but superheated 200° at a pressure of 200 lbs. per square inch absolute. Determine all quantities called for in Prob. 7. 10. Choose a point on the T^-chart at which a constant volume line intersects the saturation curve. Determine the change of quality, entropy and heat above 32° F., if the material drops to half pressure at constant volume. CHAPTER VI TEMPERATURE ENTROPY DIAGRAMS OF STEAM CYCLES 47. Complete Expansion Cycle. This cycle was con- sidered in Chapter IV and the PF-diagram was given there as Fig. 21. The diagram of this cycle drawn to ^-co- ordinates is shown in Fig. 29. The same letters are used to represent corresponding points in the two diagrams. The entropy change during the heating of the liquid is shown by the part of the liquid line between d and a, and the heat supplied during that process is represented by the area below the line da, measuring clear down to the absolute zero of temperature. The entropy change during vaporization is represented by the line ab and the heat supplied during the process is shown by the total area under that line. The adiabatic expansion of the steam is represented by the line be, such an adiabatic change fortunately being a con- stant entropy process and therefore easily drawn in this diagram. Obviously no heat is received or removed dur- ing this process, as there is no area under the line be. The entropy change during condensation is represented by the line cd and the heat rejected by the working sub- stance during this process is represented by the area under that line. 72 Fig. 29.— 7V-diagram, Com- plete Expansion Cycle. TEMPERATURE ENTROPY DIAGRAMS 73 48. Area of Cycle Representative of Work. It will be remembered that area under a line in the PF-diagram represents work in foot-pounds. That diagram, however, gives no indication of heat received or rejected and it is not possible to obtain any direct idea of efficiency from it. In this respect, the T^-diagram is much better. Area under the lines da and ab in Fig. 29 represents heat supplied the working substance. Area under the line cd represents heat rejected by the working substance. The difference between these two, or the area enclosed within the lines of the cycle, must therefore represent the heat converted into mechanical energy per cycle. This diagram therefore shows directly by areas the heat supplied, the heat rejected, and the heat converted into mechanical energy. Further, the ratio of the area representing heat converted into work, and the area repre- senting heat supplied must be the efficiency of the cycle. Remembering also that if the lines of the cycle are drawn upon a T0-chart such as that given in Fig. 28, all volume changes, heat contents and qualities at different points are shown without further work, it becomes evident that this form of representation is decidedly convenient and far superior to the pressure volume method. 49. Modifications for Wet and Superheated Steam. The complete expansion cycle is supposed to represent an idealization of what happens in a real prime mover. In real cases, however, the steam may arrive at the prime mover wet or superheated and it is desirable to investigate the method of representing such conditions as well as their effects. Wet steam corresponds to incomplete vaporization, i.e., a quality less than unity at the upper right-hand corner of the cycle. This might be shown for a given case by the location of the point b' in Fig. 29. The cycle would then be ab'c'd and a smaller amount of work would be obtained per pound of working substance as evidenced by the smaller area enclosed within the lines of the cycle, 74 STEAM POWER In the case of superheated steam, superheating occurs at constant pressure after vaporization is complete. This would be shown by the location of the upper right-hand corner of the cycle at some point b" on the constant pressure line which extends out from b. The cycle is now represented by abb"c"d and evidently has a different shape than it had in the preceding cases. Obviously the area enclosed within the lines of the cycle is greater than it was before and therefore more mechanical energy is obtained per pound of steam. 50. Incomplete Expansion Cycle. The only difference between the incomplete and complete expansion cycles is Fig. 30. — T<£-diagram, Incom- Fig. 31.— 7 T -diagram. Cycle plete Expansion Cycle. Without Adiabatic Expansion. the termination of the expansion in the former by means of a constant volume line. This is shown to T^-coordinates in Fig. 30 in which the incomplete expansion cycle is drawn in heavy lines over the one in which expansion continues to the back pressure. The constant volume line is seen to cut off a corner, thus reducing the area representing heat converted into work. The heat supplied in each case is measured by the area under the lines ea and ab. The efficiency of the cycle with incomplete expansion can therefore be seen to be less than that of the other cycle by simple inspection of the diagram. If the adiabatic expansion is terminated at a higher TEMPERATURE ENTROPY DIAGRAMS 75 pressure, as by the constant volume line c"d" in Fig. 30 r still more of the work area is lost, but the same quantity of heat is supplied, and therefore the efficiency is still lower than when the expansion terminated at c. Obviously as the point at which the adiabatic expansion is terminated moves nearer and nearer to b as shown in Fig. 31, the cycle becomes less and less efficient. If the constant volume line starts at b, there is no adiabatic expansion and the cycle becomes that previously considered as having a rec- tangular shape in the PF-diagram. This cycle has the shape indicated by abed in the T^-diagram of Fig. 31. Obviously it is least efficient of all as was previously shown by other means. 51. Effect of Temperature Range on Efficiency. It has already been stated (see p. 26) that heat engines receive heat at a high temperature, convert some of it into me- chanical form and discharge the remainder at a lower tem- perature. Inspection of the T^-diagram shows this very clearly, and, remembering that the area of the cycle measures the heat converted, these diagrams also show how raising the upper temperature (or pressure) or lowering the lower temperature (or pressure) will increase the efficiency. It can be seen readily that lowering the lower temperature will, however, be more effective in increasing the efficiency than raising the upper temperature. PROBLEMS 1. Draw a complete expansion cycle to T ^-coordinates for the following conditions (using T^-diagram for steam to get values); weight of working substance, 1 lb.; initial pressure, 125 lbs. absolute; quality at beginning of adiabatic expansion, 100% back pressure, 10 lbs. absolute. 2. Determine the following values for cycle drawn in Prob. 1 : (a) Entropy of liquid at beginning of vaporization ; (b) Entropy at beginning of adiabatic expansion; (c) Quality at end of adiabatic expansion; (d) Volume at end of adiabatic expansion ■ (e) Entropy at end of condensation. 76 STEAM POWEP 3. Show by measuring the area on TV-diagrams, the increase of efficiency resulting from the use of an initial pressure of 175 lbs. absolute and from the use of a terminal pressure of 2 lbs. absolute in place of the values given in Prob. 1. 4. Compare the efficiency of a complete expansion cycle with conditions as in Prob. 1 with a complete expansion cycle with same pressures but with a temperature of 500° F. at the beginning of the adiabatic expansion. 5. Draw an incomplete expansion cycle to TV-coordinates for the same pressures as in Prob. 1, but with adiabatic expan- sion ending at a pressure of 15 lbs. absolute. 6. Compare work and efficiency of the two cycles of Probs. 1 and 5 above. 7. Draw a cycle without expansion for the conditions of Prob. 1 to TV-coordinates and compare the work area with that obtained in Probs. 1 and 5. CHAPTER VII THE REAL STEAM ENGINE 52. Operation of Real Engine. In previous chapters the ideal steam engine was considered and several cycles upon which it might be operated were discussed. Real engines are built to operate on the same cycles, but because of certain practical considerations they only imperfectly approximate the ideal performance. Real engines must be built of iron and steel for practical reasons and these metals absorb, conduct and radiate heat so that certain heat interchanges between the working substance and engine and certain heat losses occur in practical operation. These were eliminated in the ideal case by simply assuming ideal materials not possessed of the characteristics of real metals. It is also practically impossible to generate steam in the cylinder of a real engine as was assumed to be done in the ideal case. Heat is practically obtained by the com- bustion of fuels, and the higher the temperature attained the better can the liberated heat be utilized in the genera- tion of steam. To subject the cylinder to such high tem- peratures and to control the heating and cooling as neces- sary to produce a number of cycles in rapid succession would lead to rapid wear and great practical difficulties. It has been found best to generate the steam in a boiler which is properly equipped for that purpose and then to transmit it with its contained heat to the engine, which is constructed in such a way as to utilize that heat to the best advantage. If the steam is to be condensed, as assumed in the ideal cases, it has also been found best to remove it from the 77 78 STEAM POWER cylinder and to condense it in a separate piece of apparatus properly constructed for that purpose. The entire arrangement which results from these prac- tical modifications in the case of a non-condensing engine I f///////////////////,...^ THE REAL STEAM ENGINE 79 is shown in Fig. 32. Steam is generated within the boiler at some constant pressure P\ and at the proper instant the admission valve at one end of the cylinder is opened, allow- ing steam to flow in and drive the piston outward. If there were no losses, this would be represented by some such line as ab at a height Pi on the PF-diagram of Fig. 33. Closing of the valve after the piston had moved part way out would cut off the further flow of steam, and, with con- tinued motion of the piston, the steam within the cylinder would expand. ^ „„ If no heat interchanges occurred, this expansion be would be adiabatic as in the ideal case. It will be observed that the two lines on the PF-diagram thus far produced represent equally well the corresponding two lines of the complete or incomplete-expansion cycles. The heat supplied in the boiler is the same as that supplied in the cylinder under the ideal conditions originally assumed, and the work under the line a 6 is equal to the external work done during vaporization just as in the ideal case. If difficulty is experienced in connection with the statement regarding external work, it is only necessary to picture the process in this way: Assume that each pound of steam formed in the boiler does the external work equivalent to APu by pushing the pound previously generated ahead of it as a piston, and that this motion communicated along the pipe from layer to layer results in pushing an equivalent weight (and volume) into the cylinder against the resist- ance offered to the piston's motion. When the piston arrives at the end of its stroke at the point c, the opening of the " exhaust valve," connecting the interior of the cylinder with the space in which the pressure P2 lower than P c is maintained, will permit some of the steam to blow out of the cylinder with the piston standing stationary at the end of its stroke. This would 80 STEAM POWER give a constant volume change similar to the correspond- ing line in the incomplete-expansion cycle. The return of the piston from d to e, with the exhaust valve still open, would force the remainder of the steam out of the cylinder and into the space in which the pres- sure P2 is maintained. The result, so far as the diagram is concerned, is obviously the same as in the ideal case, and if the steam were condensed within a proper vessel into which it exhausted (instead of being exhausted to atmos- phere), the result would also be the same so far as the shape of the diagram is concerned. The pressure P2 might, how- ever, be maintained at a lower value, thus giving a greater temperature range. The pressure rise ea within the cylinder would result directly from the opening of the admission valve and the admission of steam for the next cycle. But, if the working substance is to be returned to starting conditions as was dene in the ideal case, its pressure must also be raised to Pi and its temperature to a corresponding value. The pressure is raised in the case of condensing operation by means of the boiler feed pump, which picks up the condensed steam (condensate) and forces it into the boiler. The temperature of the working substance is raised by passing it through feed-water heaters or by heat absorbed directly from the heated water in the boiler. When operating non-condensing the working substance exhausted during the last part of each cycle is really thrown away by allowing it to mix with the atmosphere, but an equivalent quantity of water is fed to the boiler by the boiler feed pump and takes the place of the material lost by exhaust to atmosphere. This method of operating does not approximate the ideal as closely as does the con- densing method, but the discrepancy is not very great. 53. Losses in Real Installations. The diagram given in Fig. 33 was obtained by assuming the absence of certain practical losses and is considerably modified when real Pt THE EEAL STEAM ENGINE 81 apparatus is used. Thus the real engine, as shown in con- nection with Fig. 9, has clearance and operates with com- pression so that the clearance is filled with steam at a pres- sure indicated by the point a' in Fig. 34 when the admission valve opens. There is also always some drop of pressure along the steam pipe so that the pressure at the engine is lower than at the p,| — I — *^*5S!mEv£ boiler. Further, the admission g *?\\ /AdiataticExpanBtauoi g ^ A>/ all Material in Cylinder valve can never be made to give 1 d&q W\ «TnniaiijDryana (£ \\y. %V" -^Saturated. such a large opening into the \§^j cylinder that there is not a P j | Backer, measurable drop of pressure in FlG . 34 .-Theoretical and Real flowing through it. As a result indicator Diagrams. of these actions the highest pressure attained within the cylinder as indicated at point a in Fig. 34 is always lower than the boiler pressure Pi. As the piston of a real engine moves out it acquires a higher and higher velocity until it reaches a point near mid-stroke. The entering steam therefore must flow through the valve with increasing velocity if it is to follow up the piston and fill the cylinder, but this usually neces- sitates greater pressure drops as the piston moves out, so that the admission line generally slopes downward instead of being horizontal. There is also another phenomenon which causes this line to slope. The metal of cylinder, cylinder head and piston is in contact with comparatively low-temperature steam during the latter part of each cycle, and therefore acquires a lower temperature than that of the steam about to enter. Therefore, when the high- pressure steam enters the cylinder it gives up heat to the walls at a comparatively rapid rate, and, if initially dry saturated, this results in a great deal of condensation. Such condensation is called initial condensation. As the steam condenses after flowing into the cylinder and forms water occupying a negligibly small volume, 82 STEAM POWER it follows that steam must flow into the cylinder at a pro- portionately greater rate in order to fill the space vacated by the piston. But this results in an increased pressure drop and therefore would give a sloping admission line. When the piston has finally been driven out as far as desirable by the action of high-pressure steam, the admis- sion valve is closed, that is, cut-off occurs. This valve can not be closed suddenly; the closure is more or less gradual in all cases. As the opening becomes smaller it becomes increasingly more difficult for the steam to flow through and into the cylinder so that the pressure continues to drop at an increasing rate until the valve is finally closed. This gives the rounded cut-off shown at the point b. The loss of pressure during admission is generally said to be due to throttling or wire drawing, these terms being intended to convey the idea that the steam has to squeeze its way through the inlet openings with correspond- ing loss of pressure. When the cut-off has finally been completed, it leaves the end of the cylinder filled with a mixture of steam and water at steam temperature, and this mixture then expands as shown by the line be. At the beginning of the expansion the steam generally has a higher temperature than that of the surrounding walls and it therefore continues to give heat to those walls. Were the expansion adiabatic it would follow the dot-dash line in the figure, but, as the steam must not only convert heat into work, but must also supply heat to the walls, it condenses more rapidly than in the ideal case and its pressure and volume changes follow some such law as that indicated by the upper part of the curve be. As expansion continues, the pressure and temperature of the steam drop until some point is reached at which the temperature has become equal to that of the walls. Further expansion with drop of pressure and temperature results in reducing the temperature of the steam below THE REAL STEAM ENGINE 83 that of the walls, and then the direction of heat transfer is reversed, the hot walls giving heat to the cooler steam at an increasingly rapid rate. This heat causes re-evapora- tion of some of the water formed before and thus tends to increase the volume occupied by the material in the cylinder, with the result that the lower part of the expansion curve be approaches and generally crosses the curve which would have been attained by adiabatic expansion in non-conduct- ing apparatus. In many real engines the re-evaporation is so great that the steam is entirely dried and sometimes superheated before the exhaust valve opens. The exhaust valves of steam engines are always opened before the piston reaches the end of the stroke, as it is found necessary to do this if excessive losses are not to occur due to the difficulty of forcing the large volume of low- pressure steam through the exhaust passages. When opened early enough, the steam flows out in such quantity before the end of the stroke that the " back pressure " during the return or exhaust stroke is only a pound or two above that of the space into which the engine is exhausting. During all of the exhaust period, the steam is probably at a lower temperature than the walls to which it is exposed and re-evaporation probably continues in most cases until the closure of the exhaust valve. It seems probable that the steam retained in the cylinder after the closure of the exhaust valve is approximately dry, but little is really known regarding the quality of the clearance steam. The rise of pressure during compression has two bene- ficial effects : It helps to bring the moving parts to rest grad- ually, and it raises the temperature of the clearance steam and of the walls of the clearance space to values nearer that of the entering steam. Remembering that area on a PF-diagram represents work, it is easily seen that throttling losses and rounding of corners due to slow valve action (which cause a loss of 84 STEAM POWER diagram area) result in a loss of work. The fact that conden- sation also causes a great loss is easily shown. A given quantity of steam entering the engine with its supply of heat can, in the ideal case, do a certain amount of work at the expense of that heat. In the real case part of the heat is stored in the walls during the early part of the cycle, so that it is not available for the doing of work and is removed from the walls and carried out into the exhaust as unutilized heat during the later part of the cycle. The phenomenon can be pictured by imagining the steam as dropping some of its heat into a pocket in the walls of the cylinder when entering the engine and then picking it up again and carrying it out when leaving, so that the next charge of steam will have to fill the pocket again. The net result of condensation and re-evaporation is the obtaining of less work from a given quantity of steam than should be obtained, or the use of more steam than theoretically necessary for a given quantity of work. This effect is shown graphically by the two adiabatic expansion lines of Fig. 34. The initial condensation in real engines which are sup- plied with saturated steam generally amounts to from 20 to 50 per cent of all the steam supplied, so that it is evident that anything which will prevent part or all of this loss should do much to improve the steam consumption of engines. This subject will be discussed in more detail in later para- graphs and various methods of decreasing losses from this source will be considered. 54. Clearance. The term clearance is used in a two- fold sense; (a) to refer to mechanical clearance or the linear distance between the two nearest points of cylinder head and piston face when the piston is at the end of its stroke, and (6) to refer to volumetric clearance or the volume enclosed between the face of the valve, the cylinder head and the face of the piston when the latter is at the end of its stroke. THE REAL STEAM ENGINE 85 Steam Port Steam Chest Valve Fiston Fig. 35. — Mechanical and Volu- metric Clearances. The former is generally given in inches and varies from a very small fraction of an inch in the best engines to an inch or more in cheap and in poorly designed engines. It is indicated by a in Fig. 35. The volumetric clearance is expressed as a percentage of the piston displacement or volume swept through by the piston. It varies from 2 per cent or less in the best engines to as high as 15 per cent in the cheaper and less economical models. It is made up of the parts designated by c in Fig. 35. 55. Cushion Steam and Cyl- inder Feed. It is customary to imagine the steam operating within an engine cylinder to consist of two parts, the cushion steam and the cylinder feed. The former is that part of the total which is contained in the clearance space before the admission valve opens and serves to cushion the reciprocating parts of the engine. The cylinder feed is the steam which enters through the valve for each cycle. If the same cycle is produced time after time so that all temperature effects are repeated at regular intervals and so that all events occur at the same points in successive cycles, the quantity of steam retained in the clearance volume will be the same for successive cycles. It is impossible to measure the quantity of this steam directly and indirect methods are therefore adopted for that purpose. It is often assumed that the steam is dry and satu- rated when compression begins, as at the point e in Fig. 34. With this assumption, the weight of cushion steam can be determined by dividing the volume occupied, that is, 86 STEAM POWER Ve t by .the volume occupied by one pound of dry saturated steam at the same pressure. Thus, Cushion steam = = ; — - — =- lbs. . (29) Sp.vol. at pressure P e The weight of cylinder feed can be very accurately determined by condensing and weighing the steam leaving the engine in a given time and dividing by the number of cycles performed during the same period. It can also be determined by metering the steam entering the engine or by measuring the water fed to a boiler supplying only the engine in question. An approximate determination of the quantity of the cylinder feed can also be made directly from an indicator diagram by determining what is known as the diagram water rate. This will be considered in detail at a later point. When cushion-steam and cylinder-feed have both been determined, the weight of steam contained in the cylinder between cut-off and release can be found by adding the two quantities. Thus, W = W f +W K , (30) in which W — total weight of steam expanding in cylinder per cycle; Wf= weight of cylinder feed per cycle; and Wk = weight of cushion steam per cycle. The volume which the mixture would occupy if dry and saturated at any given pressure can be determined by multiplying W the total weight by the specific volume for that particular pressure. 56. Determination of Initial Condensation. The loss due to initial condensation is so important that it is cus- tomary to determine the amount of this loss when studying engines. This can be done with fair accuracy by means of the indicator diagram. THE REAL STEAM ENGINE 87 To make such a study it is necessary to know the total freight of material in the engine cylinder at the point of cut-off. This weight may be determined by any of the methods just given. With the weight known, the volumes which this material should occupy at different pressures if dry and saturated can be determined by multiplying by the specific volumes at the various pressures. Plotting these points on a PF-diagram and connecting them will give a saturation curve for the material in the cylinder such as the curve shown in Fig. 36. By drawing this curve on the indicator diagram ob- tained from the engine and then comparing distances such as ab and ac as explained in section 26 of Chapter III Fig. 36. Fig. 37. the quality of the steam within the cylinder at all pressures between cut-off and release can be determined. The weight of initial condensation (up to the point of cut-off) must be the total weight of water shown as existing within the cylinder at that point minus any water brought in by the steam if it was not dry when entering the engine. Should the saturation curve cross the real expansion curve, as shown in Fig. 37, it indicates that the steam oc- cupies volumes greater than the specific volumes toward the end of the expansion; the steam within the cylinder must therefore be superheated during this part of the cycle. Many formulas have been devised for giving the quan- tity of initial condensation. They are all based upon the results of experiment and generally only give reliable 88 STEAM POWER values for cases similar to those used in developing them. One formula of this sort which has been very widely tested and been found to give reliable results within its field of applicability is that devised by Robert C. H. Heck and explained in his books on the steam engine. The formula is c_ sd \fN\pe in which =«fe ra = the fraction representing initial condensation; for ordinary cases it is the fraction of the cylinder feed which is condensed during admission, but wheu compression is very high and when great weights of steam are retaiued in the clearance it is the frac- tion of all the material within the cylinder which exists in liquid form at the time of cut-off; c = a coefficient, which varies between 0.25 and 0.30 with ordinary engine types. Its value is unknown for certain new types such as the Una-flow. N = engine speed in revolutions per minute (R.P.M.); s = a constant for any engine, equal to nominal surface in square feet divided by nominal volume in cubic feet. The nominal surface is the area of the inner walls and the ends of a cylinder with diameter equal to the internal diameter of the cylinder and with a length equal to the stroke of the engine. The nominal volume is the cubic contents of such a cylinder; s = y-(2-^-f-4j in which D and S represent diameter and stroke of engine in inches; = a temperature function obtained from Table II as there indicated; p = the absolute pressure in cylinder in pounds per square inch just after completion of cut-off; THE REAL STEAM ENGINE 89 e = cut-off ratio, that is, ratio of cylinder volume opened up by time cut-off has just been completed, to the total piston displacement. TABLE II For Finding Values of .0 for Use in Heck Formula ki~ kz when ki and ki are chosen from table for highest and lowest pressures existing in cylinder V k V k V k p k V k V k 1 175 15 210 50 269.5 90 321.5 160 389 230 441 2 179 20 220 55 277 100 332.5 170 397 240 447.5 3 183 25 229 60 284 110 343 180 405 250 454 4 186 30 238 65 291 120 353 190 413 260 460.5 6 191 35 246 70 297.5 130 362.5 200 420 270 467 8 196 40 254 75 304 140 371.5 210 427 280 473 10 200 45 262 80 310 150 380.5 220 434 290 479 57. Methods of Decreasing Cylinder Condensation. Before discussing methods of decreasing the loss due to cylin- der condensation it will be well to consider what things may be expected to determine the extent of such loss. The condensation is due directly to the transfer of heat from one body to another at lower temperature, and anything which tends to increase the total amount of heat thus trans- ferred will increase the total condensation. It is therefore evident that the ratio of steam condensed to steam supplied will be greatest when: (a) The time of contact is greatest; (b) The ratio of surface exposed to volume enclosed is greatest, and (c) The temperature difference is greatest. The time of contact can be controlled to a certain extent by controlling the speed of the engine and, with other things equal, the higher the speed the lower should be the condensation. The ratio of surface exposed to steam to the volume occupied by steam has a great influence on the amount of 90 STEAM POWEK condensation which occurs. The surface of the clearance space, including the interior surfaces of all ports or passages leading to the valves, seems to have the greatest influence, and the clearance space which is most nearly a short cylinder without connected passages may be expected to give the least initial condensation. The size of the engine is also important in this connec- tion. The area exposed does not increase as rapidly as does the volume inclosed when the diameter of a cylinder is increased, and therefore large cylinders give smaller ratio of surface to volume and therefore a smaller percentage of steam condensed. Large engines thus have a decided advantage over small engines. The shape of the cylinder also has an effect. The longer the cylinder with respect to its diameter the more favorable its performance. The point at which cut-off occurs is also intimately connected with the condensation loss. In a given cylinder with a given clearance the total condensation within the clearance space may be assumed practically constant if speed and temperature remain about the same. But if the cut-off is made later larger quantities of steam are admitted per stroke, and hence the fraction of the total cylinder feed which is condensed decreases. The temperature differences depend on upper and lower pressures, that is, on the pressure range. The inner surfaces of the walls follow as rapidly as possible the tem- perature changes of the steam within them. Thus their average temperature is somewhere between the upper and lower temperatures of the steam. If now, with a given upper steam pressure and therefore temperature, the lower pressure be reduced, the average wall temperature also will be reduced, and therefore the differences between the temperature of the entering steam and the average tem- perature of the walls will be increased with a resulting in- crease in condensation loss. THE REAL STEAM ENGINE 91 The methods of decreasing this loss can now be con- sidered. They are given below under separate heads with brief explanation when necessary. (a) Clearance spaces should be properly designed so that the minimum surface is exposed. (6) The proportions of cylinder (diameter and stroke) and the speed of the engine should be so chosen that the condensation loss is reduced to a minimum. (c) The engine should be so proportioned that when delivering its rated power the cut-off occurs at such a point as to make the percentage of cylinder condensation the minimum consistent with other requirements. (d) The cylinder should be surrounded by spaces filled with air or by materials which are poor conductors of heat so as to decrease loss by radiation, because all heat lost in this way must be supplied by the condensation of steam within the cylinder. Such metallic parts as cannot be " lagged" in this way should be polished because polished surfaces radiate less heat than dull surfaces under like conditions. (e) The cylinder may be surrounded by a steam jacket, that is, a space filled with steam similar to that supplied the cylinder. The use of such a jacket sometimes results in a considerable saving and at other times in a great loss. The cylinder proportions, speed and pressure range seem to be the determining factors, and most long-stroke cylin- ders operating at low rotative speed and with small pressure ranges are jacketed. (/) The engine may be compounded, that is, the expan- sion of the steam may be made to occur in two or more cylinders taking steam in series. This results in decreas- ing the pressure range in each of the cylinders and effects a decided saving under proper conditions. Compounding will be considered in detail in a later chapter. (g) The engine may be supplied with superheated steam. If the steam is sufficiently superheated it can give up part or all of its superheat to heat the cylinder walls, and thus no 92 STEAM POWER condensation need occur. Heat interchanges between metal and superheated steam also appear to be less rapid than is the case when the steam contains water, so that a saving results from this source also. Tests made with saturated and with superheated steam indicate that from 7° to 10° of superheat are generally required to prevent 1 per cent of initial condensation. Results differ greatly with the character of the engine, with its economy on saturated steam, with its valve gear, etc. Superheats of from 25° to 50° can generally be used with any well-designed engine, but higher temperatures usually require specially constructed engines. With proper cylinder and valve construction the maximum permissible temperature is set by the lubricating oil. At the present time the max- imum steam temperature is thus limited to 600° F. or less. (h) The engine may be built to operate on the Una-flow cycle described later. In this case a very ingenious modifica- tion of the engine results in decreasing the temperature dif- ferences between walls and steam, with resulting diminution of heat transfer. 58. Classification of Steam Engines. Steam engines, are classified on many different systems, the one used in any particular case being determined largely by circum- stances. The principal methods of classification are indi- cated in the following schedule : Classification of Steam Engines On the basis of rotative speed. (a) Low speed; (6) Medium speed; (c) High speed. On the basis of ratio of stroke to diameter. (a) Long stroke; - (6) Short stroke. On the basis of valve gear. A. Slide valve; (a) D-slide valve; (b) Balanced slide valve; (c) Multiported slide valve; (d) Piston valve. THE REAL STEAM ENGINE 93 B. Corliss valve; (a) Drop cut-off; (b) Positively operated. C. Poppet valve. On the basis of position of longitudinal axis. (a) Vertical; (6) Inclined; (c) Horizontal. On the basis of number of cylinders in which steam expands. A. Single expansion or simple engine. B. Multi-expansion engine, (a) Compound expansion; (b) Triple expansion ; (c) Quadruple expansion. On the basis of cylinder arrangement. (a) Single cylinder; (6) Tandem compound; (c) Cross compound; (d) Duplex. On the basis of use. (a) Stationary engines; (b) Portable engines; (c) Locomotive engines; (d) Marine engines; (e) Hoisting engines. 59. Rotative Speeds and Piston Speeds. High-speed engines operate at a comparatively high rotative speed and are characterized by a short stroke in comparison with the diameter of the cylinder, the stroke generally being equal to, or less than, the diameter. The piston speed, by which is meant the feet travelled by the piston per minute, generally falls between 500 and 700. It is not considered advisible to allow piston speeds of stationary steam engines to exceed about 750 feet per minute for ordinary constructions and the great majority of engines give much lower values. The piston speed will obviously be given by the formula S = 2LN, (32) in which S = piston speed in feet per minute; L = stroke in feet; and N = revolutions per minute, 94 STEAM POWEK and it is evident from this formula that as the rotative speed is increased the piston speed will increase unless the length of stroke is proportionately decreased. As a result, high-speed engines have short strokes in com- :$o 20 ■Sl8 o \ * \ \ \ H IGH SP EE D EN Gl NES \ / \ £ b. \' f > <6 ,* y y #; 4 £ y sto h £ ipe 3d *$ £> y& ^ vy <^ a/ 7 / c f/ *■ *'/ 375 350 325 300 275 250 225 * 200 175 150 125 100 8 10 12 14 16 18 20 22 24 26 Diameter of Cylinder (Inches) Fig. 38. — Proportions of High Speed Engines. parison with their cylinder diameters and slow-speed engines have long strokes. The characteristic relations between cylinder diameter and stroke, rotative speed and piston speeds of high-speed engines are given in Fig. 38. High-speed engines are generally fitted with some THE REAL STEAM ENGINE 95 "S 8 R. P. M. Piston Speed (S) Ft. per Min. S o »o o >ra o so CO W3 »o -* -h \ \ / \ / \ / \ \ / \ / V \ / \\ / / V / \\ / \ \ / \ \ / o © \ / \ s / UJ l\ V (1 p*\ / v -z. II 1 Stt Q UJ UJ d\ CO <$ <: V £\ _j « \ \ \ \ \ \ \ \ \ \ y v \ / y a o CM o Fig. 68. VII; but the means by which such diagrams are obtained from operating engines was not given. Indicator diagrams showing the pressure and volume changes experienced by steam in the cylinders of real 115 116 STEAM POWER engines are obtained by means of an instrument known as an indicator. The operation of obtaining such diagrams is known as indicating the" engine. An external view of one form of indicator is shown in Fig. 68 and a section through the instrument is given in Fig. 69. The method of connecting an indicator to the Pi'ston Rod- "Connected to Engine Cylinder Fig, 69. cylinder of a steam engine and one method used for driving it are illustrated in Fig. 70. The indicator is intended to draw a diagram showing corresponding pressures and volumes within the engine cylinder and must, therefore, contain one part which will move in proportion to pressure variations and another which will move in proportion to volume changes. The one may INDICATOR DIAGRAM AND DERIVED VALUES 117 be called the pressure-measuring and the other the volume- measuring device. The pressure-measuring device generally consists of a piston, such as shown in the figure, working with minimum friction in a small cylinder and fitted with a spring which will resist what may be called outward motion (upward in the figure). The cjdinder containing this piston is coupled to a short pipe connected with the clearance space of the engine and, whenever the indicator cock in this connection is open, the steam acting on the engine piston Fig. 70. — Method of Attaching and Operating an Indicator. will also act on the indicator piston. Steam of any given pressure will drive the indicator piston out against the action of the spring until the pressure exerted by the spring is equal to that exerted on the face of the piston. The indicator piston will thus move out different distances for different pressures, and, through the piston rod and pencil mechanism, will move the pencil point to various heights corresponding to different steam pressures. The pencil mechanism is so arranged that the point traces a straight vertical line on the drum as the indicator piston moves in and out. Springs are made to certain definite scales, thus there 118 STEAM POWER are, for instance, 10-lb., 25-lb., 50-lb. and 100-lb. springs. The number which is known as the scale of the spring designates the steam pressure in pounds per square inch which is required to move the pencil point 1 inch against the action of such a spring. With a 100-lb. spring in the indicator, a steam pressure of 50 pounds per square inch acting on the indicator piston would drive the pencil up a distance of half an inch, a pressure of 100 pounds per square inch would give 1 inch of motion and so forth. The volume-measuring device is of an inferential kind. It simply indicates the position attained by the engine piston at the time when a given steam pressure existed in the cylinder and the volume occupied by the steam can be calculated from piston position and cylinder dimensions. The position of the piston is indicated by connecting the cord wound around the drum to some part of the engine which is rigidly connected to the piston. The crosshead is commonly used for this purpose and, since the motion of this member is generally much greater than the circum- ference of the drum, it is necessary to use a reducing mechanism of some sort. This mechanism must be very accurate, so that it moves the drum as nearly as possible in proportion to the motion of the engine piston. The pencil point moves up and down as the pressure within the cylinder varies, and the drum rotates under the point in proportion to the motion of the engine pis- ton, so that the combination of the two motions brings the pencil point to successive positions on the drum which indi- cate successive corresponding values of steam pressure and piston position. By mounting a piece of paper, known as a card, on the drum and pressing the pencil point upon this paper, the successive positions occupied by the pencil point will be recorded in the form of a series of curves and straight lines. If the drum is rotated with the lower side of the indica- tor piston connected to atmosphere, the pencil will trace INDICATOR DIAGEAM AND DERIVED VALUES 119 a horizontal line. This is known as the atmospheric line and is used as a reference for locating the pressure scale. If the indicator cylinder is then connected with the engine cylinder and the drum is rotated by the reducing mechanism, a diagram similar to that of Fig. 71 will be drawn upon the card. The atmospheric line indicates the height assumed by the pencil when atmospheric pressure acts on the piston and, knowing the value of the existing atmospheric pressure (barometer reading) and the scale of the spring, a line at a height representing zero pressure can be drawn on the card. This line is indicated in Fig. 72. The length of the card between the lines a and b is proportional to the length of the engine stroke and there- Atmospheric Line ■ Fig. 71. i N J a V\ V ^ 7 Atmospheric Line Line of Zero Pressure ^, ' ' i i i i i i ., . i i Fig. 72. fore to the piston displacement, that is, to the volume swept through by the piston. Knowing the clearance volume of the engine as a percentage or fraction of the piston displacement, this fraction of the length of the diagram can be laid off from the end of the diagram to give a line of zero volume. This line is also indicated in Fig. 72. With the line of zero pressure and the line of zero volume drawn in, all values of steam pressure and volume occupied by steam can be read directly from the diagram, and it thus forms a picture of what occurs within the real engine cylinder. The indicator diagram is used for a number of purposes, the more important being: 120 STEAM POWEE (1) The determination of the energy made available within the cylinder, that is, the indicated horse-power, I.h.p. (2) The determination of the amount of initial conden- sation and of heat interchanges between walls and cylinder. (3) The determination of what is known as the diagram water rate. (4) The study of the operation and timing of valves. The second one of these uses has already been considered in Chapter VII, the others are treated in succeeding sections. 64. Determination of I.h.p. The lines of the indicator diagram show by their height the pressures or forces acting on the engine piston as it moves. But the product of force by distance is equal to work and these lines can be used there- fore for determining the net work done by the steam upon the piston. In Fig. 73 is shown the upper part of the diagram, the curved lines represent- ing the successive pressures in pounds per square inch which acted on the left face of the piston while it moved outward. If the average pressure could be deter- mined and multiplied by the area of the piston face, this product would be the average total force acting on the piston. Multiplying this by the distance traveled would give the work done by the steam upon the piston. Expressed in the form of an equation, E = poXaXLftAbs., (33) in which Eo = work done upon piston by steam during outstroke ; po = mean pressure (in pounds square inch) acting on piston during outstroke ; 1 -Length of Diagram s»-» >^ in Inchas 1 illlilv 1 a 1 1 1} 1 1 5 1 ! 1 1 1 J ■:-f^ Fig. 73. — Positive Work Area. INDICATOR DIAGRAM AND DERIVED VALUES 121 a = area of piston face in square inches; and L = stroke of piston in feet. For the instroke shown in Fig. 74, the work done by the piston on the steam is given by the similar expression m Ei=piXaXL ft.-lbs., (34) Fig. 74. — Negative Work Area. in which Ei and p% represent work done and mean pressure respectively. The net work done by the steam upon the piston per cycle is then, # C ycie =E -Ei= (p - Pi) oL f t.-lbs. . . . (35) The values of po and pi can be found directly from the diagram by dividing the areas A and A t respectively by the length I and then multiplying by the scale of the spring, giving A Po and I A* X scale of spring, p* = -7-Xscale of spring, 122 STEAM POWER so that, Po-Vi = V = — ] — *X scale of spring . ... (36) = areaof f agram X S cale of spring. . (37) The value of p evidently can be determined very simply from the indicator diagram, and the work per cycle can be found when p is known by substituting in the following equation, obtained by putting p for po — pi in Eq. (35), # C ycie = ;pXaXLft.-lbs (38) The pressure p is known as the mean effective pressure and is often represented by M.E.P. If n cycles are produced per minute, the net work done by the steam upon the piston per minute will be E m m = pXaXLXn, (39) which is generally rearranged to read, E m [n=pLan, (40) in which form the group of letters forming the right-hand member is easily remembered. Since 33,000 foot-pounds per minute are equivalent to one horse-power, it follows that the power made avail- able as shown by the indicator diagram, that is, the indicated horse-power, must be, T , pLan N Lh - p - = 337W (41) in which p = mean effective pressure in pounds per square inch ; L = stroke of piston in feet ; a = area of piston in square inches; = (diam. cyl. in inches) 2 X7r/4 = .7854d 2 ; and n = number of cycles per minute. INDICATOR DIAGRAM AND DERIVED VALUES 123 If an engine cylinder takes steam on one side of the piston only, that is, if the cylinder is single acting, the num- ber of cycles produced per minute is equal to the number of revolutions per minute, but it should be noted that for other arrangements this is not necessarily true. In the case of double-acting engines which receive steam at both ends of the cylinder, the number of cycles produced is equal to twice the number of revolutions. It should also be noted that the symbol a represents the area of the piston face upon which the steam acts. If a piston rod extend from the face of the piston to and through the cylinder head (as is always the case at the crank end of double-acting cylinders), the area of the piston rod must be subtracted from that of the piston to obtain the area on which the steam really acts. When a tail rod is used, a correction must be made for each side of the piston. In the case of double-acting engines the indicated horse-power may be determined in two ways: It may be figured separately for the two ends of the cylinder, or the values for the area and pressure may be averaged for the two ends and the value of n chosen equal to twice the revolutions per minute. The former is generally the more accurate method. It will have been observed that the area of the indi- cator diagram must be determined before the mean effective pressure can be found. This area is generally measured by means of an instrument known as a planimeter, and this is the most accurate method. It occasionally happens, however, that a planimeter is not available when the value of the indicated horse-power is desired. Under such circumstances an approximate determination of the area of the indicator diagram can be made by the method of ordinates. For this purpose the length of the diagram is divided into an equal number of parts, usually ten, as shown in 124 STEA.M POWER Fig. 75 and vertical lines y\, 2/2, 1/3, etc., are drawn at the center of each of the parts into which the diagram has been divided. The mean ordinate or height is then found from the equa- tion, = 2/i+?/2+?/3+?/4+etc. number of vertical lines' (42) and the mean effective pressure is then determined by multiplying y m by the scale of the spring. An indicator diagram similar to that shown in Fig. 76 is occasionally obtained. The small loop on the end repre- sents negative work, since the pressure of the steam which w m m Fig. 75. Fig. 76. does work upon the piston is lower than that which resists the return of the piston. When using a planimeter, this area is automatically subtracted from that of the rest of the diagram, but care should be taken to see that this is also done when the method of ordinates is used. ILLUSTRATIVE PROBLEM 1. Determine the I.h.p. of a double-acting steam engine, having a cylinder 8 ins. diameter, length of stroke, 12 ins., running at 100 R.P.M., the mean effective pressure (M.E.P.) on the piston being 45 lbs. Neglect the area of the piston rod. I.h.p. pLan _(pXa) lbs.X(Ln) ft. per min. 33,000 = 33,000 ft.-lbs. per min. (45X8X8X.7854) lbs. XffX 100X2) ft. per min. 33,000 INDICATOR DIAGRAM AND DERIVED VALUES 125 2260 lbs. X 200 ft. per min. 33,000 = 14 nearly. 2. The I.h.p. of a double-acting engine is 14, the R.P.M. =100; M.E.P. =45 lbs.; length of stroke = 12 ins. Find the diameter of the cylinder, neglecting area of piston rod. First determine the area of the piston from the formula vLan 33,000 I.h.p. Lh - P '=33^0 ° r "' pXLXn ' 33,000X14 „ A . rf ° = 45XlX100X2 =51 - 4sq - m - = T ; ,7 I 51 - 4 — - '—- = Vq5A=S ins. (approx.), 7854 65. Conventional Diagram and Card Factors. It is often necessary to approximate the mean effective pressure obtained in the cylinder of an engine when no indicator diagrams are available. The most common case is when an engine is being designed to carry a certain load and it is desired to determine the necessary cylinder dimensions and speed. If the probable mean effective pressure can be determined, the dimensions and speed can be found from the equation, pLan I.h.p. per cylinder end = I.p.h. 33,000' by rewriting it pLnX0.7854:d 2 33,000 from which d? 33 000I.hp. 0.7854 pLn v y Since n is equal to revolutions per minute for one cylinder end, the product of L by n must be equal to half the piston 126 STEAM POWER speed of the engine, and a proper value of this product can be chosen for substitution in the equation. If a proper value for p can then be predicted the only unknown remain- ing will be the diameter d, and this can be found by solving the equation. The prediction of the mean effective pressure is made either by drawing upon recorded experience in the form of values obtained in similar engines previously constructed or by means of what is known as a conventional indicator diagram. The conventional diagram is drawn with upper and lower pressures equal to those expected in the case of the real engine, and all expansions and compressions are drawn as rectangular hyperbolas. The equation of the rectangu- lar hyperbola is PiV^PnVn, (44) in which subscript 1 indicates initial conditions and sub- script n represents any later conditions with the same mate- rial in the cylinder. This law is assumed because it is the simplest and, as a rough average, gives values as close to those actually attained as do any of the more complicated laws. The diagram may be drawn as nearly as possible like the one which the engine may be expected to give or it may be drawn with various simplifications which remove it more and more from the approximation to an actual indicator diagram. In any case, the mean effective pres- sure is determined from this diagram and this value is then multiplied by a corrective factor, the value of which has been determined by experience. This corrective factor is called the diagram factor or card factor and it is realty the ratio of the area of the diagram the engine would really give to the area of the conventional diagram used. The simplest form of conventional diagram is drawn INDICATOE DIAGRAM AND DERIVED VALUES 127 by neglecting the clearance volume and has the shape shown in Fig. 77. The upper line is drawn horizontal at a height representing the highest pressure expected and of such a length (compared with the length of the diagram) as will approximately represent the fraction of the stroke at which cut-off is to occur in the real engine. The expansion curve is then drawn in as a rectangular hyperbola and extended until the end of the diagram is reached. The next line is drawn vertical and the lower line of the diagram is drawn horizontal at a height representing the pressure expected in the space into which the engine is to exhaust. This simple diagram can be divided into the three areas shown and the value of the work represented by these areas can be determined from the equations given below, the first and last of which should be self-evident from what has preceded. The equation for the work represented by area A 2 can be determined very easily by means of integral calculus. The equations are, Fig. 77.— Conventional In- dicator Diagram. and Ai represents P\V\ ft. -lbs.; A2 represents P\V\ log e ^- = Pi~Fi log e r ft.-lbs., As represents P2V2 ft.-lbs., in which P represents pressure in pounds per square foot and V represents volume in cubic feet. The total area is then equal to the sum A1+A2 — As and the net work is equal to a similar sum of the right- 128 STEAM POWER hand members given above. The net work must also equal the mean-effective pressure P m multiplied by the total volume change, so that and PmV 2 = PiVi+PiV 1 \oger-P 2 V2, . . (45) P m =Pi^+Piploger-P 2 . . . (46) ^(ft+ft 10 ^) - ^' • • • (47) 1 Vi . . and substituting - for ■=- this gives r v 2 l±^_p 2 (48) Pm = P v 2 The ratio T^=r is called the ratio of expansion and its Vi 1 reciprocal, ■==- = — is known as the cut-off ratio. By the use of this ratio the volume terms can be disposed of and the equation above is obtained. This equation then gives the mean effective pressure in terms of upper and lower pres- sures and the fraction of the stroke at which cut-off is desired in the real engine and no cylinder dimensions need be known. Since pressures in steam-engine practice are usually given in pounds per square inch, the equation for mean effective pressure is more useful in the form Pm = J l +^l)- P2 , .... (49) in which pi and p 2 and p m are expressed in pounds per square inch absolute. For convenience in the use of this INDICATOR DIAGRAM AND DERIVED VALUES 129 equation the values assumed by the bracketed quantity are given for various conditions in Table III. TABLE III r 1 +log e r r r 1 +log e r r r 1 +log e r r 1.0 1.00 6.0 0.465 16.0 0.236 1.5 0.937 7.0 0.421 17.0 0.226 2.0 0.847 8.0 0.385 18.0 0.216 2.5 0.766 9.0 0.355 19.0 0.208 3.0 0.700 10.0 0.330 20.0 0.200 3.5 0.644 11.0 0.309 21.0 0.192 4.0 0.597 12.0 0.290 22.0 0.186 4.5 0.556 13.0 0.274 23.0 0.180 5.0 0.522 14.0 0.260 24.0 0.174 5.5 0.492 15.0 0.247 25.0 0.169 The values of the mean effective pressures obtained from this form of diagram are very much higher than are to be expected from real engines with the same initial and terminal pressures and the same nominal ratio of expan- sion. They are therefore corrected by multiplying by the proper diagram factor as selected from Table IV. It is obvious from the range of values given that the selection of a proper value for the factor depends largely on expe- rience, but such experience is quickly gained by contact with real engines and a study of the practical diagrams. TABLE IV Diagram Factors Simple slide-valve engine 55 to 90% Simple Corliss engine 85 to 90 Compound slide-valve engine 55 to 80 Compound Corliss engine 75 to 85 Triple-expansion engine 55 to 70 66. Ratio of Expansion. — The ratio of expansion used above is sometimes called the apparent ratio. It is not the 130 STEAM POWER real ratio of expansion for an engine with clearance, such an engine the real ratio of expansion is For W + Va (50) Fig. 78. in which the symbols represent the volumes indicated in Fig. 78. The numerical values of r and r' are often very different and care should be used in dis- tinguishing between them. The diagram factors referred to in Table IV are for idealized con- ventional cards without clearance as shown in Fig. 77. ILLUSTRATIVE PROBLEMS 1. Given an engine with a stroke of 24 ins. and cut-off occurring at \ stroke. Steam pressure of 160 lbs. per square inch and back pressure of 16 lbs. Assume diagram factor = 80%. Neglect- ing clearance, find the probable M.E.P. M.E.P. =p l+log e r p 2 = 160 l+log e 3 r / \ 3 <.7-16=112.0-16=96 1bs. 16 Hence probable M.E.P. =.80X96 =76.8 lbs. 2. A given double-acting engine indicates 75 I.h.p. under the following conditions: Cut-off at 20%; steam pressure, 140 lbs. per square inch absolute; piston speed, 600 ft. per minute; back pressure, 2 lbs. per square inch absolute. Assume a diagram factor for this type of engine equal to 85% ; and neglecting clearance, find a convenient size of the cylinder (diameter and stroke). INDICATOR DIAGRAM AND DERIVED VALUES 131 Solution. r= Jo =5; ^ 1+bger j / l+Iog e 5 \ 73.1 -2 =71.1 lbs. per sq.in. Diagram factor =85%. Hence probable M.E.P.=71.1X.85 =60.4 lbs. Therefore, since T , pLan 75X33,000 co n . Lh - P - = 3P00 ft= 6O«600- = 68 - 3Sq - m - d=9% ins. (approx.); and since 2Ln = 600, assume L = 1 ft. hence n=300R.P.M. The engine is rated 9.5X12 ins., running at 300 R.P.M. 67. Determination of Clearance Volume from Diagram. It was shown in a preceding paragraph that the clearance volume of a cylinder must be known in order to draw the line of zero volumes on the indicator diagram. This volume can be determined accurately for any real engine by weighing the quantity of water required to fill the clear- ance space, but this procedure is often impossible and an alternative, though approximate, method is often resorted to. This method is graphical and depends upon the assump- tion of the law of expansion and compression. As in the case of the conventional diagrams, expansion and compres- sion are assumed to follow rectangular hyperbolas. It is a property of this curve that diagonals such as aa and bb drawn for rectangles with their corners on the 132 STEAM POWER curve all pass through the origin of coordinates as shown in Fig. 79. If two points a and c are selected on the expansion curve of a real diagram and a rectangle is drawn upon them as shown in Fig. 80, the diagonal bd extended will pass through the origin of coordinates, if the expansion follows the assumed law. The point at which this diagonal cuts the zero pressure line must therefore be the point through which the vertical line of zero volume is to be drawn. If the original assumption were correct, this construc- tion would give the same point when different locations of the points a and c were chosen and when used on the Fig. 79.— Rectangular Hyperbola. Fig. 80. compression as well as on the expansion line. In reality- it will generally give as many different locations for the origin as are chosen for the rectangle abed. It is customary to construct this rectangle of fair size and to locate it near the center of the expansion curve. 68. Diagram Water Rate. When an engine is run on saturated steam, part of the steam supplied is generally condensed upon the cold metal walls surrounding it. The indicator diagram therefore shows the volumes assumed by the mixture of steam and liquid water in the cylinder, but, since the volume occupied by the liquid is negligible, it may be assumed to show the volumes occupied by the part of the mixture which exists in vaporous form. INDICATOR DIAGRAM AND DERIVED VALUES 133 Assuming that the vapor is saturated, the volume occupied by one pound at various pressures can be found from the steam tables and, therefore, the weight existing in the cylinder can be calculated. The weight of steam determined in this way is known as the indicated steam, the diagram steam or the diagram water rate. The diagram water rate is generally determined for a point such as z in Fig. 81 just after cut-off, though some engineers prefer to use a point nearer the lower end of the expansion curve. The volume occupied by the steam con- tained in the cylinder at point z is equal to V z and its weight can be determined by dividing this volume by the specific volume V 2 for the existing pressure P z . Thus, calling the weight of steam in the cyl- inder w z , V z Fig. 81. w >=vl- (51) This quantity of steam is a mixture of cylinder feed and clearance or cushion steam and the weight of the latter must therefore be subtracted from w z to obtain the weight of cylinder feed w f . Assuming the cushion steam dry and saturated at the point k, the weight of cushion steam is Wt V*' (52) so that the weight of cylinder feed per cycle as shown by the diagram at the point z is Wf=W z -W t = y- y~. ...... (53) The formula is generally modified to give the steam consumption per indicated horse-power hour, instead of 134 STEAM POWER per cycle, and it is also expressed in different terms as a matter of convenience. For this purpose let ya = clearance volume divided by piston displacement per stroke lei ~v 2/ z = piston displacement to point z divided by piston dis- placement per stroke Jl V yt = piston displacement to point k divided by piston dis- placement per stroke J± V a = area of piston in square inches; p=mean effective pressure in pounds per square inch; L = stroke in feet ; and n = number of cycles per minute. The piston displacement is then tj^L cubic feet and the volumes at z and k are given by and '*Xffi) + (v«xj§ v>xm)+(y*xm): Substituting these values in Eq. (53) gives the indicated cylinder feed per cycle as w f aL / yz+yci y k -hyci \ , . = m\~T 2 vr) (54) Multiplying by the number of cycles per hour (60 Xn) and dividing by the indicated horse-power, }* gives 00, uuu INDICATOR DIAGRAM AND DERIVED VALUES 135 the diagram water rate, or steam shown by the diagram per I.h.p. hour as Wa 13,750 V •■+ya yt+y a v 2 v t ). (55) in which form the equation involves only values which can be determined directly from the diagram without any knowledge of the engine dimensions. The value obtained for w d will vary as the location of points z and k are varied because of the quality changes occurring during expansion and compression, and it is obvious that the diagram water rate is in no sense an accu- rate measure of the real water rate of the engine. It is, however, often useful for comparison with the real water rate, the ratio giving an indication of the loss by conden- sation and leakage. Average values for real water rates are given in Chapter XI. ILLUSTRATIVE PROBLEM Given the diagram shown in Fig. 82 and the following data from an actual test, find the diagram water rate for point c, and for point n. Double-acting steam engine having: Average piston area =28.9 sq.- in.; Length of stroke = 8 in.; R.P.M.=237; I.h.p. =8.75; M.E.P. =31.6 lbs.; Clearance = 13% ; Beginning of compression = 29% ; Weight of condensate per hour =371 lbs.; Quality at throttle =95%; Sp. vol. at c=7.8; Sp. vol. at n = 12.57; Sp. vol. at K =38.4 cu.ft. per lb.; Assume xt = 100%. & ,_,_ -4 — i-V i —r\-fc--i— 1 ■ Ti !.:•»! ;f v Fig. 82. ia6 STEAM POWER Solution. Substitution in Eq. (55) gives 13,750 ftjc+yci yt+ya m)c= ^~\~v c vT 13,750/ 0.39+0.13 _0. 29+0.13 \ " 31.6 \ 7.8 38.4 / = 24.2 lbs. per I.h.p. per hour at point c. 13,750 fyn+ya yt+ya ( ^ = ~3ix r^r~~^r 13,750 /0.638+0.13 _ 0.29+0.13 \ " 31.6 \ 12-57 ~ 38.4 / = 21.83. lbs. per I.h.p. per hour at point n. 371 Real water rate =——X 0.95 =40.2 lbs. 8.75 69. T^-diagram for a Real Engine. In Chapter VI the T ^-diagrams of the various ideal cycles were given and attention was called to the fact that these diagrams were particularly useful, because they showed certain things which were not apparent from the more common PV- diagrams. It has been customary for many years to draw Te used. The result of using five cylinders is shown in Fig. 88, and it is evident that the clearance surfaces exposed to COMPOUNDING 147 high temperatures, the temperature ranges per cylinder and the ratios of expansion per cylinder are all small. The gain in econ omy should therefore be correspondingly great. There are two limits to the possible multiplication of cylin- ders in this way. (1) As the number increases the radiating surface and therefore the heat lost by radiation increases. The extent of this effect can be appreciated by noting that every cylinder with the exception of the low-pressure cylinder is really an unnecessary addition, because the cycle could be produced entirely in the low-pressure cylinder. On the other hand, the surfaces of cylinders which operate at high temperature are small as compared with that which would be exposed to this temperature if the entire cycle were pro- duced in the low-pressure cylinder. (2) As the number of cylinders is increased, the first cost, the complexity and the cost of lubrication and attend- ance are all increased so that, for each installation, some number will be found beyond which the interest on the investment and the added cost of operation and mainte- nance would more than balance the saving of fuel. The second limit mentioned is the more important commercially, as it is the first one reached. For ordinary operating conditions in stationary power plants expansion in two cylinders generally gives the most economical results. The total ratio of expansion is generally between 7 and 16, that is, the volume of steam at release in the L.P. cylinder is from 7 to 16 times the volume at cut-off in the H.P. cylinder. For large pumping stations and large marine installations, expansion in three cylinders is generally considered the most economical, and total ratios of expansion of 20 or more are used. Four and five cylinders have been used, but the resultant gains do not seem to warrant any extensive installation of such units. Engines using more than one cylinder for the expansion 148 STEAM POWER of steam in the way just described are called multi-expan- sion engines, or compound engines, and the use of multi- expansion- is spoken of as compounding. Custom has almost confined the use of the term compound engine to those in which only two cylinders are used in series as indicated in Fig. 89, and such engines are often spoken of as 2x engines. Engines in which three cylinders are used in series are called triple-expansion or 3x engines. With four and five cylinders in series the engines are known as quadruple or 4x and quintuple or 5x, respectively. In the case of triple-expansion engines of large size, H.P. Exhaust and L.P. Admission To Condenser Fig. 89. Fig. 90. the volume of the low-pressure cylinder required generally becomes so great that it is found economical to use two low-pressure cylinders instead of one. The flow of steam in such an engine is represented diagrammatically in Fig. 90. This type is known as a four-cylinder, triple-expansion engine. All multi-expansion engines are generally operated condensing, and the choice of type is determined partly by the character of work to be done and partly by economical considerations. In all cases the boiler pressure must be chosen to suit the type of engine used. The pressures ordinarily used with the different types are given in Table V. COMPOUNDING 149, TABLE V Boiler Pressure Commonly Used Type of Engine. Boiler Pressure. Pounds per Sq.in. Gauge. Simple 80 to 125 High-speed compound 100 to 170 Low-speed compound 125 to 200 Triple expansion and higher 125 to 225 73. The Compound Engine. The term compound engine will be used hereafter in the commercial way as reierring to a 2x engine. Such engines may roughly be divided roughly into two types, receiver and non-receiver engines. The latter are often called Woolf engines, after the man who first used this construction. A receiver engine has a vessel known as a receiver located between the two cylinders and so connected with them that the high-pressure cylinder exhausts into the receiver and the low-pressure cylinder draws its steam from the receiver. By using a receiver the cylinders are made independent of each other so far as steam events are concerned; the high-pressure cylinder can exhaust at any time without reference to the events occurring in the low-pressure cylinder. A Woolf type has practically no receiver, the high-pres- sure cylinder exhausting directly into the low-pressure cylinder through the shortest convenient connecting pass- age. As the high-pressure cylinder must exhaust directly into the low-pressure cylinder it follows that cut-off must not occur in the latter until compression starts in the former; i.e., very near the end of the stroke. An engine with a receiver of infinite size would give a horizontal exhaust line for the high-pressure cylinder and a horizontal admission line for the low-pressure cylinder, since the small amount of steam given to or taken from the 150 STEAM POWER receiver would have no appreciable effect upon the pressure within that vessel. Neglecting throttling losses, the high- pressure and low-pressure cards would therefore fit together as originally indicated in Fig. 86. With receivers of finite size there are pressure changes during exhaust by the high- and admission to the low-pres- sure cylinders, and real valves and connections also cause certain throttling losses, so that the lines representing these events are not horizontal nor do they exactly coincide. A diagrammatic arrangement of the Woolf engine is given in Fig. 91 with idealized diagrams obtained by Fig. 91. assuming hyperbolic expansions, no clearances, and no throttling losses. The pistons must make their strokes together in such engines, but they may move in the same direction, as shown in the figure, or in opposite directions. The ideal diagram would be that shown at (a) by the lines AbcdCDA. The idealized high-pressure diagram is abcda and the idealized low-pressure diagram is ABCDA. The exhaust line da of the high-pressure diagram and the admission line BC of the low-pressure diagram are pro- duced at the same time. Corresponding points on these two lines represent the common pressures assumed by the steam not yet exhausted from the high-pressure cylinder, the steam in the small connecting passage and the steam COMPOUNDING 151 already admitted to the low-pressure cylinder. As the movement of the low-pressure piston opens up volume faster than the high-pressure piston closes up volume, the volume occupied by the steam continues to increase as the low-pressure piston moves out, and its pressure there- fore decreases. The two diagrams are shown back to back at (b) in the figure and the horizontal line xX connects corresponding JTrom Boiler Fig. 92. points on the exhaust of the high pressure and the admission of the low pressure cylinders respectively. Compound engines are also divided into two types on the basis of cylinder arrangement. When the axes of both cylinders coincide as shown in Fig. 92 they are called tandem compounds. When the axes are parallel as shown in Fig. 89, the engines are spoken of as cross-compound engines. 74. Cylinder Ratios. The idealized diagrams of a com pound engine with infinite receiver volume are shown in Fig. 93 by abed and ABODE. The height of the high-pressure exhaust line is the same as that of the low-pressure admission line and represents the receiver pressure j)r. The value of the receiver pressure is determined by the point chosen for cut-off in the low-pressure cylinder. Thus if cut-off in the low-pressure cylinder is made to occur earlier, as at some point C", the admission line for this Fig. 93. 152 STEAM POWER cylinder must move up to B'C and the receiver pressure must rise correspondingly. The exhaust pressure in the high-pressure cylinder would also rise an equal amount. Changing the point of cut-off in the low-pressure cylinder also produces another result. As the receiver pressure rises the work area of the high-pressure diagram is obviously de- creased, while that of the low-pressure diagram is increased. In a simple engine the area of the diagram becomes smaller the earlier the cut-off, and it should be noted that just the reverse of this occurs in the low-pressure cylinder of a com- pound engine. It is evident that the choice of the receiver pressure or of the point of cut-off in the low-pressure cylinder determines the relative areas of the high-pressure and low-pressure diagrams and it also determines the relative size of the two cylinders. The diagram of Fig. 93 shows that late cut-off in the low-pressure cylinder calls for a larger high-pressure cylinder than does early cut-off. The ratio of the piston displacement of the low-pressure cylinder to that of the high-pressure cylinder is called the cylinder ratio. Designating this ratio by R, and using other symbols as in Fig. 93, B=J|. ....... (61) The cylinder ratios chosen for real compound engines vary greatly in different designs and no given ratio has been proved the best for a given set of conditions. Normal practice gives the average values listed in Table VI, but cylinder ratios as high as 7 have been used with excellent results. TABLE VI Cylinder Ratios for Compound Engines Cylinder ratio Initial pressure (gauge) non-condensing . Initial pressure (gauge) condensing 100 3| 120 100 120 4* 150 COMPOUNDING 153 The cylinder ratio to be used in a given case may be determined by any one of several considerations or by a combination of them, the latter being more often the case. Thus it may be deemed desirable to obtain the same amount of work from both cylinders; or to obtain equal temperature ranges; or to have cut-offs occur at the same fraction of the strokes; or to have the same total load on the two piston rods during admission; or to obtain the maximum possible uniformity of turning effort at the crank. The con- sideration of equal work is generally regarded as the most important. 75. Indicator Diagrams and Mean Pressures. The idealized diagrams for a compound engine with clearance, with incomplete expansion in both cylinders, and without compression - - tp' i\a ■ rvu - i ■ Fig. 94. are given in Fig. 94. Ihe nominal total ratio of expansion would be L L -^-l H , but the total ratio of expansion taking account of clearance is Total ratio of expansion = L ni L , . . (62) and the cylinder ratio is The mean effective pressures can be found from each of the diagrams in the ordinary way and the indicated horse-power of each cylinder determined therefrom. The indicated horse-power of the engine is then equal to the sum of the values obtained for the separate cylinders. It is often convenient to refer the mean effective pres- sure of all cylinders to the low-pressure cylinder as though this were the only cylinder acting. In the simple form of diagram, such as that shown in Fig. 93, it is obvious that this could be obtained by measuring the area AbcDEA, 154 STEAM POWER dividing by the length AE and multiplying by the scale of the spring, just as though the diagram were all produced in one cylinder with the piston displacement equal to V L . In the case of the diagrams given in Fig. 94 a similar method could be adopted, or the mean effective pressure of each cylinder could be determined separately and then the equivalent pressure which would give the same result on the low-pressure piston could be determined analytically. Assume for this purpose that the mean effective pressure of the high-pressure is equal to p H pounds per square inch, that the mean effective pressure of the low-pressure cylinder is equal to p L and that the cylinder ratio is R. The strokes of all cylinders of a multi-expansion engine are generally equal, so that the piston areas are in the same ratio as the cylinder volumes (piston displacements). In the case of a 2x engine, therefore, the area of the low-pressure piston is R times as great as that of the high-pressure piston, and the pressure required on the low-pressure piston to do the same work as that done by pressure pu on the high- pressure piston will be ^— . R In the case of a 2x engine therefore the total M.E.P. referred to the low-pressure cylinder is Vr^+Vl (64) This mean effective pressure acting on the low-pressure piston only would give the same indicated horse-power as is obtained with the two cylinders of the engine. In designing compound engines it is customary to determine the size of the low-pressure cylinder as though it were to do all the work expected of the engine by receiving steam at the highest pressure available and exhausting it at the lowest. The mean effective pressure which would thus be assumed to exist is the referred value p R just ex- plained. Having found the size of the low-pressure cylinder COMPOUNDING 155 and the value of the referred M.E.P. the size of the high- pressure cylinder can be determined so that the work done by each cylinder will be just half of the total for which the engine is being designed. This size will have to be such that the high-pressure mean effective pressure referred to the low-pressure cylinder (i.e., p H +R) is equal to half the total mean effective pressure referred to that cylinder. That is, the size will have to be so chosen that f=f (65) ILLUSTRATIVE PROBLEM A double-acting compound engine is capable of developing 500 I.h.p. The stroke is 18 ins.; revolutions per minute, 175; mean effective pressure referred to L.P. piston, 45 lbs. per square inch; cylinder ratio, 3^. Find cylinder diameters. From pLan I.h.p. 33,000' 500X33,000 aLP - = 45Xl.5Xl75X2 =700(aPPrOX ' ); that 4. — =30 ins. (approx.), with the cylinder ratio equal to 3j, «h.p. =-7tt =200 sq.ins., 3.5 d H .p. = \ i™ = 16 ins. (approx.). 76. Combined Indicator Diagrams. When a compound engine is indicated, the diagrams of the two cylinders as drawn by the indicator are not directly comparable. The scales of pressure and volume are different on the two dia- grams, and correction must be made for this fact before the 156 STEAM POWER diagrams can be compared. It is customary to do this and to draw the average high-pressure and low-pressure diagrams on the same set of coordinates in order to determine how well they approximate the ideal diagram that would be obtained in one cylinder operating between the extreme limits of pressure. Diagrams approximating those that would be obtained from high- and low-pressure cylinders are shown at (h) and (I) respectively, in Fig. 95, and the result of drawing Fig. 95. both to the same scales is shown at the left of this figure. The curves xn and xi show the variations of quality along the two expansion curves. Drawing the two diagrams to the same scales in this way is known as combining the diagrams and the result is known as a combined diagram. The curves SS and S'S' added to the combined diagram are saturation curves. They do not, in general, form a continuous curve, because of the different quantities of steam contained in the two clearances and because any COMPOUNDING 157 moisture in the high-pressure exhaust is generally removed in the receiver. The volumes occupied by clearance steam at initial pressures are indicated by the points b f and B' respectively. The lengths b'S and B'S f approximately represent the volumes that would be occupied by cylinder feed when in each cylinder if dry and saturated. A combined diagram for a triple-expansion engine is shown in Fig. 96. The heavy lines give diagrams con- structed so as to represent as nearly as possible what may be expected to occur in the cylinders of such an en- gine, assuming perfect valve action and hyperbolic expan- sions and compressions. The dotted diagrams indicate the shapes that would be drawn by indicators applied to the real cylinders. The numer- ous sharp angles are due to overlapping of events, one cylinder suddenly starting to draw from a receiver while another is exhausting. It will be observed that the dotted diagrams do not contain any of these sharp angles, but that their general outline forms a fair average of them. The curve cd is a rectangular hyperbola drawn as a continuation of the assumed hyperbolic expansion line of the high-pressure cylinder. The failure of the expansion lines of the other cylinders to fall upon this curve is ex- plained by quality changes, different quantities of clearance steam in the different cylinders and withdrawal of moist- ure from steam exhausted to receiver before admission to the following cylinder, — .d Fig. 96. 158 STEAM POWER PROBLEMS 1. Find the size of the cylinders of a double-acting compound engine, which is to give 600 I.h.p., when using steam at a pressure of 150 lbs. per square inch absolute, and having a back pressure of 2 lbs. per square inch absolute. The cylinder ratio is to be 4, and the total ratio of expansion 12, piston speed 750 ft. per minute, and R.P.M. =150; diagram factor is 80%. 2. Given a 200 H.P. compound Corliss engine with cut-off in the H.P. cylinder at 60% stroke. Ratio of expansion is 7; clearance is 7%; card factor is 70%; pressure at the H.P. cyl- inder is 165 lbs. absolute. Find (a) Cylinder ratio; (b) Theoretical and actual M.E.P.; (c) Determine size of four engines, and select the best one. l-f-% Q\ Note - /= % (c.o.)°+%ci Xcyl - ratio - 3. Given a compound engine 18X40 ins., having a stroke of 28 ins. Steam pressure is 165 lbs. per square inch absolute; cut-off in H.P. cylinder occurs at 62% stroke; clearance equals 16%; back pressure equals 5 lbs. ; R.P.M. equal 150. Find (a) Cylinder ratio; (b) Ratio of expansion; (c) Actual M.E.P.; (d) I.h.p. CHAPTER X THE D-SLIDE VALVE 77. Description and Method of Operation. The simple D-slide valve, shown in place in Fig. 97, is so named because of the similarity of its section to the letter D. It is located in the steam chest, rides back and forth upon its seat and Fig. 97. serves to connect the two ports alternately with steam and exhaust spaces respectively in order to give the neces- sary distribution of steam. The valve has to perform the following functions for each end of the cylinder during each revolution of the engine : (1) It connects the proper port to the steam space or 159 160 STEAM POWER steam chest at such a time that steam can enter the cylinder as the piston moves away from the head. (2) It shuts off this port and thus cuts off the supply of steam when the piston has completed a certain definite fraction of the stroke. (3) It connects the port with the exhaust cavity shortly before the piston reaches the end of the stroke, thus effecting " exhaust " or " release "; and (4) It shuts off the port again when the piston has com- pleted the proper fraction of the next stroke, thus trapping in the cylinder the steam which is compressed during the remainder of the stroke. Engine Crank rl Main Co nnecting Fig. 98. It is obvious that the valve must be reciprocated upon its seat and that its motion must be connected with that of the piston in some way so that the proper phase relation may be retained. This could be effected by the system shown diagrammatically in Fig. 98, a small crank operating on the end of a connecting rod giving the valve its short stroke just as the main crank fixes the longer stroke of the piston. Such an arrangement would, however, be very inconvenient with many real engines, as the valve would be located too far from the center line of the cylinder. It is customary to use what is knbwn as an eccentric for the purpose of operating the slide valve. The parts and arrangement of an eccentric, together with an illus- tration of the way in which it is mounted on the shaft of Strap Fig. 99.— Parts of Eccentric. 161 162 STEAM POWER THE D-SLIDE VALVE 163 L an engine are shown in Figs. 99, 100 and 101. The motion it gives the valve is exactly the same as that imparted by the crank first assumed, and it can easily be shown that it is the exact equivalent of such a crank. Assume, for example, a crank such as that shown in Fig. 98 with a length of arm or throw equal to a. If the crank pin is made larger while other parts of the crank remain the same, as shown in Fig. 102, the crank mech- anism is not essentially altered; the mo- tion which it would impart to a connecting rod is not changed. If this process of enlarging the pin be continued Fig. 101. — Eccentric on Vertical Engine. (<0 (d) 1 Fig. 102. — Equivalence of Crank and Eccentric. until the pin has become large enough to surround the shaft and if the crank arm be then removed so that what was the crank pin is fastened directly on the shaft, an 164 STEAM POWER ^sssssssss^s^s Fig. 104. — Slide Valve without Lap, Live Steam Space. Exhaust Cavity Connected to Exhaust J?ipe Fig. 105. THE D-SLIDE VALVE 165 eccentric results. It is the exact equivalent of the original crank; its center, which is the center of the crank pin, revolves about the center line of the shaft in a circle with a radius a just as in the original mechanism. The eccentric makes it possible to place a short crank (short arm) upon a large diameter shaft without having to cut the shaft away as shown in Fig. 103, and it is therefore very useful for driving valves. 78. Steam Lap. The simplest possible form of D-slide valve would just reach the outer edges of the ports when in its central position as shown in Fig. 104. The crank driving it (that is the crank equivalent to the eccentric which would probably be used in a real case) would have to be located 90° ahead of the engine crank in the direction of rotation, as can easily be seen by consulting Fig. 105, which illustrates the mechanism in various critical positions. The illustration shows that such a valve would give full stroke admission, thus producing a rectangular cycle which has already been shown to be very inefficient as a means of obtaining work from the heat used in i> • , Steam-Space IOrmmg Steam. ^Valve Travel? If cut-off is to occur before the - ^^P^T^"-^— ^ - end of the stroke, the edge of the ^^ JXdm valve must return and close the port cSy* before the piston reaches the end of its stroke. But since the crank mechan- - Piston Travel ism does not permit the valve to Fig. 106. remain stationary in any one posi- tion, such early cut-off could only occur if the valve over-traveled, as shown in Fig. 106, and this would un- fortunately result in connecting the working end of the cylinder to exhaust and in admitting steam to the other side of the piston at such a time as to oppose the piston's motion. The solution of the difficulty lies in making the valve longer, so that when in its central position it overlaps the outer edges of the ports as shown in Fig.. 107.. 166 STEAM POWEE The amount of overlap of the outer edge is called the out- side lap, and when steam is admitted by the outer edges of the valve, as in the case under discussion, it is also called the steam lap. With such an arrangement the valve must be drawn out of its central position by the amount of the lap when the piston is at the end of its stroke as shown by a in Fig. Fig. 107.— Steam and Exhaust Lap. Fig. 108. — Lap a and Lap Angle a. 108 in order that steam may be admitted just as the piston starts to move. It follows that the crank driving the valve must be more than 90° ahead of the engine crank and that it must be ahead by the angle required to move the vaive a distance equal to the outside lap. This angle, represented in the figure by a, is called the lap angle. 79. Lead. In real engines it is further desirable to start the admission of steam just before the piston arrives at the end of its stroke. This assists in bringing the moving parts to rest, raises the pressure in the clearance to full value before the piston starts, and gives a wider opening through which the steam can flow during the early part of the stroke, thus reducing wiredrawing and loss of area at the top of the diagram. If the valve is to open before the piston reaches the end of its stroke, the crank driving it must be shifted still further ahead of the engine crank. It must be shifted ahead by an angle which will draw the valve through the distance which will give the desired opening of valve with the piston at the end of its stroke as shown by b in Fig. 109. The angle required, indicated THE D-SLIDE VALVE 167 by j8, is known as the angle of lead, and the width of the steam opening with engine crank on dead center, i.e., the distance b, is known as the lead. The lead varies from less than t^ in. on small engines and with low speeds up to over J in. on large engines and with very high speeds. 80. Angle of Advance. The eccentric or valve-operating crank must be ahead of the engine crank by an angle equal to 90°+ angle of lap a + angle of lead j8, as can be seen Fig. 109. — Lead b and Lead Angle /?. by an inspection of Fig. 109. The sum of a and is called the angle of advance and will be represented by 8. This is the number of degrees in excess of 90 by which the eccen- tric leads the engine crank. Fig. 109 shows that cut-off in an engine fitted with a valve having lap and lead must occur when the engine crank has turned through an angle equal to 180 — 2a, because the valve will then have returned to the closed position. Apparently, cut-off can be made to occur at any point in the stroke by properly choosing the value of a, but it will be discovered later that the exhaust events set a limit to increase in the value of this angle and hence do not per- mit of cut-off occurring earlier than a certain fraction of the stroke. 81. Exhaust Lap. Inspection of Fig. 105 will show that the simple valve without lap originally assumed will give no compression, because the cylinder end is connected to the exhaust cavity for the entire stroke. Inspection of all 168 STEAM POWER the changes which have been suggested in the subsequent paragraphs will show further that if the inner edges of the valve are left in the original positions the exhaust events will be considerably distorted in the case of a valve having steam lap and lead. This trouble may be remedied by moving the inner edges of the valve closer together, making the exhaust cavity in the valve shorter and giving inside lap as shown in Fig. 107 by b. When the inner edges of the valve control exhaust, as in the case of the valve under discussion, this inside lap is also called exhaust lap. The length of the valve, the lap and the lead are gen- erally chosen so as to give the desired arrangement of admission and cut-off and then the exhaust edges are so located as to give desirable release and compression. In some forms this necessi- tates the use of an exhaust cavity in the Fig. 110. valve such as that shown in Fig. 110. The amount by which the edges of the valve fail to meet the inner edges of the port is spoken of as negative inside lap. This dimension is indicated by c in the figure. It should be noted particularly that all measurements of lap are made with the valve central on its seat and that the measurement of lead is made with the piston at the end of its stroke, i.e., with the engine crank on dead center. 82. The Bilgram Diagram. The action of all slide valves could be studied by means of drawings of the actual mechanism, as has been done in preceding paragraphs, but such a method is time and space consuming. Numerous diagrams such as the Elliptical, the Sweet, the Zeuner and the Bilgram have been developed for the purpose of simpli- fying and expediting such studies and, when properly understood, they are very convenient. The scope of this book does not permit a discussion of all of these diagrams THE D-SLIDE VALVE 169 and, since the Bilgram diagram is probably the most gener- ally applicable, attention will be confined to it. The construction of this diagram is illustrated in Fig. 111. The point represents the center of the engine crank shaft and the two circles drawn about this point as a center represent respectively the paths traveled by the Fig. 111. pin of the valve crank and the pin of the engine crank. These circles are drawn to any convenient scales. The diagram is conventionally drawn in such a way that the line OM represents the head end dead center position of the crank and in all subsequent paragraphs the relative positions shown by the small sketch in Fig. Ill will be assumed. The cylinder will be assumed to the 170 STEAM POWER left of the shaft and the engine will be assumed to run " over." With the crank in position OM , the eccentric (equivalent crank) must be in the position OB, ahead of the crank by an angle 90°+«+j3 = 90+<5. The valve must then be displaced to the right of its central position by an amount represented by the distance DB, if a small correction for " angularity " of the valve connecting rod be neglected. As rotation continues, horizontal distances corresponding to this line will always give the instantaneous valve dis- placements. For position OB' , for instance, the valve dis- placement will be D'B' '. If the angle 8 is now laid off above OX, locating the point Q as shown, a perpendicular QE dropped upon OX from this point will equal in length the line DB, and will therefore show the valve displacement when the crank is in head end dead center position OM. This must be true, because the triangles QOE and BOD are similar and have the sides OQ and OB equal to the radius of the same circle. The perpendicular QE is really a perpendicular dropped upon the extension of the line representing the crank posi- tion, and it is a general property of this diagram that a line starting at Q and perpendicular to the line representing any chosen crank position (or an extension of that line) will show by its length the displacement of the valve when the crank is in the chosen position. Thus assume the engine crank to rotate through the angle 7 to the position OM'. The eccentric will have rotated to B' and the valve dis- placement will be represented by D'B' . A perpendicular drawn from Q upon OX', the extension of the crank posi- tion, gives QE' equal to B'D' and hence representing the valve displacement to the same scale. This construction drawn for different crank positions OA, OM, OM x , OM 2 , etc., is shown in Fig. 112, the dash- dot radial lines about Q representing the various values of the valve displacement. The number on each of thesf. THE D-SLIDE VALVE 171 lines indicates the crank position to which it corresponds. It will be seen that the displacement increases in value until the crank position OM3 is reached, after which it decreases again. Steam Lap Circle- Steam Lead Fig. 112. Since the opening ment minus the lap, as by which the valve is found by subtracting placement the amount head end dead-center to lap plus lead, and is to steam is equal to the displace- shown in Fig. 109, the actual amount open for any crank position can be from the corresponding valve dis- of lap possessed by the valve. For position, the displacement is equal shown by QE in Fig. 112. Subtract* 172 STEAM POWER ing the lead EF, the remainder FQ gives the lap of the valve. A circle drawn about Q with radius equal to QF (or a circle drawn about Q and tangent to the line L) will cut off of the lines representing valve displacement the amount representing the part of each displacement used in over- running the lap of the valve. The remainders, that is the parts of the lines radiating from Q in Fig. 112 which are outside of the lap circle, must then represent the amounts by which the valve port is actually open. It will be observed that the valve is open by the amount of the lead when the crank is on dead center, position OM. The crank position for which the valve displacement is just equal to the lap, and hence at which the valve is just begin- ning to open, can be found by drawing a tangent through to the lower side of the lap circle and then extending it to give the crank position OA in Fig. 112. As the crank rotates clockwise from this position, the valve opens wider until, when position OM3 is reached, the greatest valve opening exists. Further rotation results in partial closure of the valve and, when the crank has finally rotated into position OC, the valve has just closed, that is, cut-off has occurred, the displacement being just equal to QG, the steam lap. Thus this diagram, as so far developed, indicates crank positions for admission and cut-off and the values of valve displacement and valve openings for all intermediate crank positions. ILLUSTRATIVE PROBLEM A certain valve has an external steam lap equal to 1J ins. The lead is ^ in. and the throw of the eccentric is 2\ ins. (a) Con- struct such parts of the Bilgram diagram as are necessary to indicate ^head end" crank positions for admission, maximum valve opening and cut-off. (b) Indicate on this diagram the amount of valve opening at various crank positions between admission and cut-off. (c) Determine the value of the angle of advance. THE B-SL1DE VALVE 173 Draw a circle with radius equal to the eccentric throw, 2\ ins., using any convenient scale. This circle is designated by abed in Fig. 113. Draw about the same center another circle of any convenient size. Draw in the horizontal diameter ac and extend as shown. On the right-hand side of the circle draw the line ef, Steam Lap Circle Fig. 113. parallel to the horizontal axis and a distance above it equal to the lead, Y6 i n -, to the same scale as that chosen for eccentric circle. The steam lap circle must have its center Q on the upper right- hand quadrant of the eccentric circle, and it must be tangent to the line ef. Its radius must equal the steam lap, If in. to scale. Therefore, with compass points set the proper distance apart, find the center Q, about which a lf-in. radius circle will just be tangent to the line ef, and draw the steam lap circle. 174 STEAM POWER The crank position at admission is found by drawing the line AO so that, if extended, it is tangent to the lower side of the steam lap circle. The crank position at cut-off is found by drawing the line H.E. Adm I.E. Compression Fig. 114. M""0 in such position that it is tangent to the upper part of the steam lap circle. The crank position for maximum valve opening is found by drawing the line M"0 in such position that a line through QO will be perpendicular to it. The amount of valve opening at this THE D-SLIDE VALVE 175 crank position is shown by the length of the part of this per- pendicular line outside of the steam lap circle, i.e., the distance Og interpreted according to the scale chosen for eccentric and steam lap circles. When the crank is in position M'O, the length of hi, interpreted to scale, gives the amount by which the valve is open to steam. When the crank is in position M'"0, the length of jk, inter- preted to scale, gives the amount by which the valve is open to steam. The angle indicated by 5 is equal to the angle of advance because of the property upon which the construction of this diagram is based. 83. Exhaust and Compression. The exhaust edge events can be shown on the Bilgram diagram by a method similar to that used for the steam edge events. The direction in which valve displacements occur are indicated in the upper part of Fig. 114 in which the crank and eccentric circles have been drawn to such scales that they coincide. In- spection of the small sketch in the lower part of the figure will show that head end release must occur when the valve has traveled a distance equal to the inside lap to the left of its central position. A crank position OR drawn tangent to the lower part of a circle about Q with radius equal to the inside lap will, therefore, be the crank position at re- lease. Clockwise rotation from this position will result in a wider opening to exhaust until position OM\ is reached, after which the valve will begin to close. Final closure will occur when the crank reaches position OK, the exten- sion of which is tangent to the top of the exhaust lap circle. At that time the valve will have returned (moving from left to right) and will still have to move a distance equal to the exhaust lap before attaining a central position. ILLUSTRATIVE PROBLEM Given the exhaust lap of a D-slide valve equal to f in.; the steam lap \\ ins.; the throw of the eccentric, 2 ins.; and the lead \ in. Find the angle of advance, the maximum port opening to steam and to exhaust, and the crank positions of cut-off, release, compression and admission for the head-end of the cylinder. 176 STEAM POWER Draw the eccentric (and crank) circle with a radius equal to 2 ins., and draw the horizontal diameter as in Fig. 115. Draw a horizontal line in the upper right-hand quadrant at a distance of |+1J ins. above the horizontal diameter. Locate the point Q at intersection. Fig. 115. Draw the steam lap circle with a radius 1^ in. and the exhaust lap circle with a radius f in. The angle of advance is the angle between OQ and the hori- zontal. The maximum opening to steam is given by the distance Oa=f in. The maximum opening to exhaust is given by the distance 06 = If in. The crank positions shown are obtained by drawing lines THE D-SLIDE VALVE 177 tangent to the lap circles. A represents admission; C, cut-off; R, release, and K, beginning of compression. The piston positions at the times of these events are given to reduced scale by vertical projection. 84. Diagram for Both Cylinder Ends. The complete diagram for the head end of cylinder is shown in Fig. 114 with all critical crank positions marked. The positions for the crank end of the cylinder can be found in a similar way by constructing a diagram in which the point Q and the lap circles are located in the opposite quadrant. The resulting Fig. 116. diagram for both cylinder ends, with laps the same for both ends of the valve, is given in Fig. 116. 85. Piston Positions. The valve events might be studied entirely in conjunction with crank-pin positions, but it is more convenient and customary to consider them in connec- tion with piston positions. Piston positions corresponding to different crank-pin positions could be found by drawing the mechanism to scale for each different position as shown in Fig. 117 for piston positions 1 and 2. It is obvious that this would involve a great deal of work and that, if drawn to large scale, it would consume a great 178 STEAM POWER ! \ 8 © e deal of space. Further, it is convenient to be able to locate relative piston positions on the line which serves as the hori- zontal diameter of the crank circle of the Bilgram diagram. The method used depends upon the fact that the motion of the crosshead is exactly the same as that of the piston, so that if the motion of the crosshead end of the connect- ing rod can be followed, it will be equivalent to following the motion of the piston itself. It should also be noted that the diameter of the crank cir- cle must be equal to the stroke of the engine. Assume now, that the point b in Fig. 117 be taken to rep- resent the position of the pis- ton when it is really in posi- tion 1. When the piston has moved to position 2, the cross- head will have moved from a to a' and the crank pin from b to b' . If with a' as a center the connecting rod be swung down to the horizontal its right-hand end will arrive at the point c. The distance be must then represent the dis- tance that crosshead (and pis- ton) have moved from dead- center position because ab and THE D-SLIDE VALVE 179 a'c both represent the length of the connecting rod and c must therefore be as far to the right of b as a' is to the right of a. The point c may therefore be taken to represent pistor' position when the connecting rod is in the position a'b'. In general, if the horizontal diameter of the crank circle be taken to represent the stroke of the engine, the pis- ton position corresponding to any crank position can be found by taking a radius equal to the connecting-rod length (to the same scale as the circle) and striking an arc from the Fig. 118. crank-pin position, using a center on the horizontal line on the cylinder side of the crank circle. An approximate method is also used for finding the piston position. Instead of projecting down from the crank-pin position with an arc, such as Vc in Fig. 117, a vertical line through the crank-pin position is used. Such a line would give c' as the piston position when c is really correct. This method would give accurate results with a connecting rod of infinite length. For ordinary lengths of rod, however, the results are far from correct. The error is said to be due to the angularity of the connecting rod. The effect of the angularity of the connecting rod is shown in Fig. 118 for different positions. On the outstroke the piston is always farther ahead than the rectilinear pro- 180 STEAM POWER jection would indicate. On the return stroke the piston is always behind the position indicated by rectilinear projection. Fig. 119. 86. Indicator Diagram from Bilgram Diagram. Since the piston positions corresponding to different crank posi- THE D-SLIDE VALVE 181 tions can be determined, it is a comparatively simple matter to construct the indicator diagram which theoretically would be given by an engine fitted with a valve of certain dimensions. It is necessary to assume the upper and lower pressure and also to assume the form of the expansion and compression curves. These are generally taken as rectangular hyperbolas. The method of constructing an indicator diagram from the Bilgram diagram is shown in Fig. 119. The crank- pin positions for admission (A), cut-off (C), release (R) and beginning of compression (K) are first found. These pin positions are then projected to the horizontal diameter by means of arcs with radius equal to the connecting-rod length and with centers on the line MM produced to the left. The intersections a, c, r and k indicate the piston positions at which the corresponding events occur. These are then projected vertically downward to intersect the proper pressure lines and the card is drawn through the intersections. Diagrams constructed in the same way, but for both head and crank ends, are given in Fig. 120. A symmetrical valve was assumed, that is, one built exactly alike on head and crank ends. The diagrams show that such a valve cannot give the same results for both cylinder ends because of the effect of the angularity of the connecting rod. It is most evident in the case of cut-off. The cut-off in this case occurs just before three-quarter stroke for the head end and just after half stroke for the crank end of the cylinder. All other events are distorted in the same way, but the actual lengths of the variations are not as great as in the case of the cut-offs and therefore the distortion is not as obvious. The effect of the angularity of the connecting rod upon the diagrams can be remembered easily if it is noted that all valve events occur later with respect to piston position on the outstroke and earlier on the instroke than they would with a connecting rod of infinite length. 182 STEAM POWER It is possible to " equalize " the cut-offs, that is, make them occur at the same fraction of the stroke by using unequal steam laps at opposite ends of the valve, but this will result in still further distortion of admissions, as can be seen by constructing a Bilgram diagram for this case. Similarly, the compressions can be equalized by the use of •pression H.E. Admission C.E. Admission Fig. 120. unequal exhaust laps, but this results in distortion of the release events. Various linkages have been developed which are so arranged that they distort the motion of the valve to just the extent necessary to counterbalance the effects of the angularity of the connecting rod. The scope of this book does not, however, permit a discussion of such valve gears. THE B-SLIDE VALVE 183 87. Limitations of the D-slide Valve. The simple valve discussed in the preceding paragraphs has numerous limitations and is therefore only used on small and cheap engines, or in cases where economy in the use of steam is not essential. This valve, when used with steam entering over the outside edges as previously considered, is pressed to its seat by the live steam acting over its entire upper surface. This pressure is practically unbalanced, as the greater part of the lower surface of the valve is subjected to the low pressure of the steam being exhausted. As a result the friction to be overcome in moving the valve is very great and there is an appreciable loss from this source. Further, the shape of the valve makes necessary the use of long ports which form part of the cylinder clearance and which are alternately exposed to live and to exhaust steam with results previously discussed. These ports can be decreased in length by increasing the length of the valve, but this in turn increases the area exposed to high pressure and hence increases the friction loss. It can be shown by means of the Bilgram diagram that, if a cut-off earlier than about f stroke is desired, the angle of advance, the amount of steam lap and the size of the eccentric must all be made very great. This results not only in large friction losses, but also in very early release and compression, because of the great angle of advance. As a result, slide valves of the simple D type are seldom used when a cut-off earlier than \ to \ stroke is desired. It should be remembered in this connection that the simple engine generally gives its best economy with a cut-off of about J stroke. The drawing of lines representing the opening of the valve to steam as in Fig. 112 will show that this simple valve is further handicapped by the very slow opening and closing of the steam ports, causing a great amount of wire drawing with a corresponding loss of diagram area. In order to get an adequate opening to steam the valve 184 STEAM POWER .Steam Spaces, .Exhaust i must also be given a great displacement and, since this occurs under great pressure, it results in great friction loss. The unbalanced feature can practically be overcome by rolling up the valve and ports about an axis parallel to the length of the cylinder. This gives what is known as a pis- ton valve, shown dia- grammatically in Fig. 121 It can also be par- tially overcome by using a balance plate or ring of some kind between the top of the valve and the inside of the steam-chest cover, so arranged that live steam is excluded from the greater part of the upper surface of the valve. Valves of this type are generally called balanced slide valves and are used on many high- and medium-speed engines. The valve travel required for obtaining a given opening ■Steam Ports Fig. 121— Piston Valve. Fig. 122.— Allen Double Ported Valve. can be decreased and the rate of opening and closing can be increased by the use of multiported constructions. These are so arranged that two or more ports open or close at the same time, so that the total movement required for a given opening is divided by the number of ports and the rate of opening and closing is multiplied in the same proportion. One simple type of double-ported valve is illustrated in Fig. 122. When several ports are used the valve often becomes THE D-SLIDE VALVE 185 a rectangular frame crossed by a number of bars and is known as a gridiron valve, because of its appearance. Such valves are often combined with balance plates and give very satisfactory results. A number of designs of slide valves have been developed for the purpose of making cut-off independent of the other events. Many of these use a separate cut-off valve which either controls the steam supply to the main valve or else rides on the main valve and controls cut-off by covering ports in that valve. Devices of the latter type are called riding cut-off valves. They are either driven by separate eccentrics, or by linkage from the eccentric controlling the main valve, the linkage being so arranged as to give the proper relative motion between main and auxiliary valves. In such designs the main valve is proportioned so as to give the desired admission, release and compres- sion and the cut-off is then taken care of by proper adjust- ment of the cut-off valve, 88. Reversing Engines. It was shown in one of the early paragraphs of this chapter that the eccentric must be set 90°+ angle of advance ahead of the crank, ahead meaning in the direction of rotation. To cause the engine to revolve in the opposite direction, that is, to " reverse " the engine, it is therefore only necessary to shift the relative positions of eccentric and crank so that the eccentric leads the crank by 90°+ 8 in the new direction of rotation. This corre- sponds to shifting ahead (in first direction of rotation) through an angle equal to 180 — 25 or shifting backward through an angle equal to 180+25, as can be seen by inspec- tion of Fig. 109. In practice it is generally more convenient to use two eccentrics, one set properly for rotation in one direction and the other set properly for rotation in the opposite direc- tion. This arrangement is shown diagrammatically in Fig. 123. This figure is drawn for a vertical engine and in such position that the engine is on crank-end dead center. 186 STEAM POWER Reverse or Weight Shaft Valve Stem The point P represents the position of the center of the crank pin; the point / represents the position of the equivalent crank (center of eccentric) which / / ^v \ drives the valve for " forward," " ahead " or clockwise rotation; and the point b represents the position of the equivalent crank which drives the valve for " backing/' " reverse," or counter-clockwise rotation. The real mechanism, in one of its numerous forms known as the Stephen- son Link Gear, is shown in perspective in Fig. 124. The forward eccentric corresponds to / of Fig. 123 and the backing eccentric corresponds to b of that figure. The eccentric rods are fastened to opposite ends of a curved " link " and move the valve through a " link block " fastened to the end of the valve stem. In the position shown in the figure the link is in such position that the forward eccentric operates practically directly on the valve stem so that the valve motion is practically entirely governed by that eccentric. If the reverse shaft were to be rotated clockwise into the backing position, the " sus- pension rods " would pull the link over until the eccen- Clockwise Rotation Fig. 124. — Stephenson Link Gear. trie rod of the backing eccentric was directly under the valve stem. Under such conditions the valve motion would be controlled almost entirely by the backing THE D-SLIDE VALVE 187 eccentric and the engine shaft would rotate counter-clock- wise. If the mechanism were so set that the link block occupied a position on the link between the ends of the two eccentric rods, the valve motion would be controlled by both eccentrics and would be a compromise between the motions given by either eccentric separately. It is characteristic of this gear that the cut-off is latest when either one or the other eccentric is fully " in gear " and that it becomes earlier as the link block approaches the center of the link. With the link block in the center of the link the valve does not open at all, i.e., the cut-off occurs at zero stroke. There are numerous other forms of link gears, the best known being the Gooch, the Allan and the Porter- Allen. There are also numerous reversing mechanisms known as radial gears in which the motion of the valve is controlled by means of a " radius rod " which can be set to give the desired valve motion. • The valve motion is obtained in- directly through the radius rod from an eccentric, from the crank, or from the connecting rod. The limits of this book do not permit a detailed discussion of these forms. 89. Valve Setting. From what has preceded it will be evident that it is not only necessary that a valve and its seat and driving mechanism be correctly designed, but also that the various parts must be correctly connected up in order that the valve may move in its proper phase rela- tion with respect to the piston. Adjusting the mechanism in such a way that the proper phase relations are obtained is known as setting the valve. This can be done with fair accuracy by a simple study of the mechanism in various positions, as will be shown below, but it is always advisable to check the setting by means of indicator diagrams taken after the setting is completed. Such diagrams will often show errors of such character or size that they cannot be determined by measurement on an engine which is not operating. 188 STEAM POWER Before beginning operations it is always advisable to go over the entire engine carefully and to eliminate excessive lost motion at all pins and bearings in order that the relative positions of parts obtained while setting the valve may approximate those which will be obtained when the engine is in operation. The effect of lost motion will be appreciated after a study of Fig. 125. Assume that all parts of the mechanism are tight except the crank-pin end of the con- Fig. 125. necting rod as shown. If, for instance, the engine is rotated by hand by turning the fly-wheel, the crank will pull the piston mechanism and the piston will be drawn into the position shown in the upper half of the figure when the crank has turned through an angle a. On the other hand, when the engine is operating under steam, the piston will push the crank pin around and will occupy a position such as that shown in the lower half of the figure when the crank has been turned through the same angle a. Obviously, the piston can occupy two very different positions for the same crank position, and a valve setting based upon the conditions shown in the upper part of the figure might be THE D-SLIDE VALVE 189 very incorrect when used under the conditions shown in the lower part of the figure. Lost motion in any part of the mechanism can produce analogous results and it is therefore necessary to remove as much of it as possible before attempting to set the valve. It is practically impossible to eliminate all lost motion, as there must be sufficient clearance at all bearing surfaces to accommodate a film of oil, and this alone would make necessary the taking of indicator diagrams for the check- ing of valve settings, even if it were possible to set perfectly by measurement for stationary conditions. In general, there are two adjustments which can be made in setting a plain slide valve. The length of the valve stem or eccentric rod can be changed and the eccentric can be shifted around the shaft. It is necessary to under- stand the effects of each of these adjustments. Changing the length of the valve stem is equivalent to shifting the valve upon its seat without moving the engine as shown in Fig. 126. In this figure the valve is shown in its central position by full lines. The lap is the same at both ends. If, now, Fig. 126. the valve is worked to the right upon its stem by adjustment of the nuts shown, until it reaches the dotted position, the head-end lap will have been de- creased and the crank-end lap will have been increased by the same amount. This would make admission earlier and cut-off later for the head end and admission later and cut- off earlier for the crank end. Obviously, the effects of changing the length of the valve stem are opposite for the two ends of the cylinder. Shifting the eccentric about the shaft simply changes the time relation between valve motion and piston motion; it does not alter the valve motion itself. If difficulty is experienced in realizing the truth of this statement, it is only necessary to draw several Bilgram diagrams for the 190 STEAM POWER same valve, but with different angles of advance, and then to construct indicator diagrams for both cylinder ends in every case. It will be discovered that shifting the eccentric ahead in the direction of rotation, for instance, will make all events occur earlier with respect to piston position for both ends of the cylinder. In setting a plain slide valve which is built symmetrical about a central axis, i.e., same inside and outside lap at each end, it is first necessary to adjust the length of the valve stem. This may be done by removing the steam-chest cover so as to expose the valve and then rotating the engine slowly by hand and observing the distance traveled by the valve on each side of its central position. This is con- veniently done by observing the distance between the outer edge of the steam port and the outer edge of the valve when the valve is fully open at each end. If the valve travels further toward the head end than it does toward the crank end, with reference to the port edges, the valve stem must be shortened; if it travels further toward the crank end the stem must be lengthened. In making these adjustments it is advisable to turn the engine only in the direction in which it is going to rotate, so that any lost motion in the valve mechanism will have approximately the same effect as when the engine is opera- ting. When the length of the valve stem is correctly adjusted, the eccentric must be so set on the shaft as to give the proper angle of advance. This is commonly done by shifting it about the shaft until the proper value for the steam lead has been obtained. In order to determine the value of the lead it is necessary to be able to set the engine on each dead center. This can be done approximately by turning the engine until the crosshead has come to either end of its stroke, but it will be found by trial that the fly-wheel and shaft can be turned through a very large angle at each end of the stroke without causing an appreciable motion of the crosshead, THE D-SLIDE VALVE 191 so that this method is not very satisfactory for the purpose of adjusting the eccentric. It is customary, therefore, to work in such a way as to give a more accurate determina- tion of shaft and crank positions for dead center. The engine is rotated until the crosshead has been brought near one end of its stroke, as shown in Fig. 127, and a mark is then scribed across the crosshead and guide as at ab. An arc xy is then marked on the fly-wheel by means of a tram such as that shown, the end c being placed *Mmti& K-Trai Fig. 127 at point P on some solid part of foundation or floor. The engine is then rotated, clockwise in the figure, until the crosshead has reached the end of its stroke and returned to such a point that the marks on crosshead and guides again coincide, as shown by dotted positions in the figure. The arc x'y' is then scribed on the fly-wheel with the tram, the end c again bearing on the point P. A point z is then found by bisecting the arc ef and when this point is brought under point d of the tram the crank will obviously be at crank-end dead center and the piston at the crank end 192 STEAM POWER (a) Perfect Cards for Slide Valve Type. (b) Actual Card; Small Engine. Center Line of Valve on Center Line of Seat; Eccentric Advanced to Give Normal Lead of 0.05 inch. Engine Running Over. (c) Same Setting as (6) except Engine Running Under. Fig. 128. THE D-SLIDE VALVE 193 (d) Angular Advaice of Eccentric Increased. Valve Stem Length Same as in 6) and (c). Lead 0.375 Inch, (e) Angular Advance of Eccentric Decreased so as to Give Negative Lead of 0.5 Inch. Length of Valve Stem Unchanged. (/) Length of Valve Stem Changed; Angle of Advance as in (6). Fig. 128. 194 STEAM POWER of its stroke. A point on the fly-wheel diametrically opposite to z is next found, so that when it is brought under point d of the tram the engine will be on head-end dead center. It is probable that more accurate results are obtained by rotating the engine in a direction opposite to that in which it rotates under steam, because lost motion is then taken up in the same direction as when working, but when the whole process of valve-setting is considered it is ques- tionable whether this is the correct direction of rotation. Opinion and practice differ in this respect. In the end, the setting should be checked by the taking of indicator diagrams, so that effects of incorrectible lost motion may be finally eliminated. With the dead-center points found the engine is placed on, say, head-end dead center, and the eccentric shifted until the valve is open to steam by the desired lead. The eccen- tric is then fastened in this position and the engine turned to the opposite dead center. Because of angularity of con- nections and of irregularities in valve and seat dimensions, it generally will be discovered that the valve is not now open to steam by the same amount as at the other end. If it is desired that it should be, the valve can be shifted on its stem about half of the distance by which it is out and the eccentric can then be swung about the shaft to take up the remaining distance. The effect should then be checked by putting the engine on the opposite dead center. Valves may be set for equal leads as above, or for equal cut-offs or for any sort of a compromise desired. In any case the procedure is about the same. The length of the valve stem is adjusted, then the eccentric position is adjusted, and then refinements are effected by small changes of both adjustments. Remember always, that changing the length of the valve stem changes events at opposite cylinder ends in opposite directions, while shifting the eccentric changes all events in the same direction. The effects of various adjustments are shown by the THE D-SLIDE VALVE 195 indicator diagrams given in Fig. 128. These diagrams were taken from a small, slide-valve engine and serve very well to show the way in which the indicator discloses poor adjustments. PROBLEMS 1. Given: angle of advance, 30°; throw of eccentric, If ins.; lead, r^ in.; maximum exhaust-port opening, 1| in.; find the steam lap, maximum opening to live steam, and the exhaust lap. 2. Given: steam lap of | in.; lead of ^ in.; exhaust lap of f in.; and the angle of advance equal to 30°. Find the valve travel ( =2 X throw of eccentric) and maximum port opening to steam and to exhaust. 3. An engine has an eccentric throw of If ins.; a steam lap of | in.; and a lead of ys hi- Compression begins at £ of the return stroke. Assume a connecting rod of infinite length and find the angle of advance, the exhaust lap, and the maximum port openings to steam and to exhaust. 4. Given: valve travel, 3 ins.; steam lap, f in.; exhaust lap, \ in. ; and lead, f in. ; find maximum port opening, angle of advance, and piston positions at cut-off, release, compression, and admission for both ends of cylinder, with the length of the connecting rod equal to 4| times the length of the crank. 5. It is required to build an engine having a steam-port opening of | in., a lead of ^ in., and a connecting rod four times the length of the crank. Cut-off must occur at § stroke and release at 95% of the stroke. Find the inside and outside lap, the throw of the eccentric and the fraction of stroke completed by the beginning of compression. CHAPTER XI CORLISS AND OTHER HIGH-EFFICIENCY ENGINES 90. The Trip-cut-off Corliss Engine. The slide valve has certain limitations which can be partly, but never wholly, overcome. In most slide-valve gears, for instance, the various events occur more slowly than is desirable, and this is particularly true of cut-off. Ideal valves would open sud- denly to full opening when necessary and would close as suddenly at the proper time, and such action would give minimum throttling loss and rounding of corners of the diagram. Engines fitted with such ideal valves would therefore give indicator diagrams with maximum work area as shown by the dotted lines in Fig. 129, the full lines indicating the type of diagram obtained with the ordinary slide valve. Again, the simpler forms of slide valve involve the use of long ports connecting with the clearance space within the cylinder, thus adding greatly to the clearance surface exposed and to the cylinder condensation. These ports serve for both admission and exhaust, and their walls are therefore periodically cooled by the exhaust steam with the result that excessive condensation occurs during admission. Many attempts have been made to devise valve gears which should not be subject to the limitations of the 196 Fig. 129. HIGH-EFFICIENCY ENGINES 197 simple slide valve. Some of these have resulted in the development of the more complicated slide valves de- scribed in the last chapter, but such designs generally leave much to be desired. One of the earliest and most success- ful solutions was made by Corliss, who developed what is known as the trip-cut-off Corliss gear. The long combined steam and exhaust ports are elimi- nated by the use of four valves, two for steam and two for exhaust. These are rocking valves and are located top and bottom, at the extreme ends of the cylinder, with their longitudinal axes perpendicular to those of the cylinder, as shown in Figs. 48, 49, and 50. The exhaust valves are located below so as to drain out water of condensation. Details of valves of this type are shown in Fig. 130. These valves may each be regarded as an elementary slide valve which has a cylindrical instead of a flat face, and which is oscillated about a center near the face instead of being reciprocated, i.e., oscillated about a center at an infinite distance. The valves are operated as shown in Fig. 131 by short links from a wrist-plate pivoted on the side of the cylinder and rocked back and forth about its center by means of an eccentric operating through the linkage indicated. The locations of the various pins and the lengths of the various links are so chosen that the valves travel at high velocity when opening and closing, that they open very wide, and that they close only far enough to prevent leakage and then remain practically stationary until about to open again. Throttling losses are thus decreased and wear caused by useless motion after closure is minimized. The opening of the admission valves in this gear is effected positively by the linkage already explained, but they are closed differently. For opening, the steam link rotates the bell crank B in Fig. 132 and thus raises the latch C. The hook on the end of one of the arms of this latch engages the steam arm which is fastened on the end X98 STEAM POWER u I I HIGH-EFFICIENCY ENGINES 199 200 STEAM POWER of a rod which is slotted into the end of the valve. The valve is thus drawn further open as the wrist plate revolves, until the tripping end D of the latch strikes the cam indicated by E. This throws the hook out of engagement and thus disconnects the valve from the driving mechanism. The Q o btea Fig. 132.— Details of Corliss Trip-Cut-off Gear. valve is closed by the action of a dash pot, one form of which is shown in Fig. 131. As the steam arm rises during the opening of the valve it draws up the plunger or piston of the dash pot, leaving a partial vacuum beneath it, and, when the 20"x 48 Heavy Duty Corliss 110 Lb. Steam K.F.M. Fig. 133. valve is released by unhooking of the latch, atmospheric pressure drives the plunger down and thus causes cut-off to occur. The action of a dash pot is found to be unsatis- factory when the speed of the engine exceeds about 125 R.P.M. and most Corliss engines with trip-cut-off operate HIGH-EFFICIENCY ENGINES 201 at still lower speeds. Under such circumstances the cut- off is very rapid as compared with the piston speed, and the diagram shows a comparatively sharp corner at this point. A set of diagrams obtained from a large Corliss engine operating at low rotative speed is given in Fig. 133, and it is obvious that little throttling occurs. Because of the low speed at which these engines operate the stroke can be made long with respect to the diameter without attaining a prohibitive piston speed. The economy mentioned in Chapter VII as resulting from the use of long strokes can thus be obtained in these engines. An idea of the saving in steam effected by the partial elimination of throttling and condensation losses by means of the Corliss gear can be obtained from the curves in Fig. 134 (a) and (6), which give average performances. The position of the cam which determines the time at which cut-off occurs is controlled by the governor of the engine. When moved in the direction taken by the steam arm it causes cut-off to occur later. Variation of the point of cut-off is used in these and in most other engines to control the amount of work done per cycle in order that the engine may make available the quantity demanded at the shaft, as will be explained in a later chapter. It is there- fore desirable that the range of cut-off should be as great as possible, but it has been found very difficult to design trip-cut-off gears which will give a cut-off later than about 0.4 stroke if steam and exhaust valves are operated from the same eccentric. Later cut-off causes poor timing of the exhaust events. This has led to the introduction of Corliss engines with two eccentrics and two wrist plates per cylinder. One set operates the steam valves and the other the exhaust valves. With this arrangement the range of cut-off is unlimited. 91. Non-detaching Corliss Gears. Because of the low speed at which trip-cut-off Corliss engines are operated, 202 STEAM POWER 1 . so — -7 PROXIMATE STEAM CONSUMPTIO OF VARIOUS TYPES OF ENGINES (NON-CONDENSING) mple Single- Valve Throttling Engines (P = 100 lbs. mple Single- Valve Automatic Engines (P = 100 lbs. mple Four-Valve Automatic Engines (P = 100 lbs.) mdem and Cross-Compound Four- Valve and Corliss Engines (P = 100 lbs.) tndem and Cross-Compound Four- Valve and Corliss Engines (P = 125 lbs.) mdem and Cross-Compound Four- Valve and Corliss Engines (P= 150 lbs.) 3ntz Engine P= 133 lbs.; 92.7° Superheat; 248.5 1 R.P.M. 206. la-flow Engine, P= 150 lbs. na-flow Engine, P= 140 lbs.; Superheat, 110° F. fna-flow Engine, P= 150 lbs; Superheat, 300°. e 2 £„ II II II II 11 11 11 11 II A SSSS s s B "^^s u Is in d (N cc O 1 72 is PJ J_ T Zj 1 1 i p / / j / / r^ • > / / / s / y J raoH Jad " cTH'jl k >d> wi ■no nh urn no; ) ta^aj S 02 a) c! 'Sb H .3 '53. •8 1=1 o U 1 d o 55 a o O £ s HIGH-EFFICIENCY ENGINES 203 1 .ILL LI 1 1 APPROXIMATE STEAM CONSUMPTION OF VARIOUS TYPES OF ENGINES CCONDENSING) (1) = Simple Single-Valve Throttling Engines (P= 100 lbs.) (2) = Simple Single-Valve Automatic Engines (P = 100 lbs.) (3) = Simple Four- Valve Automatic Engines (P = 100 lbs.) (4) = Tandem and Cross-Compound Four- Valve and Corliss Engines (P= 100 lbs.) (5) = Tandem and Cross-Compound Four- Valve and Corliss Engines (P= 125 lbs.) (6) = Tandem and Cross-Compound Four- Valve and Corliss Engines (P= 150 lbs.) (») = Una-flow Engine (P= 140 lbs.). Superheat 100° F. («') = Una-flow Engine (P= 150 lbs.), Superheat = 300° F. 5 ie 2 d GQ CO i 1 ^O I j f _ / o 1 / / / / f / .-" fcC 3 5 8 8 § £ 2 •jnoH z^d d'H'I jsd •sq r j-uot^duinsuoo mv^s 204 STEAM POWER they are necessarily large, heavy and costly and efforts have been made to design gears which shall possess the advantages of the original Corliss mechanism without the limitation as to speed. In many models the Corliss valves are retained and are located in the ends of the cylinder as just described or in Fig. 135. — Non-detaching Corliss Valves Located in Cylinder Head. the cylinder heads as shown in Fig. 135. In some the wrist plate and the connecting links are also retained, but in others they are eliminated. In all engines of this type the admission valves are closed positively, the closure being effected by the same linkage that opens the valves to admit steam. Quick action is obtained by the arrangement of the operating mechanisms, the centers of rotation and the HIGH-EFFICIENCY ENGINES 205 lengths of links being so chosen that the valve travel is small when the valves are closed, that it is rapid when the valves are opening and closing, and that the valves remain practically wide open during most of the time that steam is being admitted. The advantages of small clearance and short and sepa- rate ports are attained in these arrangements and the operation of the valves is almost as perfect as that of the trip-cut-off gear. Engines fitted with these modified Corliss gears are operated at speeds considerably higher than those permissible with the older arrangement, and they may be classed with medium-speed engines. Engines of this type are generally known commercially as four- valve engines, but as this name applies equally well to the ordinary trip-cut-off gear and to others which will be described later, it is best to use some other designation. The term non-detaching Corliss engines seems to best describe them and is apparently gaining in favor. Non-detaching Corliss engines generally give diagrams intermediate between those obtained with the low-speed, trip-cut-off mechanism and those obtained from slide-valve engines with the simpler forms of valves, though the later designs very closely approximate the performances of the trip-cut-off Corliss engine. 92. Poppet Valves. Attention has already been called to the fact that the use of highly superheated steam is very effective in lessening or even eliminating initial con- densation. Experience has shown that it is very difficult to make large valves with sliding surfaces, such as Corliss valves, work well with highly superheated steam. The large castings warp so that contact surfaces do not remain true and the lack of moisture which acts as a seal with saturated steam leads to excessive leakage. Difficulty has also been experienced with the lubrication of these sliding types of valves when using highly superheated steam. 206 STEAM POWER An old form of valve known as the poppet value has therefore been adopted by some builders as a solution of the difficulties met in the use of highly superheated steam. This form of valve in four-valve arrange- ment, combined with designs in which short ports and sym- metrical cylinder castings are used, yields very econom- ical engines which can be used safely with a degree of superheat prohibi- tively high in the case of the sliding and oscillating forms of valves. Fig, 136&. — Cross-section, Lentz Engine. HIGH-EFFICIENCY ENGINES 207 Sections of a modern type of poppet valve engine are shown in Figs. 136 (a) and 136 (6), and details of the admission valve and its operating mechanism are given in Fig. 137 (a) and (6). The valves are all double-seated (double-ported or double-beat), that is, they seat at both ends and are made hollow so that the steam passes both around the outside of the valve and through the valve as shown by the arrows in Fig. 137 (6). This results in large area for passage of steam and in quick opening and To Cylinder Fig. 137a. — Admission Valve and Operating Mechanism, Lentz Engine. Fig. 1376. closing, as in the case of gridiron valves, with small actual movement of the valve. The valves are opened positively by eccentrics opera- ting through cams and rollers as shown in Fig. 136 (6) and they are closed by springs as rapidly as the return motion of the cam permits. The eccentrics are mounted on a horizontal lay shaft which is located to one side of the engine, with its axis parallel to that of the latter, and which is driven by bevel gears from the crank shaft of the engine. Since this valve arrangement gives short steam and exhaust poits, permits the use of small clearance, and 208 STEAM POWER gives fairly rapid opening and closing of valves with little throttling when open, it gives good economy when used with saturated steam. By adding superheat the economy is still further improved. The water rate of one of these engines is shown for one load in Fig. 134 (a). A simple, Lentz non-condensing engine is reported to have given a consumption of 16.13 lbs. of steam per horse-power hour with 92.7° superheat, and a pressure of 133 lbs., and this figure is materially lowered by compounding, higher super- heat, lower back pressure, etc. 93. The Una-flow Engine. A very interesting modifica- tion of the steam engine, known as the Una-flow Engine, has been developed recently. The purpose of the design is to permit a great ratio of expansion in one cylinder while at the same time reducing the losses caused by initial condensation when such ratios of expansion are attempted in single cylinder engines of earlier types. Experience with the earlier types has shown that as cut- off is made earlier (ratio of expansion increased) in single cylinder engines, the steam economy is improved until best results are obtained with cut-off in the neighborhood of one- quarter stroke. With earlier cut-off the loss due to initial condensation overbalances the gain which should result from greater ratio of expansion so that a net loss ensues. The construction of the Una-flow Engine Cylinder is shown diagrammatically in Fig. 138. Steam enters the cylinder head at A and jackets the entire cylinder end as it flows toward the inlet valve located in the upper part of the head. The inlet valve is a double-seated poppet valve, the seats being indicated at B and C. Exhaust occurs through openings in the cylinder wall at the middle of its length, the engine piston serving as a valve to cover and uncover these openings. In operation, steam is admitted through the inlet valve as in other engines up to the point of cut-off. After cut-off this steam expands until the piston begins to uncover the HIGH-EFFICIENCY ENGINES 209 exhaust ports, thus effecting release. After completely uncovering these ports the piston returns and covers them again, the instant of complete closure corresponding to Fig. 138. — Section of Una-flow Engine Cylinder. beginning of compression in the ordinary case. Further motion of the piston compresses the trapped steam into the clearance space, and brings conditions back to those existing when admission starts. With an ideal arrangement the pressure in the clearance space at the end of compression would be just equal to that of the high-pressure steam out- side the inlet valve. It will be observed that the piston is so long that at either end of its stroke it fills all that volume of the cylinder between the clearance and the edges of the exhaust ports. The result is that the only part of the cylinder wall which is used in common for the cycles occurring at opposite ends of the cylinder is that part containing the exhaust openings or ports. It will also be observed that this part is strictly a 210 STEAM POWER low temperature zone, coming in contact only with steam at exhaust temperature and pressure. This construction is almost equivalent to the placing of two single-acting cylinders end to end. It isolates each cylinder end in a sense, so that the thermal conditions resulting from the cycle carried through on one side of the piston are practically independent of those resulting from the cycle on the other side of the piston. In the ordinary form of double-acting engine the cylinder walls are common to both cycles for almost their entire extent with correspond- ing interrelation of thermal phenomena. Another important feature is the fact that high pressure, high temperature steam enters at the cylinder head while low pressure, low temperature steam resulting from expan- sion flows out of the cylinder at the point most distant from the point of entry. Steam flow is thus continuously in one direction and low-pressure steam does not sweep over parts which at the next admission will be bathed with high tem- perature steam. As a result of this construction, of the jacketing of the cylinder head with live steam on its way to the cylinder and of the long compression giving a compres- sion curve much nearer to the expansion curve than is com- monly obtained, the temperature distribution along the length of the cylinder is much more perfectly controlled than in the ordinary case. The ends of the cylinder tend to take and retain a temperature corresponding to that of the steam supply. The middle of length of the cylinder tends to take and retain a temperature corresponding to that of the exhaust steam. The length of wall between cylinder ends and center tends to assume temperatures grading from high temperature at the ends toward low temperature at the exhaust ports. As expansion of saturated steam occurs within a cylinder condensation occurs, work being done at the expense of the latent heat of vaporization thus released. This condensa- tion is distributed throughout the entire mass of steam and HIGH-EFFICIENCY ENGINES 211 at the metallic walls the water has a tendency to deposit. It should be noted that the temperature at which such water is formed must correspond at any instant to the temperature of the steam, and as the temperature of the steam decreases continuously as the pressure drops during expansion, the same thing must be true of the water formed during expan- sion. In the Una-flow design the cylinder head and walls near that head are kept hot by the steam jacket and if any water does deposit on them it is probably evaporated into steam R Steam pressure at the Throttle 150 lbs. Vacuum 27" p fc*18 W *16 o A j§14 r 6 Cylinder Diameter 12?4 ; Stroke 20; Speed 200 R.P.M. ■*•»«.. on- -conde n sing-30J ifF. i — "*"■ - r-j. Corn lensin £ 120 F-_ ( ron Jensin K 300° F- ■ Norm. Load 50 60 80 90 100 110 120 Indicated Horse Power 130 140 150 Fig. 139.— Results of Tests Made with a " Una-flow Engine. again. The engine is assumed to operate in this way and it appears to be true that the greater part of the water resulting from expansion concentrates near the exhaust ports and is swept out during the release of steam. Such action is indicated diagrammatically in Fig. 138 by the gradation of the stippling within the cylinder. At first sight it might appear that heat transferred from jacket to steam within the cylinder would surely result in loss but it should be noted that a large part of the steam within the cylinder which is in contact with steam-heated surfaces probably remains in the cylinder and constitutes the clearance steam so that heat transferred to it during 212 STEAM POWER expansion does not pass out through the exhaust valve to such an extent as it would in anengineoftfaeordinarytype. The long compression stroke, with increasing tem- perature and pressure, also has a tendency to evaporate any moisture remaining on the walls. This is certainly true for that part of the metal near the cylinder head. The entering steam thus comes in contact with hot, dry walls and with hot, dry steam and initial condensa- tion is greatly reduced, if not eliminated. Remarkably low steam consumptions have been ob- tained with Una-flow en- gines, even when simple en- gines have been used to expand steam from compar- atively high pressure to a high vacuum. Results ob- tained with one engine under different conditions are shown in Fig. 139. The mechanical con- struction of an 18 X 20-inch single cylinder Una -flow en- gine is shown in Fig. 140, a section through the exhaust ports being shown to the right of the figure. HIGH-EFFICIENCY ENGINES 213 94. The Locomobile Type. In the effort to improve the ecomony of small steam plants the Germans developed a form of plant now known as the Locomobile Type. The name came from the fact that these plants, as originally made, were mounted on wheels and intended for portable use by agriculturists and contractors. Their economy in the use of fuel proved so great that they have since been built for stationary use in sizes running well toward 1000 horse-power per unit. A locomobile of American construction known as the Buckeye-mobile is illustrated in Fig. 141, which shows a longitudinal section of the plant. The tandem compound engine is mounted on top of an internally fired boiler with the engine cylinders located in the flues which lead the products of combustion away from the boiler. The steam generated in the boiler is passed through a superheater suspended in the smoke box. The flow of steam is from the rear toward the front of this superheater (counter flow) so that the hottest steam comes in contact with the hottest gas. The steam then passes through a pipe contained within the flue to the high-pressure cylinder, which is jacketed by the hot flue gases and in which the loss of heat to metal is thus minimized. From the high- pressure cylinder the steam passes to a receiver contained in the smoke box, the receiver serving as a reheater to evaporate any condensate exhausted from the first cylinder and to superheat the steam admitted to the low-pressure cylinder. From the low-pressure cylinder, the steam passes through a feed-water heater in which it raises the tem- perature of the boiler feed and then it passes to atmosphere or to a condenser. Boiler-feed pump and condenser pump, if used, are also integral parts of the plant, being driven directly from the main engine. It' will be observed that every precaution is taken to guard against initial condensation, and to minimize loss of heat in flue gases and in exhaust steam leaving the 214 STEAM POWER HIGH-EFFICIENCY ENGINES 215 plant. The high economies achieved are due to such facts alone. Small plants of this type have given an indicated horse- power hour on a little over one pound of coal when operated condensing, whereas the best large compound reciprocating engine plants seldom do better than about 1.75 lbs. of coal per I.h.p. and often use 2 or more pounds when operated condensing. CHAPTER XII REGULATION 95. Kinds of Regulation. There are two distinctively different kinds of regulation referred to in connection with reciprocating steam engines, one of which may be called fly-wheel-regulation and the other governor-regula- tion or governing. The regulating effect of the fly-wheel has already been referred to. The turning effort exerted at the crank pin by the action of steam on the piston or pistons of an engine is not constant, and the angular velocity of the engine shaft is therefore constantly varying during each revolution. It is the function of the fly-wheel to damp these variations so that they do not exceed the allowable maximum for any given set of operating conditions. The efficiency of the fly-wheel in this respect is measured by the coefficient of fly-wheel regulation 8 W which is defined by the equation -. 'max ' mln /aa\ 0w = y , \pO) in which Fmax = maximum velocity attained by a point on fly-wheel rim or other revolving part; Fmin = minimum velocity of the same point, and F = mean velocity of the same point = - approximately. Governor-regulation is absolutely different. Its function is to proportion the power made available to the instan- taneous demand. The fly-wheel takes care of variations 216 REGULATION 217 occurring during the progress of one cycle, while governor regulation varies the work value of successive cycles. 96. Governor Regulation. If the effect of engine friction be neglected, the power delivered at the shaft of the engine will vary directly with the indicated horse- power. Such an assumption is accurate enough for the discussion which follows. The indicated horse-power of a given engine is deter- mined entirely by the value of the mean effective pressure and the number of cycles produced in a given time, since these are the only variables in the formula for indicated horse-power. The power made available by an engine might therefore be varied by varying the mean effective Fig. 142.— Throttling Governing. Fig. 143.— Cut-off Governing. pressure, or by varying the number of cycles produced in a given time, or by a combination of both processes. All of these possibilities are used. In ordinary station- ary power plants the mean effective pressure is generally varied. In the case of pumping engines, working against a constant head, but required to deliver different quantities of water at different times, the number of cycles per minute is generally altered by changing the speed at which the engine operates. In locomotive and hoisting practice both the number of cycles per minute (speed) and the mean effective pressure are varied as required to meet the instantaneous demands. These variations may be effected manually as by the driver of a locomotive, in which case the engine may be said to be manually governed. Or, they may be brought 218 STEAM POWER about mechanically, as in the case of most stationary power- plant engines, in which case the engine may be said to be mechanically governed. In some instances a combination of manual and mechanical governing is used. 97. Methods of Varying Mean Effective Pressure. The mean effective pressure increases and decreases with the area of an indicator diagram of constant length, so that the mean effective pressure can be changed by any method which will change the area of the diagram. Two methods are in use and they are illustrated in Figs. 142 and 143. The first causes a variation in area by changing the value of the initial pressure. This is generally done by chang- ing the opening of a valve in the steam line just outside of the steam chest. It is called throttling governing, and the valve is called a throttling or throttle valve. The latter name is also commonly used for the valve located near the engine, which is used to shut off the supply of steam entirely when the engine is not in operation. The second method, illustrated in Fig. 143, is known as cut-off governing. The variation of cut-off determines the amount of steam admitted to the cylinder per cycle and is used to measure out the quantity required for the load which happens to exist at any instant. Cut-off governing is used on most modern stationary engines and is exclusively used in large reciprocating engine power plants. 98. Constant Speed Governing. Most engines used for such purposes as the operation of mills and the driving of electrical and centrifugal machinery are required to run at practically constant speed irrespective of the load. They are furnished with mechanical governors which so regulate the power made available that there shall never be any appreciable excess or deficiency which would respectively cause an increase or a decrease in speed. These mechanical devices always contain some sort of tachometer which moves whenever the speed of the engine exceeds or falls below the proper value. The tachometer REGULATION 219 is so connected to the valve gear that it decreases the power-making ability of the engine whenever the speed starts to increase and it increases the power-making ability if the speed drops. Since the valve gear must have a different position for each load in order that it may throttle or cut off as necessary to suit that load, it follows that the tachometer which controls the position of the valve gear must also have different positions for different loads. But tachom- eters assume positions dependent on speed, and therefore different loads can only be obtained if the tachometer and the engine to which it is connected operate at different speeds for different loads. Constant-speed governing is therefore an anomaly. The device which is supposed to maintain constant speed irrespective of load must be operated at different speeds, as the load varies, in order that it may maintain the valve gear in the different positions required to handle the differ- ent loads. All so-called constant-speed engines have their highest speed when carrying no load, and the speed gradually decreases to a minimum as the load increases to a maxi- mum. The total variation is generally between 2 and 4%. The efficiency of a governor in this respect is measured by means of the coefficient of governor regulation, 8 G , which is defined by the equation n 2 — ni , 8g= , (67) n in which ri2 = highest rotative speed attained by the engine; n\ = lowest rotative speed attained by the engine, and n = mean speed n 2 +ni . . = — — approximately. 99. Governors. The mechanical devices which are used for controlling the power-making ability of an engine 220 STEAM POWER as described above are known as governors. There are many varieties and only a few of the more prominent can be described. (a) The Pendulum Governor. One of the earliest forms of governor used on steam engines is illustrated in Fig. 144. It is often called a fly-ball governor. This governor is driven by gearing, chain or belt from the engine, and the weights assume some definite position for each different speed, thus drawing the collar to different positions. The valve gear is connected to this collar and is moved correspondingly. A similar governor is shown in Fig. 131, which also indicates the way in which the collar is connected to the valve gear in the Corliss type of engine. The governor rods are moved as the collar moves and they in turn alter the position of the knock-off cam, and thus vary the time at which cut-off occurs. As the speed increases due to a decrease of load, the governor weights and collar move up, and this shifts the cams so as to produce earlier cut-off and decrease power-making ability. (6) Shaft Governors. On medium- and high-speed engines fitted with some form of slide valve it is found best to use what are known as shaft governors. They are gen- erally carried within the fly-wheel of the engine, operate in a plane passing through the rim of the wheel at right angles to the shaft, and operate upon the eccentric in such a way as to vary the cut-off with speed (and load) changes. Fig. 144. REGULATION 221 Fig. 145. One simple form of such a governor is shown in Fig. 145. The eccentric is not mounted directly upon the engine shaft, but is carried by a pin P in the fly-wheel and is slotted so that it can swing back and forth across the shaft, about P as a center. Its position at any time is determined by the position of the governor weights W, which draw the eccentric down (in the figure) as they move out. The center of the eccentric is indicated by a heavy dot in the figure, and it will be seen that this center would travel in the arc of a circle about P, as the weights moved. If the path of the eccentric center is drawn on a Bilgram diagram, it will be found that this motion is equivalent to decreasing the length of the eccentric crank and increasing the angle of advance, resulting in earlier cut-off as the weights move out with increasing speed and decreasing load. Other events will also be changed as the eccentric swings, and some of these changes are occasionally unde- sirable. Numerous designs have been developed in which the eccentric is so guided as to produce various sorts of rela- tions between the different steam and exhaust events. All can be divided into two classes, those in which the eccentric swings about a fixed center variously located, and those in which the center of the eccentric is guided to move in a straight line. All can be studied by plotting the path of the eccentric center (path of Q) on the Bilgram diagram. The Rites Inertia Governor is a form of shaft governor so designed as to act very quickly with change of speed, 222 STEAM POWER and to be very powerful, so that it can shift heavy parts. It is shown in place in the wheel in Fig. 146. With changes in speed it acts like a governor of the type just described, swinging (with increasing speed) about a fixed point P in the wheel as its center of gravity G moves outward under the action of the centrifugal effect C and against the action of the spring. This motion shifts the center of the Fig. 146. eccentric from E toward e, giving the desired variation in cut-off. Superposed upon this action is that of inertia. Assume the wheel and governor to be rotating clockwise at a given constant speed. If the engine speed is suddenly increased, the wheel will move faster, but the governor bar will tend to continue rotating at the same speed because of its inertia. It will thus lag behind the wheel, rotating about P and bring- ing about an earlier cut-off. The position thus assumed will later be maintained by centrifugal effect if the new speed REGULATION 223 is maintained. The particular advantage resulting from using inertia in this way is speed of action. In many forms of governor the inertia of the moving parts actually resists the efforts of the governor to assume the new position required by changed load and speed, whereas in this form the inertia of the governor parts is used to increase the speed with which they move to the new position. CHAPTER XIII THE STEAM TURBINE 100. The Impulse Turbine. One of the oldest of modern water wheels is the tangential or impulse wheel shown diagrammatically in Fig. 147. Water flowing from a reservoir above the wheel passes through a nozzle and the Fig. 147. — Tangential or Impulse Wheel. jet, moving at high velocity, strikes buckets on the rim of the wheel and causes the latter to revolve. Theoretically the velocity of the water in the jet would be v = V2gh feet per second, .... (68) in which g = gravitational constant, 32.2, and h = head in feet as shown in the figure. The kinetic energy possessed by the moving water would be K = w v* 7"2' 224 (69) THE STEAM TURBINE 225 in which iv represents pounds of water discharged per second and g and v have the same meanings as above. If the buckets of the wheel could reduce the velocity of the water to zero they would absorb all of this kinetic energy and (assuming no losses within the buckets and the bearings of the wheel) would make all of it available at the shaft for the doing of useful work. Any fluid moving at velocity v and striking buckets in the form of a jet would possess kinetic energy in quantity given by Eq. (69) and would drive the wheel in the same way. Steam might therefore be used instead of water with exactly the same results, and steam is so used in what are known as impulse steam turbines. Experience shows that steam will flow at high velocity from any opening made in the s^> steam space of a boiler or from any open-ended pipe con- nected to such a boiler. This is commonly said to be due — ^ _j | to the high pressure within C^M^,,^^ gJL, the boiler, the spectator pic- turing the process as the driving out of part of the steam by the high-pressure steam within the boiler, just as though the part leaving were a solid piston and were driven vS^^^Pf ^ Fig. 148. Fig. 149. out as is the piston of an engine during admission, as shown in Fig. 148. An hydraulic analogy is given in Fig. 149. The vessel shown is supposed to be fitted with a piston, and it is assumed to be possible to exert any desired pressure upon the piston. Any such pressure exerted is the exact equivalent of some given head of water and the resultant jet velocity would be given by Eq. (68) by substituting for h the head in feet equivalent to the pressure exerted upon the piston. When an " elastic " fluid such as steam is being con- 226 STEAM POWER sidered it is, however, necessary to take account of other factors. The steam within the boiler exists at a high pressure ; after issuing it exists in the atmosphere at a lower pressure. But low-pressure steam contains less heat than does steam at high pressure, and this difference must exist in some form, as it is energy and could not possibly have been destroyed during the flow. Experiment shows that steam after flowing into the atmosphere from a boiler in this way has exactly the same characteristics as though it had expanded adiabatically behind a piston through the same temperature range, ex- cepting for the fact that it has a very high velocity, which it would not possess if expanded behind a piston. Experi- ment further shows that, if small losses be neglected, the kinetic energy possessed by a jet of steam is exactly equal to the energy which would be turned into work if that steam acted on a piston as in an ordinary engine. A complete picture of the process of flow can then be made by assuming the steam flowing out in the form of a piston driven by high-pressure steam, as before, and adding to this the idea that this piston expands adiabatically as it travels from the region of high to that of low pressure. This expansion liberates heat contained within the piston or plug of steam and this heat is used in imparting addi- tional velocity to the moving steam which is giving up this heat. The result of using such a jet upon a theoretically perfect tangential or impulse wheel would be to rob the jet of all this energy. But the energy possessed per pound of steam in the jet is just the same as that shown under the upper lines of a complete expansion cycle using one pound of steam. The area under the upper horizontal line of the PF-diagram of the cycle as shown in Fig. 21 may be assumed to represent the work done upon one pound of steam (flow- ing out) by another pound which is being evaporated and pushing out the first in order to make room for itself. The THE STEAM TURBINE 227 area under the expansion curve in the PF-diagram repre- sents the energy converted into velocity energy by the adia- batic expansion of the flowing steam. The lower hori- zontal line represents the negative work during condensa- tion to water at the lowest pressure and temperature, and the left-hand line represents the pumping of this water back into the boiler and the raising of its temperature to the value tDiarihragm Fig. 150. — Early Form of Impulse Turbine. maintained within the boiler. The complete expansion cycle is therefore the cycle upon which the impulse steam turbine operates and, as a matter of fact, it is the theoretical cycle of all steam turbines. The ideal impulse turbine would therefore be acted upon by a jet which possessed available kinetic energy represented by the area of the complete expansion cycle. If the buckets could entirely remove this energy, that is, could reduce the velocity of the jet to zero, the same amount 228 STEAM POWEK of energy could theoretically be made available at the shaft of the turbine. An example of a simple form of impulse steam turbine is given in Fig. 150, in which the essential parts of an early form of Kerr turbine are shown. The wheel, the diaphragm and nozzles are all inclosed within a casing. The space on one side of the diaphragm is connected to the steam pipe and that on the other is in communication with the space into which the exhaust steam is to be exhausted. Another form of impulse turbine is shown in Fig. 158. It will be described later. 101. Theoretical Cycle of Steam Turbine. It was shown in the preceding section that the steam turbine operates on the complete expansion cycle. If a turbine could remove from the steam passing through it and convert into mechan- ical form all the energy which it is theoretically possible to convert it would therefore make available mechanical energy represented by the area of the PF-diagram of the complete expansion cycle. The area of the corresponding T »-"' - a 3 O pE< 0.5 lb. pressure. This is practically at 540° F. absolute, or about 80° F. THE STEAM TURBINE 231 Second, take from the steam table the volumes of one pound of steam at, say, 200 lbs., 140 lbs., and 100 lbs. absolute pressure when superheated to the values shown by this vertical line. These will be about 2.75 cu.ft., 3.58 cu.ft., and 4.67 cu.ft., respectively. Plot these volumes with corresponding pressures on a PF-chart as shown in Fig. 152. Third, take from the TV-chart the pressures at which the vertical line intersects different volume lines in the wet steam region and plot volumes against pressures on the PF-chart. Fourth, draw a smooth curve, as shown, through all points so determined. Fifth, draw horizontal top and bottom lines and a vertical line at the left of the diagram. This vertical line should be to the right of the pressure axis by an amount representing the volume of one pound of water, but the volume is so small that it cannot be plotted to any ordinary scale. 102. Nozzle Design. It was stated In preceding sec- tions that the energy which would be converted into work by the introduction and adiabatic expansion of steam behind a piston is converted into kinetic energy when steam flows out of an orifice or nozzle and that an ideal impulse turbine could absorb all this kinetic energy from the jet, bringing it to rest and making the energy available in the, form of useful power at its shaft. It is, therefore, of interest to determine the velocity which a jet will acquire under different conditions. This could be done by evaluating the area of a diagram, such as that of Fig. 152, and then putting this value m place of K in Eq. (69) and solving for v, but it can be done much more accurately and expeditiously in other ways. The heat energy which can be converted into kinetic energy of the moving jet and which can later be con- verted into useful work by the turbine wheel is represented by the area enclosed within the lines of the complete expansion cycle when drawn on the T^-diagrams. That is the area abed in Fig. 153, for instance, for the case of wet steam at the beginning of expansion. But this area is equal to that representing the heat supplied minus 232 STEAM POWER that representing the heat rejected, that is, Q1-Q2, so that K(inB.t.u.)=Qi-Q2. .... (70) The values of Qi and Q 2 can be found very readily by plotting the points c and d upon a T-chart for steam and observing the constant heat lines upon which they fall, a b 1 ' c d \ / i \ > ■2 1 \ ' Fig. 153. or they can be obtained even more conveniently from what is known as a Mollier Chart for steam. In this chart, entropy above 32° F. is plotted against heat above 32° F. as shown in Fig. 154. An adiabatic expansion on this chart is shown by a horizontal line, since this shows a constant entropy change just as a vertical line on the Tcf> chart shows a constant entropy change. If a point is found in this chart giving conditions corre- sponding to those at point c in Fig. 153, the value of Q x THE STEAM TURBINE 233 «2 Mo.T}ua »> 234 STEAM POWER can be read directly under that point on the horizontal axis. A horizontal line drawn from that point to the terminal-pressure line will give the point corresponding to d of Fig. 154 and the value of Q2 can be read on the hori- zontal axis immediately below that point. The difference between the two readings gives the value of the kinetic energy K or of the mechanical energy which an ideal tur- bine could make available, but the expression will be in British thermal units and not in foot-pounds. This value of the kinetic energy, i.e., K = Q\ — Q2, may then be placed in Eq. (69), giving, X = 778(Qi-Q 2 )=£ft.-lbs., . . . (71) since Qi and Q2 refer to one pound of steam, or K = 77MQ-Q 2 )=7rft.-lbs., • • • (72) when w represents the number of pounds of steam flowing per second. Solving cither Eq. (71) or Eq. (72) for v gives, v = V77SX2g(Q 1 -Q 2 ) = V / 778X64.4(Q 1 -Q 2 ) = 22WQi-Q 2 feet per second. . . (73) The design of a nozzle consists simply in choosing such sections that the desired amount of steam may flow through it with the desired pressure drop, as the velocity obviously is determined by that pressure drop. This is very con- veniently done by working in terms of one pound of steam, since all formulas and charts are generally given on that basis, and then multiplying the cross-sectional areas found by the number of pounds of steam required. Assume for instance that it is desired to design a nozzle THE STEAM TURBINE 235 to pass one pound of steam per second with an initial pres- sure of 100 lbs. per square inch abs., and a terminal pressure of 60 lbs., the steam being initially dry and satu- rated. The Mollier chart shows that Qi is equal to about 1187 B.t.u. per pound of steam, while Q2 is equal to about 1147 B.t.u. The velocity with which a jet would issue from a theoretically perfect nozzle under these conditions may then be found by using Eq. (73). This gives v = 22Wl 187 -1147 = 1416 feet per second. The shape of the entrance end of the nozzle is generally made such that the steam will enter it without great dis- turbance and the shape beyond that point is determined by methods which will be explained below. The cross- section of the discharge end must be such as to pass the required quantity at the velocity found above to be equal to 1416 feet per second. This is easily done by deter- mining the volume of steam discharged. Drawing the adiabatic expansion on the T<£-chart will give the quality at the end of the expansion; or, the quality can be determined by finding what quality a pound of steam at 60 lbs. pressure must have to give it a heat content of 1147 as found above. With the quality known the ter- minal volume per pound can be found by multiplying the quality by the specific volume at terminal conditions. Thus for the case under discussion the quality will be about 96.7% and as the specific volume at 60 lbs. is 7.17 cubic feet, the volume to be passed per second, per pound of steam is 0.967X7.17 = 6.94 cu. ft approximately. If the velocity is 1416 feet per second the area per pound of steam must be 6.94-^-1416 = 0.0049 sq.ft. The exact shape of the nozzle is determined by deciding upon the way in which pressure, or velocity, or volume 236 STEAM POWER shall change as the steam passes through' it. Suppose, for instance, that a nozzle is to be constructed of the length shown by ab in Fig. 155, and that the pressure is to vary along its length as shown. Assume also that the nozzle is to pass 10 lbs. of steam per second. Taking initial pressure as 100 lbs. and terminal as' 60 lbs., the conditions a ^100 90 80 70 2 60 *^«^ f> s / i, 7 ?2 \t> b x3 * / /N£ ^ tfj, / <^v? / fVa riati 3ii o: Vel< >city 1400 1200 1000 800 600 400 200 Length of Nozzle Fig. 155. — Nozzle Design. will be the same as in the problem above. The discharge area will have to be 10X0.0049 sq.ft. or 0.049 sq.ft. The area at the plane X2 must be that required to pass the steam when it has the velocity resulting from expansion from 100 down to 64 lbs., just as though the nozzle ended at that point. This can be found just as the terminal area was found above. Similarly the sections at x\ and x can be found by figuring velocity and area for expansions to THE STEAM TURBINE 237 74 and 90 lbs., respectively. If the various areas required are determined in this way, the nozzle will have a longi- tudinal section about as shown by the dotted lines in the figure and the variation of velocity will be about as shown by the curve. If the shape of a nozzle is determined in the same way for a case in which the terminal pressure is less than about 0.58 of the initial pressure, the nozzle will be found to have a very different shape. This is shown in Fig. 156. The nozzle is known as an expanding nozzle and the smallest section is known as the neck. The pressure P» in the neck is always equal to about 0.58 Pi and the velocity in the neck is always equal to just over 1400 feet per second. It is therefore the section at the neck which determines the quantity of steam which a nozzle will discharge if expanding to a pressure equal to or lower than 0.58 Pi. 103. Action of Steam on Impulse Blades. It has been stated that the steam acting in an impulse type of turbine delivers energy to the wheel of the turbine by giving up its kinetic energy. In an ideal turbine the steam jet would be brought to rest and would thus give up all of its kinetic energy. In real turbines it is impossible to bring the jet to rest, as practical design problems prevent it. There is there- fore always a loss in real machines because of the residual or terminal velocity of the steam as it leaves the wheel. Thus let the black section in Fig. 157 represent the section of a bucket or blade sticking out radially from the rim of a Fig. 156. — Expanding Nozzle. 238 STEAM POWER wheel, the wheel revolving about the axis indicated by the dot dash line but located behind the plane of the paper, see Fig. 158. If minimum loss by eddying is to be experi- enced at the point at which the steam jet enters the blade, the jet must enter the blade along a tangent to the curve of the in- side at the entrance edge. This direction is shown by the line marked v r in the figure. Were the bucket stationary, the steam jet would move as shown hyv r , but as the bucket moves ahead, and, so to speak, runs away from the jet, the steam must really travel in a direction such as that indicated by v a in order to strike the bucket in the direction indicated by v r . The conditions governing the flow of steam into a bucket are the same as those governing the speed with which and direction in which an individual runs toward and jumps upon a moving vehicle. He will experience least shock when he is moving ahead at the same rate as the vehicle at the instant when he gets on board. His motion must therefore be made up of two, one toward the vehicle and the other in the direction of the vehicle's travel. In the case of steam flowing onto a blade as shown in Fig. 157, the various velocities are so related that when drawn to scale the real or absolute velocity of the steam, v a , and the real or absolute velocity of the blade, v b , form two sides of a triangle of which the closing side represents v r , the velocity of the steam relative to the bucket. The value and direction of v T is obviously obtained from v a by geometrically subtracting the velocity of the bucket. Fig. 157. THE STEAM TURBINE 239 After entrance, the steam flows around the inner curve of the blade and is finally discharged with the same rela- tive velocity as that with which it entered, and at an angle set by the tangent to the inner curvature of the discharge edge of the blade as shown by v R . But, since the steam has been moving ahead with the same velocity as the bucket during the entire time that it was in contact with the bucket, it is also moving ahead with a velocity Vb when it leaves the wheel. Its real or absolute velocity is then v A , which is found by combining v R and v b as shown in the figure. The kinetic energy possessed by the jet when entering wv the blade is equal to ~~ ft.-lbs., and that which it possesses WVa when leaving is — = — ■. Obviously, the energy removed VOV it VOV a while passing over the blade is -~ — •. If the blade were theoretically perfect, it would be so constructed that v A 2 would be zero and all of the kinetic energy would then be removed. This is practically impossible in a real mechan- ism, and there is always a loss due to the residual velocity v A . The best that can be done is to so choose the angle of jet and blade, and the velocity of blade with respect to the steam, that the actual numerical value of v A is made as small as possible. Designs usually work out in such a way that this occurs when the blade velocity is equal to about 0.47 of the abso- lute velocity of the steam jet. 104. De Laval Impulse Turbine. The expanding nozzle already described was first used by De Laval in an impulse type of turbine. The essential elements of this device are shown in Fig. 158. The nozzles are arranged at such an angle to the plane of the wheel that the steam jets strike radially arranged blades at the proper angle to enter without much loss. The blades deflect the jets as shown and 240 STEAM POWER absorb the greater part of their kinetic energy, so that the steam flows away from the wheel with low absolute velocity. As many nozzles are used as are required to make avail- steam In Nozzle Turbine Shaft . tNozzles- Fig. 158. — Single Stage, De Laval Impulse Turbine. able the amount of energy desired at full load, and pro- vision is made for shutting off one or more nozzles by hand when conditions do not warrant the use of all. Governing for ordinary variations of load is effected by throttling the steam flowing to the nozzles in use, thus altering the initial pressure as necessary. THE STEAM TUEBINE 241 A section through the wheel and casing of such a tur- bine " direct connected " to a centrifugal pump is given in Fig. 159. The steam flows into the live steam space through a throttle valve controlled by the governor; the valve and connections are not shown in the illustration. From the live steam space the steam flows through nozzles not shown, and into the exhaust steam space, thus acquir- ing a high velocity. The buckets of the wheel are located just in front of the discharge ends of the nozzles and the steam moving at high velocity must pass through them before moving on toward the exhaust outlet. 105. Gearing and Staging. It has been stated that the most efficient operation with ordinary designs is obtained when the blade speed is equal to about 0.47 of the absolute steam velocity or, roughly, half the velocity of the imping- ing jet. To get high economy in the use of steam, large pressure drops are used and very high jet velocities result. When the buckets of a turbine are operated at peripheral speeds equal to half these jet velocities one of two diffi- culties is often met. The stresses induced in the wheel structure by centrifugal effects become so high as to offer serious difficulties in design, or the rotative speed of the unit becomes too high for direct connection to the machine which is to be driven. One method of partly overcoming the latter difficulty is to operate the turbine at or near the theoretically desir- able speed and transmit the power to the driven machine through gears which decrease the rotative speed to the necessary extent. This method was used with all of the early De Laval turbines which were of comparatively small capacity. It is now being successfully applied to marine propulsion and other purposes for which large units are used. It is only a partial remedy in the case of large units, however, as the gears necessary for the desired reduction and the size of the turbine wheels would both become excessive. 242 STEAM POWER 3 H m 'bJO THE STEAM TURBINE 243 Another and very common method is known as com- pounding or staging, . This may be of two varieties. The pressure drop in each stage may be limited to that which will give a reasonable velocity and a number of such stages may be put together in series on one shaft. This would give -one set of nozzles and a wheel for each stage, the steam discharged from one wheel with very low velocity expanding to a lower pressure through the nozzles of the next stage and impinging upon the wheel of that stage with the resultant high velocity. Such an arrangement is known as pressure staging or pressure compounding, and is extensively used in large turbines. The pressure staging method is illustrated in Fig. 160 as applied to the De Laval type of impulse turbine. The combined increase in diameter of wheels and increase in length of blades gives the necessary increase in area to pass the larger volumes of steam as the drop of pressure continues from stage to stage. Instead of staging on a pressure basis, staging on a veloc- ity basis may be used. In such a case the drop in pressure through one set of nozzles is great and the resultant veloc- ity high. The steam moving at this high velocity is then directed upon the buckets moving at such peripheral velocity that they absorb only part of the kinetic energy of the steam, discharging it with a lower absolute velocity than that with which it entered, but one which is too high to be thrown away. The steam then passes through a set of stationary vanes which direct it upon the blades of a second wheel, in passing through which it gives up still more of its kinetic energy with a corresponding further decrease of velocity. If the velocity still possessed by the steam warrants it, a second set of stationary guide vanes and a third set of moving buckets can be supplied for further reducing it and by carrying this velocity staging through a sufficiently great number of stages any initial velocity 244 STEAM POWER could be absorbed theoretically without the use of wheels with high peripheral speeds. Practically, losses due to friction, eddying and other sources limit the number of velocity stages to two or three. THE STEAM TURBINE 245 ■cdi Second Wheel i ^ o Fig. 161.— Early Form of Curtis Turbine. 246 STEAM POWER Velocity staging is combined with pressure staging in the Curtis type of turbine. A section through part of an early design of vertical turbine of this type as built by the General Electric Company is shown in Fig. 161. The turbine illustrated had four pressure stages and each pres- sure stage had two velocity stages. Many varieties of impulse turbines have been developed and all of the larger ones employ several wheels and sets of nozzles and diaphragms to obtain the necessary staging. The same result has been obtained in some of the smaller models by discharging the steam from nozzles on to a set of buckets which are able to absorb only a fraction of the kinetic energy, catching it at discharge and returning it for another passage through the buckets, and so on until the greatest practical fraction of the kinetic energy has been absorbed. A vertical section through a large, horizontal turbine of the impulse type is given in Fig. 162. Units of this sort are built with different numbers of stages depending upon both the total pressure drop for which they are designed and the thermal efficiency which is desired. It is obvious that the number of stages required to give a certain thermal efficiency will increase with the total pressure drop (that is, the difference between steam pressures at entrance and exit respectively) if the peripheral speed permitted is to remain the same. Conversely, if other things remain equal, increasing the number of stages increases the thermal effi- ciency up to the point where increasing losses overtake further possible gains. Commercially, the number of stages used in any given case is determined as a sort of compromise between first cost of unit, operating reliability and money value of thermal efficiency. Twenty-two stages are about the upper limit at the present time and the great majority are built with a smaller number. The impulse type is built in all sizes between a unit THE STEAM TURBINE 247 3A{J(XiOUJ9A.OO 248 STEAM POWEE Fig. 163. Elementary Reaction Turbine. capable of developing a few horse-power and a unit capable of developing about 60,000 horse-power. 106„ The Reaction Type. If high-pressure steam or other fluid be forced into a de- vice arranged as shown in Fig. 163 and free to revolve about a vertical axis, the jets blowing out of the nozzles will cause the mecha- nism to revolve in the direction indicated by the arrow. This rota- tion is said to be due to the reaction of the jets, and the mechanism there- fore constitutes a simple form of reaction turbine. By increasing the number of nozzles any amount of steam could be dis- charged and therefore any amount of work could be obtained. ,, This multiplication of nozzles can, however, be more conveniently accomplished by fastening radial vanes to the periphery of a wheel as shown in Fig. 164, the space between any two vanes constituting a nozzle through which the steam can discharge. By mounting such a wheel within a casing as shown in Fig. 165 it forms a simple reaction turbine. One of the characteristic differences between the impulse and the reaction types lies in the distribution of pressures. In the impulse type the nozzles are fastened into a stationary part of the turbine and the drop of pressure occurs entirely within the nozzles. The wheels are therefore immersed in a space in which a uniform lower pressure exists. In the reaction type, on the other hand, the nozzles are carried on the wheel and Fig. 165. THE STEAM TURBINE 249 250 STEAM POWER there must be a higher pressure on one side of the wheel than there is on the other. Since there must also be me- chanical clearance between the blade tips and the interior of the casing, it follows that the reaction type will be handicapped by considerable leakage which does not exist in the impulse type, excepting as some of the jet may " spill " over the ends of the blades in the latter. The difference of pressure on the two sides of the wheel also causes a tendency toward motion of the wheel along the shaft, or of the wheel and shaft, in a direction away from the higher pressure. Many unsuccessful efforts have been made to design efficient reaction turbines, but no pure reaction type has yet been commercialized. The turbines commonly called reaction turbines are really combinations of re- action and impulse types. One example of what is commercially called a reac- tion turbine is shown in Fig. 166. Alternate rings (or rows) of stationary and movable blades guide the steam as it expands from the high pressure at one end to the low pressure at the other. The stationary blades project inward from the interior surface of the stationary casing and the movable blades project out- ward from the external surface of the cylindrical rotor. The rotor blades act like those of an impulse turbine in partly reversing the direction of jets of steam which reach them with comparatively high velocities, but they also act Fig. 167. THE STEAM TURBINE 251 like the movable nozzles of a reaction turbine since the steam in passing through them expands and acquires kinetic energy, giving a reaction on discharge. The sta- tionary blades serve to redirect the steam so that it strikes the next set of moving blades at the proper angle and they also serve as nozzles in which velocity energy is acquired. This is shown diagrammatically in Fig. 167, in which S denotes stationary, and M movable blades. The Parsons type, illustrated in Fig. 166, may be described as a multistage type in which impulse and reaction are utilized in conjunction. The balance pistons shown in the figure are used to balance the end thrust caused by the difference in pressure existing on opposite sides of the wheels in the case of reac- tion turbines. Each piston is of such a diameter that it presents a surface equal to the blade surface acted upon by one of the unbalanced pressures, and by connecting across as shown in the figure a high degree of balance is secured. The overload valve is used to admit high-pressure steam to the low-pressure blades for carrying excessive overloads. The larger area of the passages through these blades per- mits an abnormal amount of high-pressure steam to pass, thus giving a high load-carrying capacity with decreased economy. 107. Combined Types. The clearance at the ends of the stationary and moving blades in the Parsons type of turbine permits considerable steam to leak by, as previously explained. This clearance must have almost the same length (measured from blade tip to opposing metal) in all stages in order to insure freedom from rubbing, but it is more detrimental in the high-pressure stages than in the low. The high-pressure blades are much shorter than the low- pressure blades and a leakage length of a certain amount is therefore equal to a greater fraction of the total blade length. The density of the high-pressure steam is also so 252 STEAM POWER much greater than that of the low-pressure steam that many more pounds can leak through an opening of a given size in a given time. In discussions of this character, it should not be forgotten, however, that leakage area is determined by the dimension already referred to multiplied into a circumference and that the circumference is much greater at the lower end. Because of these and other reasons many manufacturers have come to the conclusion that the impulse type is best for the high-pressure end of the turbine and the reaction type for the low-pressure end. Many such combinations have been produced and they are giving very good results. 108. Steam Consumption of Steam Turbines. It is exceedingly difficult to compare the steam consumption of turbines and reciprocating engines in a general way. Roughly the steam consumption of the better varieties of the two types is of the same order for comparable con- ditions with the advantage probably slightly in favor of reciprocating engines in the smaller sizes and in favor of turbines in the larger sizes. It has been shown that the steam turbine operates on the complete expansion cycle while the reciprocating engine operates on a cycle with incomplete expansion. The tur- bine therefore has a certain theoretical advantage because its cycle is such as to convert into work a greater amount of heat per pound of steam used between given upper and lower pressures. This choice of different cycles rests on a sound founda- tion. This can be appreciated best after studying Fig. 168 which shows the volumes assumed by steam expanded adiabatically from an initial pressure of 150 pounds gauge and 100° F. superheat. It will be observed that at the lower pressures the volume increases very rapidly with a small drop of pressure. If an attempt were made to expand steam in a reciprocating engine down to exhaust pressure the cylinder would have to be increased in size by a very large amount in order to accommodate the rapidly increas- THE STEAM TURBINE 253 ^*-" J s s J 8 £ S • S o 1 s « •§ < a; oi W o •si Ph I ^ ■gqySQ sanoui-aansseaj 254 STEAM POWER ing volume of steam. As previously stated in Section 40, the actual amount of additional energy recovered from the steam would not be worth enough to balance the increased friction loss of the larger parts and the increased invest- ment in the larger engine. Further, some pressure differ- ence is required to cause the steam to flow through the comparatively restricted exhaust ports and valves in the short time available and absolutely complete expansion is thus really impossible even if it were desirable. With the steam turbine, experience has shown that it is often economical to expand the steam down to a low back pressure and as there are no restricted exhaust ports and valves no pressure differential is required. This is of great importance as a very small pressure drop at low pressures makes available a very large amount of energy in com- parison with the result of a corresponding pressure drop at higher pressures. For example, trial on a T(f> or Mollier Chart will show that saturated steam in expanding with constant entropy from 200 pounds absolute to 0.5 pound absolute makes available almost as much energy in dropping from 15 pounds to 0.5 pound as it does in dropping from 200 to 15. Further, such trial will show that at the extremely low pressures a very small pressure drop liberates a relatively tremendous quantity of energy. As a result of these characteristics of steam, the turbines which are built to give low steam consumption are designed to operate with very low back pressures, that is very " high " vacuums. In the larger sizes this introduces real diffi- culties in design. The entire volume of steam at the lowest pressure has to flow through the last set of blades and these must be extremely long or carried on a wheel of large diam- eter, or both, in order to give the necessary area for passage of the steam. Such difficulties have led to numerous " double flow " designs in which the steam is expanded from initial pressure to some lower pressure in the ordinary way THE STEAM TURBINE 255 and is then introduced into a section in which it can divide and flow both ways through two opposed but similar sets of blades. One double-flow arrangement, as applied to an impulse turbine, is shown diagrammatically in Fig. 169 i < fe and another, applied to a reaction turbine, . is shown in Fig. 170. In some of the larger sizes of turbines the unit has been broken up into two or more parts. Each part is a complete turbine but is built for a smaller pressure range than that 256 STEAM POWER THE STEAM TURBINE 257 through which the steam is expanded by the complete unit. The " high pressure unit " receives steam from the boilers and expands it down to some intermediate pressure. It exhausts into one, or two, " low pressure units " which expand the steam down to the lowest pressure and discharge it to the condenser. This arrangement is similar to compounding as applied to reciprocating engines but it has an entirely different purpose. Compounding of turbines in this way offers greater flexibility of design. The designer can choose dif- ferent rotative speeds for the high and low pressure units instead of having to use some sort of compromise as is neces- sary when all elements are carried on a common shaft. Compounding offers another advantage which is of impor- tance in the larger sizes. The size and complexity of each of the parts is less than if the entire turbine were built in a single unit and it is logical to assume that this adds to the operating reliability of extremely large units, if all other features are alike. The very real thermal advantage to be gained by using low back pressures with steam turbines is shown by results of tests which indicate that lowering the back pressure by one inch of mercury will increase the economy by from 3 to 10 per cent, depending upon the type of tur- bine, the back pressure under consideration and upon other factors. Superheat is also very effective in improving the thermal efficiency of the steam turbine. In general, every ten degrees of superheat causes a saving of 1 per cent in the weight of steam required for a given output. 109. Low Pressure Turbines. Experience has shown that reciprocating engines are fully the equal of turbines in the high pressure ranges, in many cases they are even superior. The turbine, on the other hand, has the advan- tage at low pressures and in cases where great ratios of expansion are used. It was at one time suggested that 258 STEAM POWER these characteristics should be recognized by building mixed plants, using reciprocating engines for the first part of the expansion and exhausting these into turbines at or near atmospheric pressure. Under ordinary conditions this is not an economical solution as the investment is so high that any thermal gain which can be obtained is not sufficient to balance the increased capital charges to say nothing of increased com- plications. However, the scheme has been used to advan- tage in increasing the capacity and thermal efficiency of reciprocating plants which were already installed. The ability of the turbine to handle low pressure steam to advantage has given rise to the use of low pressure turbines in many different ways. As examples, the ex- haust of hoisting engines, steam hammers, and other apparatus commonly exhausting at atmospheric pressure is now frequently led to one or more low pressure tur- bines in which it is expanded with the recovery of a large amount of power from what would otherwise be waste steam. When low pressure turbines are used in this way it fre- quently happens that the demand for steam on the part of the turbine and the make of steam on the part of the primary user are so different from instant to instant that some device must be used to store steam between the two. The device used for this purpose is known as a regenerator. It consists of some sort of closed vessel in which steam can be mixed with and condensed in hot water. When the make exceeds the demand of the turbine the pressure and temper- ature within the regenerator rise; when the demand of the turbine exceeds the make the pressure and temperature within the regenerator fall. 110. Steam Turbo-generators. The steam turbine is ideally suited to the driving of electric generators of the alternating current type as the desirable speeds for the two devices fall in the same general range of values. This fact, STEAM TUEBINE 259 coupled with the ease with which such units can be con- structed in large capacities, the comparatively low cost and the high thermal efficiency attainable, has resulted in 3200 J100 82000 I II': STEAM TURBINE PROGRESS Dot ted I Jnes -Con lpou idM ichii es i u / / ffi 28000 / / M 24000 P. CD 1 \^ S>- / / V / 1 1 y Turbir © O O O N> // B.T.U. used b O O O O I ^ SS«T r^ is*** C ,teaij f o 3 03 "3 280 STEAM POWER way down over the water-cooled tubes. The condensate, mixed with gases and vapors, is drawn from the bottom of the shell by the wet-vacuum pump and discharged to the hot well. The condensing water is forced through the tubes of the condenser by means of the reciprocating circulating pump, entering the lower tubes at the right-hand end in the figure, making two " passes " through the condenser and leaving at the top. Because of the path of the water a condenser of this type is sometimes called a two-pass or double-flow condenser. With the arrangement illustrated, the steam which condenses upon the upper tubes falls as a rain from tube to tube until it finally settles at the bottom and is drawn off. The outer surfaces of the lower tubes are therefore practically covered with water and this has two disad- vantages. First, these tubes carry the coolest circulating water and they therefore cool the condensate coming in contact with them while the water flowing through them is unnecessarily heated. Cooling of the condensate means a lower hot-well temperature than would otherwise be obtained, but if the condensate is to be used for boiler feed, the temperature of water in the hot well should be maintained as high as possible, since this water will eventually have to be heated to boiler temperature with a correspond- ing expenditure of heat. Second, tubes which are being used to cool water covering them are of little use as condens- ing surface, and hence such surface in a condenser is com- paratively inactive. The ideal arrangement would carry away the liquid condensate as fast as formed, leaving the tubes first entered by the condensing water to act as the final condensing and cooling surfaces, thus bringing gases and non-condens- ible vapors into contact with the coolest surfaces just before entering the vacuum pump. Numerous designs which approximate this ideal have been developed recently and CONDENSERS AND RELATED APPARATUS 281 they give better results than do the earlier and simpler forms. The improvement is shown by the values of con- densing surface per developed horse-power of engine. In early designs it was customary to supply 2 \ sq. ft. of tube surface or more per horse-power. Some of the most recent installations are giving better vacuums with only 1 sq. ft. per horse-power. One of these condensers passes tne condensate through a set of tubes so located that the engine exhaust strikes them before impinging on any tubes carrying condensing water. This results in a partial condensation of the exhaust and raises the temperature of the condensate within the tubes to very near that of the exhaust, thus heating the boiler feed to a temperature practically corresponding to the exhaust temperature of the engine. Surface condensers are commonly operated with a vacuum of from 24 to 26 ins. of mercury when used with reciprocating engines and with a vacuum of 28 to 29 ins. when receiving the exhaust of steam turbines. When operated at the lower vacuums wet-vacuum pumps are generally used, but the best types of dry-air pumps must be installed in combination with well-designed condensers when the higher vacuums are sought. An installation of a surface condenser and necessary auxiliaries as applied to a steam turbine of moderate size is shown in Fig. 185. The steam exhausted by the turbine enters the top of the condenser shell and spreads out over the tubes. As it is condensed it gravitates to the lower part of the shell and finally flows into the " hot well " attached to the lower part of that shell. It is removed from the hot well by the " hot well pump " which discharges it to storage tanks or heaters, depending on the layout of the plant. The cooling water is supplied by the " circulating pump " shown. After making two " passes " through the tubes, it flows away at the " circulating water overflow." The noncondensible gases are drawn from an " air box " 282 STEAM POWER near the bottom of the condenser by means of an hydraulic air pump which will be described later. They flow through the pipe indicated as " dry air suction/' enter the " hydrau- lic vacuum pump," and are discharged (in intimate mixture with water) into the " sealing tank." Here the noncon- densible gases separate and escape through the " sealing tank vent." The water required by the hydraulic vacuum pump is circulated by the " hydraulic supply pump " and Turbo-Generator Atmosphere Supply Water fckalinf; Tan Wake-up Overflow Fig. 185. — Surface Condenser and Necessary Auxiliaries as Applied to Steam Turbine of Moderate Size. any loss is made good through the line marked " supply water make up." In the case shown, the hot well pump and the hydraulic supply pump are driven by a single steam turbine, the exhaust of which is carried to the feed water heater and heats the condensate on its way to the boilers. 118. Vacuum Pumps or Air Pumps. In an earlier sec- tion, attention was called to the fact that some means must be provided for removing noncondensible gases, or " air," CONDENSERS AND RELATED APPARATUS 283 from condensers. This may be done by a pump which handles both the condensate and the noncondensible mate- rial. Such a pump is known as a " wet vacuum pump." A section of a pump of this kind is shown in Fig. 176 and another in Fig. 184. For larger installations, and particularly those in which a very high vacuum is desired, it is customary to use one pump for handling the condensate and another pump for the noncondensible gases. The latter is known as a " dry vacuum pump " or more commonly, an " air pump." The earliest dry vacuum pumps were merely recipro- cating air compressors which worked between a high vacuum and atmospheric pressure. That is, they received " air " at the pressure existing in the condenser and compressed it to such a pressure that they could discharge it to atmos- phere. Such pumps are still in use. They are commonly arranged with crank and fly wheel and are known as " Rota- tive Dry Vacuum Pumps " or " R. D. V. Pumps." The air cylinder of such a pump is shown in Fig. 186. Air enters at the flange indicated and flows into the right- hand end of the cylinder through the upper set of valves as the piston moves to the left. On the return stroke the air is discharged without compression through the lower set of valves, and travels through a passage outside the cylinder to the space being opened up on the other side of the piston, entering through the valves shown at the left and above the bottom of the cylinder. When the piston makes its next stroke to the left, the air entrapped in the left-hand end of the cylinder is compressed and discharged through the valves at the bottom. Such a pump is de- scribed as " two stages in one cylinder." The right-hand end merely serves as a " loader " for the left-hand end and all compression occurs in the latter. This " loading " or " two stage " feature is introduced for the purpose of increasing the volumetric efficiency, that is, the amount of air handled per revolution. 284 STEAM POWER With the arrangement shown, the right-hand end of the cylinder and its clearance are completely filled with air at suction pressure at each stroke. That part filling the piston displacement is air drawn from the condenser and that part filling the clearance is air left over from the previous return stroke. All of the air filling the piston displacement is Water Jacket Discharge Valves ufy/y/A Second Stage Air Suction Fig. 1876. — Hydraulic Vacuum Pump. In the type shown in Fig. 187, water supplied by a centrifugal pump enters the air pump at the hurling water inlet. It flows out of the head of the air pump through the annular nozzle indicated and enters the revolving jet wheel with a high velocity. The water leaves the wheel CONDENSERS AND RELATED APPARATUS 287 in the form of a number of jets of approximately rectangular section, each jet traveling a helical path leading it into the throat of the discharge tube. In traveling across the space between the revolving wheel and the throat, the jets entrap air between them and carry this air into the discharge tube. In passing through the lower part of the discharge tube, the velocity of the mixture decreases and the pressure increases so that the mixture can be discharged into the sealing tank against a small head of water plus atmospheric pressure on the sur- face of the water. In the tank the air separates and escapes through the vent. The water is used over and over, any loss by evapo- ration or spillage being made up as convenient. In the steam jet type, one or more steam jets entangle or entrain the condensible gases while moving at a high velocity and the mixture then passes through an expanding tube or nozzle in which its velocity is reduced with a cor- responding increase of pressure. The design is such that the pressure rises to a value sufficiently above atmospheric to make possible the discharge of the mixture to the atmos- phere. Such a device is shown diagrammatically in its simplest form in Fig. 188. Steam at a high pressure enters the expanding steam nozzle designed to discharge it at a high velocity and with a pressure slightly lower than that carried in the condenser. The jet of steam passes through the space indicated as " entraining space " and into the " discharge tube." While passing through the entraining space, the jet picks up or entrains some of the noncondensible material held in that space and the mixture enters the discharge tube. The taper or flare of the discharge tube below the " neck " is so proportioned that the high velocity of the mixture is reduced and the pressure correspondingly in- creased to a value equal to or greater than atmospheric. The entraining space is connected to that part of the 288 STEAM POWER High Pressure Steam Inlet Steam Nozzle ■^fAir Inlet or Suction Flange hroat or Neck Diffusor or Discharge Tube condenser in which the nonconclensible gases collect. As this material is constantly removed from the entraining space by the steam jet, a continuous flow from con- denser to air pump results. Experience has shown that a single stage pump like that illustrated in Fig. 188 is not the best possible arrangement. As a con- sequence, two and three stage pumps have been pro- duced, the two stage being the type now generally used. Such a pump is illustrated in Fig. 189, and in Fig. 190 is shown one way in which it is connected into a con- denser installation. The device shown in Fig. 189 is called the Radojet Vacuum Pump, taking its name from the peculiar ar- rangement at the second stage which is described later. Instead of using one steam nozzle leading steam into the entraining space as shown in Fig. 188, a group of small nozzles is used. This is done for the purpose of breaking the steam up into a number of small jets, thus increasing the amount of air entrained by a given weight of steam. The mixture discharged from the diffusor of the first stage enters the second entrainment space. Steam admitted at the second stage steam inlet expands through the pecul- iarly shaped nozzle indicated and is discharged at high velocity in the form of a circular sheet. This sheet entrains the mixture discharged from the first stage and carries it Fig. 188.— Steam-jet Type Vacuum Pump. CONDENSERS AND RELATED APPARATUS 289 Steam Supply, First Stage Steam Strainer, First Stage h First Stage TSteam Nozzles First Stage Entrainment Space S 'Air Suction Second Stage Steam Strainer' Collecting Space dscharge Fig. 189. — Cross-sectional View of the Radojet Vacuum Pump. into the diffusor shown. This diffusor consists of two cir- cular plates which are close together near the steam nozzle and separate gradually toward their circumferences. The mixture from the second diffusor is discharged into the 290 STEAM POWER collecting space shown and flows out at the discharge flange. Exhaust Steam Inlet Fig. 190. — One Method of Connecting Radojet to Surface Condenser. The radial flow from the second stage steam nozzle and through the second stage diffusor gives the device its trade name. In the arrangement illustrated in Fig. 189, the second stage steam has to entrain and compress not only the " air " coming from the condenser but also all of the steam used in the first stage. This requirement leads to the use of large quantities of second stage steam. When thermal efficiency is of sufficient significance, it is customary to separate the first and second stages by interposing an " inter cooler." This is merely a small surface condenser in which first stage steam and possibly vapors brought over from the condenser can be liquefied, leaving only saturated air to be handled by the second stage jet. When such an intercooler is installed, some or all of the condensate from the main condenser is generally used as circulating water, so that the heat given up by steam and CONDENSERS AND RELATED APPARATUS 291 air is caught in Mi is condensate and returned to the boilers. 119. Water Required by Contact Condensers. The weight of circulating water required varies with the type of condenser and with the conditions of operation, such as initial temperature of water, vacuum desired, etc. It can be determined approximately by calculation and the values thus found must then be increased by such factors as experi- ence has shown to be necessary. In contact condensers the water and the condensate are discharged as a mixture and therefore have the same average discharge temperature. Let ^1 = initial temperature of injection water in F.°; fe ~ temperature at which mixture is discharged in F.°; X = total heat above 32° F. of steam as exhausted; W = pounds of injection water per pound of exhaust steam. Assuming the exhaust steam to be dry saturated, each pound of steam in condensing to water at a temperature of fe degrees must give up an amount of heat equal to X minus the heat of the liquid at t°2 or roughly X— (fe — 32) B.t.u. This same quantity must be absorbed by the in- jection water, while its temperature rises from h to fa degrees. Each pound of water can then absorb approxi- mately fo — ti) B.t.u. and the pounds of injection water per pound of steam will be w _ \-k+S2 i2 — h The value of fe would be that corresponding to the absolute pressure in the condenser if it were not for the air and similar gases which exert some pressure. It is generally 10 or more degrees F. below the temperature corresponding to the vacuum. Values of fe in the neigh- 292 STEAM POWER borhood of 110° to 125° F. are customary with recipro- cating engines and values as low as 80° are used with high vacuums in connection with steam turbines. The weight of water used per pound of steam as given by Eq. (78) will vary between about 15 for very low initial and moderate discharge temperature to about 50 with average initial and moderate discharge temperature. Ex- perience shows that it is necessary to add 10 per cent or more to the values of W obtained from equation (78) to obtain the weight of water which will probably be used. ILLUSTRATIVE PROBLEM Find the quantity of water theoretically required per pound of steam condensed in a contact condenser in which a vacuum of 25.5 ins. of mercury is maintained when the barometer reads 29.5 ins. of mercury. The initial temperature of the water is 60° F. The absolute pressure in the condenser is 29.5—25.5=4.0 ins. of mercury and the steam tables give for this pressure, X = 11 15.0 and to = 126. Substituting in Eq. (78) gives w 1115.0-126 +32 ^ = 7^ — ^ =15.5 approximately. 126—60 120. Weight of Water Required by Non-contact Condensers. In the case of non-contact condensers there is no definite relation between the discharge tempera- ture of the cooling water and that of the condensate. Experience shows that the discharge temperature of the circulating water is generally from 10 to 20 or more degrees lower than the temperature corresponding to the vacuum. The temperature of the condensate (hot-well tempera- ture) is often 15 or more degrees below that correspond- ing to the vacuum, but good design makes the hot-well temperature very closely approximate that corresponding to the vacuum. CONDENSERS AND RELATED APPARATUS 293 Assuming ^1 = initial temperature of injection water in F.°; i 2 = final temperature of injection water in F.°; t c = temperature at which condensate is discharged, i.e., hot- well temperature, in F.°; \= total heat above 32° F. of steam as exhausted, and W = pounds of injection water per pound of exhaust steam. The weight of water which must be circulated per pound of steam can be found as in the case of the contact con- denser. It is given by w = \-tc+Z2 (7Q) i2 — h Values in the neighborhood of 25 lbs. of water per pound of steam are common with low vacuums and 50 or more pounds are often used with vacuums over 28 ins. of mercury. ILLUSTRATIVE PROBLEM A surface condenser receives circulating water at a temper- ature of 65° F. and discharges it at a temperature of 80° F. It maintains a vacuum of 28.0 ins. with the barometer at 29.5, and the temperature of the condensate discharged to the hot well is equal to 85° F. Find the quantity of circulating water theoretically required. This vacuum corresponds to an absolute pressure of 29.5—28.0 = 1.5 ins. of mercury. Assuming this all due to steam (neglecting presence of air) the value of X may be taken from the steam table as 1100.1 B.t.u. Substitution 'in Eq. (79) then gives w 1100.1-85+32 " = en — az =69.9 approximately. oU — OO 121. Relative Advantages of Contact and Surface Con- densers. The contact types are as a rule much cheaper 294 STEAM POWER than the surface condensers, and they are less subject to serious depreciation, the tubes of surface condensers often corroding excessively in very short intervals of time. On the other hand, the injection of the cooling water into the condensing space in contact types results in the intro- duction of large quantities of dissolved gases, and much of this material is liberated under the reduced pressure, thus tending to increase the condenser pressure, that is, decrease the vacuum. Where pumps are used to carry away the mixture with contact condensers, these pumps have to handle a much larger quantity of water than the corresponding pump in a surface condenser, and the work of pumping this water out of the vacuum into the atmos- phere combined with the additional work required of the pump which handles the " air " may partly balance the advantage of lower first cost of the contact type. A surface condenser must always be installed where it is desirable to use the condensate as boiler feed, and it is generally used when very high vacuums (low absolute pressures) are to be maintained. The surface condenser is at a serious disadvantage, however, when required to handle the exhaust of reciprocating engines. The exhaust from such engines always contains large quantities of lubricating oil carried out of the cylinder, and unless this material is separated before the exhaust enters the con- denser it is deposited on the outer surfaces of the tubes and decreases the conductivity of those surfaces. Such oil can be eliminated to a great extent before the exhaust enters the condenser by means of oil separators, which are generally made up of a series of baffles upon which the steam impinges and upon which the oil is caught. 122. Cooling Towers. The large quantity of circula- ting water required by condensing plants is often an item of great economic importance. When such plants are located near a river or near tide water, the circulating water can generally be procured for the cost of pumping. CONDENSERS AND RELATED APPARATUS 295 When they are located in the middle of cities or in regions where water is scarce, the cost of water may be excessive or it may even be impossible to obtain a continuous supply equal to the demand of the condensers. In such cases the condensing water is often circulated continuously, being cooled after each passage through the condensers. This cooling is generally done by exposure of a large surface to the air. The resultant evaporation of some of the water with the absorption of its latent heat of vaporization cools the remainder so that it can be used again. This sort of cooling may be effected by running the water into a shallow pond of large area, or by spraying it into the air over a small pond or reservoir or by passing it through a cooling tower. Cooling towers are large wood or metal towers generally filled with some form of baffling devices. The hot water is introduced at the top and spread into thin sheets or divided up into drops as it descends. Air enters at the bottom and flows upward, cooling the water by contact and by the partial evaporation which results. The cir- culation of air may be natural, i.e., due to the difference of temperature between the. air inside and out, in which case a stack is fitted to the top of the tower; or the cir- culation may be forced by fans located in the base of the tower. In the latter case the apparatuses called a forced- draught cooling tower. CHAPTER XV COMBUSTION 123. Definitions. Certain substances are known to chemists as compounds, because they can be separated by chemical processes into simpler substances. Thus water and many of the most familiar materials known to man are compounds which can be separated into two or more simpler materials. Those substances which cannot be further broken up by the processes used in separating compounds are called elements; they are regarded as elemental, as the stones of which the compounds of nature are built up. About eighty-three of these elements are now known, but many of them are comparatively rare. Pure metals are all elements; the oxygen and nitrogen which are mixed to form the greater part of the atmosphere are elements; carbon, which forms the greater part of most fuels, is an element. In many cases the combination of elements to form compounds is accompanied by the liberation of heat, and some of these combinations are used by the engineer for the purpose of obtaining heat in large quantities. When the elements which occur in fuels, such as coal, wood and petroleum, combine with oxygen, the process is spoken of as combustion. The quantity of heat liberated when a pound of any material combines with oxygen (burns) is called the heat value or calorific value of that material. Fuels contain a great number of elements, but only three of these ordinarily take part in combustion and are therefore spoken of as combustibles. They are carbon, hydrogen and sulphur. The sulphur content is generally 296 COMBUSTION 297 very small, and the carbon and hydrogen are therefore the most important constituents. The combustion of each of these elements will be con- sidered in detail in the following sections, but before this can be done two other ideas must be developed. The smallest particle of an element which can be conceived of as entering into combination to form a com- pound is known as an atom of that element. It has been found that the atoms of each element have an invariable and characteristic mass. The lightest atom is that of hydrogen, and its weight is considered unity. The atom of carbon is twelve times as heavy as that of hydrogen and carbon is therefore said to have an atomic weight equal to twelve. Similarly the atomic weight of nitrogen is four- teen and that of oxygen is sixteen. The smallest particle which can be formed by the com- bination of atoms is known as a molecule. Like or unlike atoms may combine to form molecules. Thus two hydro- gen atoms combine to form a molecule of hydrogen, and hydrogen gas as it ordinarily exists may be pictured as made up of a collection of such molecules. Similarly, gaseous oxygen and gaseous nitrogen may be pictured as collections of molecules which are made up of two like atoms. When unlike atoms combine to form a molecule, they form a molecule of a compound. Obviously a molecule of any compound is the smallest particle of that compound which can exist. For convenience, the different elements are represented by abbreviations; thus oxygen is represented by 0, nitro- gen by N, hydrogen by H, carbon by C and sulphur by S. When these abbreviations are written in chemical equa- tions, such as will be given later, they stand for an atom of the substance. Hence in a chemical equation would mean one atom of oxygen. The symbol O2 is used to mean two atoms of oxygen in combination, hence, one molecule 298 STEAM POWER of oxygen. The symbol 2O2 means two groups of two oxygen atoms in combination, hence two molecules of oxygen. The simplicity and elegance of this system will become apparent as the chemical equations which follow are de- veloped and explained. 124. Combustion of Carbon. Carbon can unite with oxygen or burn to form two different compounds — carbon monoxide (CO) and carbon dioxide (CO2). The monoxide is formed by the combination of one atom of oxygen with one atom of carbon; the dioxide, by the combination of two atoms of oxygen with one of carbon. The dioxide, therefore, contains twice as much oxygen as does the monoxide. Carbon burned to carbon monoxide has not combined with the largest possible quantity of oxygen, and combus- tion is therefore said to be incomplete in such cases. When, however, carbon dioxide is formed, the carbon has combined with as much oxygen as possible and combustion is said to be complete. It will be shown later that much more heat is liberated when the dioxide is formed than when carbon burns to the monoxide. Hence, when liberation of heat is the object of combustion, the process should be so conducted as to result in the formation of the maximum quantity of dioxide and the minimum amount of monoxide. 125. Combustion to CO. The combustion of carbon and oxygen to form the monoxide can be represented by the equation C+0 = CO, (80) or by the equation 2C+0 2 = 2CO (81) The former is the simpler and will be considered first, but the latter is the more perfect and indicates more to the trained eye than does the simpler form. The simple equation states that one atom of carbon combined with one atom of oxygen to form one molecule COMBUSTION 299 of carbon monoxide. It can, however, be so interpreted as to show much more than this. The carbon atom is twelve times as heavy as the hydrogen atom, while the oxygen atom is sixteen times as heavy as that of hydrogen. The equation C + = CO, therefore, shows that an atom, which is twelve times heavier than the hydrogen atom, unites with one which is sixteen times heavier than the hydrogen atom to form a molecule which is 28( = 12 + 16) times heavier than the hydrogen atom. In other words, the weights of carbon and oxygen 12 3 1 combining are in the ratio of t^ = t = 7T- If a sufficient lb 4 lg number of carbon atoms to weigh one pound be used, a quantity of oxygen weighing lj lbs. will be required to combine with them to form carbon monoxide. The re- sultant carbon monoxide will contain the pound of carbon and the 1J lbs. of oxygen and will therefore weigh 2f lbs: The same weight relations would hold irrespective of the weight of carbon used, and the simpler equation may therefore be put 1 weight of C+l| weights of = 2^ weights of CO. (82) ILLUSTRATIVE PROBLEM To illustrate the use of this equation, assume that 9 lbs. of carbon are burned to carbon monoxide and that it is desired to find the weight of oxygen used, and the weight of the product. The weight of oxj^gen used must be 1^ times the weight of carbon, that is, 1^X9 = 12 lbs. The weight of the product must be 2\ times the weight of the carbon, that is 2f X9 =21 lbs.; or, it must be the weight of the carbon burned plus the weight of the oxygen used, that is, 9+12 =21 lbs. In general, the oxygen used for combustion is obtained from the atmosphere, which may be considered as a median- 300 STEAM POWER ical mixture of oxygen and nitrogen in unvarying porportions. These proportions are roughly, 0.23 of oxygen to 0.77 of nitrogen by weight, or 0.21 of oxygen to 0.79 of nitrogen by volume, as shown in Table VIII. The weight of air which contains one pound of oxygen is therefore 234-0 77 ' — = 4.35 lbs., and this weight of air contains 4.35-1=3.35 lbs. of nitrogen. In the problem previously considered it was found that 12 lbs. of oxygen would be required to burn 9 lbs. of carbon to CO. The total weight of air required to obtain this oxygen will be 12X4.35 = 52.2 lbs. and it will contain 52.2 — 12 = 40.2 lbs. of nitrogen. By simple arithmetical calculations of the type just given all the weight relations involved in the combustion of C to CO can be determined. The volume of air required in any given case can be found by multiplying the weight of air by the specific volume as given in Table VIII. Thus, in the illustrative problem already considered, it was found that 52.2 lbs. of air would be required to burn 9 lbs. of C to CO. The volume of this air at 62° F. would be 52.2X13.14 = 685.9 cu.ft. It is found that a quantity of heat equal to about 4500 B.t.u. is liberated per pound of carbon burned to CO; that is the calorific value of C burned to CO is 4500 B.t.u. Returning now to Eq. (81), which was said to be more useful than the simpler form given as Eq. (80), it will be necessary to consider a rather simple law of gases. It has been shown experimentally that equal volumes of all gases contain the same number of molecules when at the same temperature and pressure. This statement is known as Avogadro's Law.. It has also been shown that the mole- cules of gaseous oxygen contain two atoms. The equation in question, 2C+0 2 = 2CO COMBUSTION 301 can therefore be read, two atoms of carbon combine with one molecule of oxygen to form two molecules of carbon mon- oxide. But, if every molecule of O yields two molecules of CO it follows from Avagadro's law that the CO formed will occupy twice the volume of the oxygen used if measured at the same temperature and pressure. If the equation be imagined as containing a numeral 1 before the O2, it will be obvious that the coefficients of the terms represent- ing gas molcules give volume relations directly. This equa- tion therefore gives both volume and weight relations. TABLE VIII Properties of Air Considering it to consist only of nitrogen and oxygen. Relative Proportions. Ratio of N to 0. Ratio of Air to O. Exact. Approx. Exact. Approx. Exact. Approx. By weight . . By volume. . / 0.766 N 1 0.234 / 0.791 N 10.209 0.77N 0.23 O 0.79 N 0.21O 3.27 3.78 3.35 3.76 4.27 4.76 4.35 4.76 Spec. wt. at Atmos. Press. (Lbs. per Cu.ft.) Spec. Vol. at Atmos. Press. (Cu.ft. per Lb.) At 32° F. At 62° F. At 32° F. At 62° F. 0.08072 0.07609 12.39 13.14 Weight of air containing one pound of oxygen is approximately 4.35 lbs. 126. Combustion to CO2. The combination of carbon and oxygen to form the dioxide is represented by the equa- tion C4-0 2 = C0 2? (83) 302 STEAM POWER which shows that one atom of carbon (twelve times heavier than hydrogen) combines with two atoms of oxygen (each sixteen times heavier than hydrogen) to form a molecule of CO2, which is forty-four times heavier than an atom of hydrogen. Therefore the weight of carbon and oxygen 12 3 1 combining are as =^ = 7^2, so that 2| lbs. of oxygen are required to burn a pound of carbon to carbon dioxide. Writing this in the form of an equation, gives 1 weight of C+2§ weights of = 3f weights of C0 2 . . (84) The weight of air required can readily be found by multiplying the required oxygen by the number 4.35, previously shown to be the number of pounds of air con- taining one pound of oxygen. Thus, the required air is 2|X4. 35 = 11.57 pounds per pound of C burned to CO2. This number is commonly rounded out to 12 in engineering literature. The equation given shows volume relations directly. It is evident, therefore, that one molecule of O yields one molecule of CO2, and hence that the volume of the product is exactly equal to the volume of the oxygen used in forming it if measured at the same temperature and pressure. This is a very important relation, and is often made use of in engineering calculations. Experiment shows that when carbon burns to the dioxide about 14,600 B.t.u. are liberated per pound of carbon burned, that is, the calorific value of C burned to CO2 in 14,600. 127. Combustion of CO to CO2. Since carbon which has burned to carbon monoxide has not combined with the greatest possible quantity of oxygen, the monoxide can take up more oxygen by burning to the dioxide. This process is represented by the formula 2CO+0 2 = 2C0 2 , (85) COMBUSTION 303 which shows that two molecules of monoxide combine with one molecule of oxygen to form two molecules of the dioxide. The volume of CO2 formed is therefore equal to that of the CO burned. So far as the ultimate result is concerned, it makes no difference whether carbon is burned directly to CO2 or is first burned to CO and then the CO is burned to CO2. The total oxygen used per pound of carbon burned to CO2 and the total heat liberated per pound of carbon burned to CO2 are the same in both cases. Thus, for the oxygen, one pound of C burned to CO2 requires 2| lbs. of oxygen; but one pound of C burned to CO requires 1J lbs. of oxygen, and 1J lbs. more will be required when this CO is burned to CO2. The result is therefore the same. For heat liberated, one pound of C burned to CO2 liberates about 14,600 B.t.u.; but one pound of C burned to CO liberates about 4500 B.t.u. and 10,100 B.t.u. are liberated when this CO is burned to CO2. Since the sum of 4500 and 10,100 is equal to 14,600 the result is again the same. Data on the combustion of C to CO and CO2 and the combustion of CO to CO2 are collected in convenient form in Table IX. 128. Conditions Determining Formation of CO and CO2. Excluding certain complicated considerations which are not of great importance in steam-power engineering, it may be said that when carbon is being burned at a certain rate (pounds per unit "of time) the amount of oxygen brought into contact with ,the carbon determines whether the caibon burns to CO or to CO2. If enough or more than enough oxygen to burn the carbon to CO2 is brought into contact, that oxide will be formed. If there is not enough to burn all the carbon to the dioxide, both oxides are formed in cer- tain proportions, which can be calculated. Since combustion to CO yields only 4500 B.t.u. per 304 STEAM POWEK pound of C and combustion to CO2 yields 14,600 B.t.u. per pound of C, the importance of supplying sufficient oxygen to burn all carbon to the dioxide in cases where the liberation of the maximum quantity of heat is desirable is obvious. In actual practice the oxygen is furnished by supplying air and it is found necessary in most cases to supply more than the amount of air theoretically re- quired in order to insure burning all, or even nearly all, of the carbon to the dioxide. This comes from the great difficulty met in obtaining contact between the oxygen of the air and the carbon which is to be burned, that is, in bringing all the oxygen of the air into intimate contact with the fuel being burned in real apparatus. TABLE IX Combustion Data for Carbon (Per pound of carbon.) Product. Oxygen Required. Nitrogen Accompanying Oxygen. Pounds. Cu.ft. at 62° F. and 14.7 Lbs. Pounds. Cu.ft. at 62° F. and 14.7 Lbs. CO 1.333 2.667 1.333 16.0 32.0 16.0 4.46 8.92 4.46 60 1 C0 2 fromC CO. from CO.... 120.2 60.1 Air Required. Quantity of Product (N not included). Heat Liber- ated. Product. Pounds. Cu.ft. at 62° F. and 14.7 Lbs. Pounds. Cu.ft at 62° F. and 14.7 Lbs. CO 5.79 11.58 5.79 76.1 152.2 76.1 2.33 3.67 3.67 32.0 32.0 32. oj 4,500 14,600 10,100 per lb. of C in CO 4,300 per lb. of CO C0 2 from C C0 2 fromCO.... COMBUSTION 305 The air in excess of that theoretically required to burn all the carbon completely is spoken of as excess air. In the form of an equation, this statement is equivalent to Air supplied — air theoretically required = excess air. (86) It is customary to express the quantity of excess air in terms of a numerical factor known as the excess coefficient. This coefficient is denned as the number by which the quantity of air theoretically required must be multiplied to give the quantity of air actually used. In the form of an equation this gives Excess coefficient X air theoretically required = air actually used. . (87) ILLUSTRATIVE PROBLEM Taking data from the illustrative problem previously considered, assume that 9 lbs. of carbon are burned in air to C0 2 . Each pound theoretically requires 11.57 lbs. of air, so that the theoretical air-supply for this case would be 9X11.57=104.13 lbs. If in a real case 150 lbs. of air are supplied, the excess coefficient is equal to 150-^ 104.13 =1.44. 129. Flue Gases from Combustion of Carbon. The gases resulting from the combustion of fuels are known in engineering as the products of combustion or flue gases, because they are the gases passing through the flues or passages leading from furnaces in which fuel is burned and to the stacks which serve to carry off the gases. It has already been shown that the CO2 formed by the combustion of carbon has the same volume as the oxygen which is used in forming it. Therefore, if the air supplied in a given case just equaled that theoretically required for combustion to CO2 and if all of the oxygen were used, the CO2 formed would merely replace the oxygen in the air. The theoretical proportions of the flue gas would then be 0.21 of C0 2 and 0.79 of N by volume. 306 STEAM POWER If real flue gases obtained by burning carbon in air are found to contain less than 21 per cent of C0 2 , the combustion has evidently not yielded theoretically perfect flue gases. The trouble may be due to an excess or to a deficiency of air. If there is an excess of air there will be oxygen present in the flue gases; if there is a deficiency there will be CO present in the flue gases. An analysis of these gases for oxygen and for CO would therefore indicate the source of trouble and the remedy to be provided. The curve to the right of the central vertical line in Fig, 191 shows the theoretical decrease in volume per 35 \ | 25 > 8 15 > O £5 \ \ \ \ \ 'A & . \ <% \ 50 40 30 10 Deficiency (in per cent) 50 100 150 200 250 300 Excess (in per cent) 2 3 1 Fig. 191. — Effect of Air Supply on Flue Gas Analysis. cent of CO2 in flue gases as the excess air increases. The single numbers 1, 2, 3 and 4 indicate the excess coeffi- cients corresponding to the various percentages of excess air. The curves to the left give the theoretical decrease in volume per cent of CO2 and the theoretical increase in volume per cent of CO as the air supplied is decreased below that theoretically required for complete combustion. 130. Combustion of Hydrogen. Hydrogen combines with oxygen, or burns, to form water. The equation for this reaction is 2H 2 +0 2 = 2H 2 0, ....*.. (88) COMBUSTION 307 which indicates that two molecules of hydrogen combine with one molecule of oxygen to form two molecules of water. In terms of volumes, two volumes of hydrogen combine with one of oxygen to form two of gaseous water, that is, water in the form of highly superheated vapor. As the water is cooled down it will obviously approach and finally reach the liquid condition, with a rapid de- crease in volume quite different from that experienced by a gas under similar conditions, so that the volume rela- tions hold only at high temperatures. The weight relations can be calculated as in other cases, starting from the fact that four weights of hydrogen combine with thirty-two weights of oxygen to form 36 weights of water. The weights of hydrogen and oxygen are therefore in the relation of -^ = |. The heat liberated when one pound of hydrogen burns to water is equal to about 62,000 B.t.u. This is the quantity of heat which could be obtained if one pound of hydrogen at, say, room temperature, and mixed with the requisite quantity of oxygen, were ignited and the resultant water were then cooled down to the initial temperature. During the cooling of the water it would partly or entirely condense and thus give up some or all cf its latent heat of vaporization. This heat would obviously be included in the calorific value just given. In many pieces of engineering apparatus in which hydrogen is burned the products of combustion are not cooled to such an extent that the water is condensed. The latent heat of vaporization would not be liberated under such conditions, but would remain bound up with the water vapor. When the water is not condensed the heat liberated is only about 52,000 B.t.u. per pound of hydrogen. This number is known as the lower calorific value of hydrogen, while 62,000 is known as the higher calorific value. Data on the combustion of hydrogen are given ip Table X. 308 STEAM POWER TABLE X Combustion Data for Hydrogen (Per pound of hydrogen) Oxygen Required. Nitrogen Accompanying Oxygen. Product. Pounds. Cu.ft at 62° F. and 14.7 Lbs. Pounds. Cu.ft. at 62° F and 14.7 Lbs. H 2 8 96 26.8 361 Air Required. Quantity of Product (N not included). Heat Product. Pounds. Cu.ft. at 62° F. and 14.7 Lbs. Pounds. Cu.ft. at 62° F. and 14.7 Lbs. Liberated. H 2 34.8 457 9 Liquid 0.144 / 62,000 I 52,000 131. Combustion of Hydrocarbons. Many of the fuels used by the engineer contain compounds of hydrogen and carbon which are called hydrocarbons. One of the best examples is methane (CH4), which forms the greater part of all the so-called natural gas. All of these hydrocarbons burn to CO2 and H2O if the supply of oxygen is great enough. If there is a deficiency of oxygen, combustion is incomplete and generally results in the formation of CO2, H2O, CO, C in the form of soot, and other products which need not be considered here. For complete combustion the requisite oxygen and air can be determined as in previous cases by means of chemical equations. Thus for methane the equation is CH 4 +202 = C02+2H 2 0, (89) which shows that sixteen (12+4) weights of methane combine with sixty-four (2X2X16) weights of oxygen to form forty-four (12+32) weights of carbon dioxide and thirty-six (4+32) weights of water. COMBUSTION 309 The calorific value of hydrocarbons is generally assumed to be equal to the sum of the heat values of the carbon and hydrogen contained in one pound of the material. Thus, if C represent the fraction of a pound of carbon contained in one pound of the hydrocarbon and if H represent the fraction of a pound of hydrogen contained therein, the common assumption would make the higher calorific value of the hydrocarbon (C X 14,600) + (HX 62,000) B.t.u. . . (90) The results obtained in this way do not generally check well with the experimentally determined values, and it is best to use the latter when they are available. 132. Combustion of Sulphur. Sulphur forms several different oxides, but when burned under engineering con- ditions it is generally assumed to form only the dioxide SO2. The chemical equation for such combustion is S+0 2 = S0 2 , (91) and since the atomic weight of sulphur is 32, this equation shows that equal weights of sulphur and oxygen combine to form the dioxide. The combustion of sulphur to SO2 liberates about 4000 B.t.u. per pound of sulphur. 133. Combustion of Mixtures. It is often necessary to obtain approximate calorific values of combustible materials which, without great error, can be considered as mixtures of combustible and non-combustible elements. If there is oxygen present in the mixture it is assumed to be combined with hydrogen in the form of water, so that the uncombined or available hydrogen per pound of material is given by the expression Available H = H-^, (92) 310 STEAM POWER in which H and O respectively represent the fractions of a pound of hydrogen and oxygen in one pound of material. The calorific values of such a mixture containing car- bon, hydrogen and sulphur would then be given approxi- mately by the equation Higher B.t.u. = 14,6000+62,000 (h-^-) +4000S, (93) in which the letters stand respectively for the fractions of a pound of each of the elements present in one pound of the mixture. Similarly the lower calorific value would be (approximately) Lower B.t.u. - 14, 600C +52,000 /h-^) +4000S, (94) and the oxygen required will be Pounds of = 2fC+8(H-^)+S. . • (95) 134. Composition of Flue Gases. It was shown in Section 129 that the flue gases resulting from combustion of carbon to carbon dioxide with the theoretical amount of air would consist of 21 per cent carbon dioxide and 79 per cent nitrogen by volume. It was also shown that vari- ation from this composition in any real case can be inter- preted to show the cause of such variation. It is necessary to note that the figures given apply only to the case of pure carbon. If, for instance, hydrogen is burned with air, no carbon dioxide can result, since no carbon is present and the composition of the flue gases must therefore be quite different from what was indicated above. As a matter of fact, the flue gases resulting from the com- bustion of hydrogen with the theoretical air supply would, if cooled down to ordinary temperatures, consist of nitro- gen saturated with water vapor, All water vapor in excess COMBUSTION 311 of that required to fill the space occupied by the nitrogen (at the existing temperature) would condense out as liquid water. All real fuels contain carbon and hydrogen and often sulphur as well. The theoretical composition of flue gases obtainable with real fuels is therefore quite different from that indicated for pure carbon. The theoretical air supply for such fuels contains just enough oxygen to burn the com- bustible constituents and the quantity of nitrogen which must accompany that oxygen. Irrespective of total quan- tities involved, the oxygen must represent about 21 per cent of the volume of the air and the nitrogen 79 per cent. After combustion is completed, some of this oxygen has been converted into carbon dioxide occupying the same volume as the oxygen from which it was made (when reduced to the same pressure and temperature). Obviously with real fuels the carbon dioxide must occupy less volume than all of the oxygen and therefore must form less than 21 per cent of the volume of the final products of combustion. The figures given below will serve to indicate the varia- tions which occur with real fuels. In each case a typical analysis has been assumed for the fuel named and it has beeen assumed that it is burned with the theoretical air supply. It should be noted that the percentage of carbon dioxide given holds only for the particular analysis of fuel assumed. Per Cent Carbon Fuel. Dioxide in Flue Gases. Bituminous Coal 18.4 Wood 20.1 Petroleum Oil 15 .4 Natural Gas 11.7 Blast Furnace Gas 25 . 1 The high value for blast furnace gas is explained by the fact that this gas contains a large amount of carbon dioxide. 312 STEAM POWER so that this quantity appears in the flue gas in addition to that formed by combustion of other constituents of the gas. 135. Temperature of Combustion. If combustion of any material could be carried on inside an ideal vessel which did not absorb nor transmit heat, the heat liberated during the combustion could not escape from the space within the vessel. If the vessel contained initially only the combustible and the oxygen or air required to burn it, the products of combustion would be the only material contained within the vessel after the completion of combustion. Under such circumstances the heat would be used in raising the temperature of the products of combustion, and the process could be pictured as though all of the combustion occurred first, forming the products of combustion without change of temperature, and then the liberated heat raised the temperature of these products. Knowing the weight of each of these products and the quantity of heat required to raise the temperature of one pound of each of them one degree, the amount of heat required to raise all of them one degree could be found by multiplying the two known values. Thus, if carbon had been burned in oxygen to CO2 with the theoretical oxygen supply, the vessel would contain only carbon dioxide. To raise the temperature of one pound of carbon dioxide one degree requires an amount of heat equal to the specific heat of that gas. Therefore, if W represents the weight of CO2 formed and C represents its specific heat, the amount of heat required to raise the temperature of all of the CO2 one degree would be W-C B.t.u. If Q B.t.u. were liberated by the combustion, the temperature rise in degrees would therefore be given by Temp, rise = t|L (96) COMBUSTION 313 and if the initial temperature had been to degrees, the final temperature would be < = s 60° F., the mean specific heat of C0 2 is 0.27, that of H 2 is ,61, that of N is 0.27, and that of is 0.24. CHAPTER XVI FUELS 136. Commercial Fuels. In engineering practice any- thing which is combustible and which can be procured in large quantities at a reasonable cost is called a fuel. The principal commercial fuels are: f(l) Coal. a. Solid 1 (2) Wood and wood wastes. 1(3) Vegetable wastes. {(1) Crude petroleum or natural oil. (2) Various products made from petroleum. (3) Methyl and ethyl alcohol. (1) Natural gas. (2) Artificial or manufactured gases. c. Gaseous Coal is by far the most extensively used fuel because of its abundance and relative cheapness in most localities. However, in oil-producing regions the crude oil and some of the products made from it are more often the commonly used fuel, particularly if good coal is not mined in the immediate vicinity. Wood is, in general, too valuable to be used exclusively as a fuel excepting on the frontiers where wooded terri- tory is being opened up and where coal cannot be pro- cured excepting at prohibitively high cost. Wood wastes, on the other hand, are very often used for fuel in the indus- tries producing them. Vegetable wastes, like wood wastes, are essentially of local value, being practically entirely consumed by the industries producing them. 317 318 STEAM POWER Natural gas is in many respects an ideal fuel, and is extensively used for power production in some localities. The diminution in the quantity available, the consequent rise in the price, the great economy achieved in burning this gas in gas engines and the increased use of the gas for domestic purposes are, however, gradually eliminating this fuel from the steam-power field. Artificial gases have never been extensively used for the generation of steam, as it is generally cheaper to burn the materials from which the gases are made, rather than to convert them into gas and then to burn the gas under boilers. This condition may change in the future when better markets have been opened up for some of the by- products which can be obtained from artificial gas plants. 137. Coal. The word coal is used as the name of a great group of natural fuels which consist of more or less metamorphosed vegetable remains. At one end of the group is the material known as peat, which is only slightly changed from the original vegetable substance; at the other end is the graphitic anthracite which has undergone such radical metamorphosis that practically all of the original vegetable material excepting carbon and ash has been eliminated. A common, rough classification of the coals in the order of age, or of completeness of carbonization is, 1. Peat or turf. 2. Lignite (brown or black). 3. Sab-bit aminous coal. 4. Bituminous coal. 5. Semi-bituminous coal. 6. Semi-anthracite. 7. Anthracite. 8. Graphitic anthracite. The divisions are not at all exact, as they depend partly upon chemical composition and partly upon physical properties. FUELS 319 Another classification of a more exact variety is that given in Table XI and partly illustrated in Fig. 192, which gives what is known as Mahler's curve. It is for United States coals only. The terms used in this classification are explained in subsequent paragraphs. 60 70 80 90 $Fixed Carbon in the Combustible 100 Fig. 192. — Mahler's Curve for United States Coals. TABLE XI Classification of Coals Division. Graphitic Anthracite Semi-anthracite .... Semi-bituminous. . . Eastern bituminous. Western bituminous Lignite Per cent of Fixed Carbon in Combustible 100 to 97 97 to 92.5 92.5 to 87.5 87.5 to 75 75 to 60 65 to 50 under 50 Per cent of Volatile Matter in Combustible. to 3 3 to 7.5 7.5 to 12.5 12.5 to 25 25 to 40 35 to 50 over 50 Calorific Value, B.t.u. per Pound of Combustible. 14,600 to 14,900 to 15,300 to 15,600 to 15,800 to 15,200 to 13,700 to 14,900 15,300 15,600 15,900 14,800 13,700 11,000 The graphitic anthracite occurs in very small quantities and mostly in Rhode Island, With a few minor exceptions the anthracites occur only in Eastern Pennsylvania and the 320 STEAM POWER semi-anthracites are almost entirely confined to the western edge of this field. The semi-bituminous coals are found on parts of the eastern border of what is known as the Appalachian coal field, extending from central Pennsylvania through the intermediate States to the northern part of Alabama. The greater part of this enormous bed consists of eastern bitu- minous coal. Western bituminous coals are found in large beds in the central part of the United States, principally in the States of Illinois, Indiana and Kentucky on the east of the Mississippi River, and in Iowa, Kansas and Texas to the west of that river. Lignite is found in small quantities in nearly all of the western half of the United States and in large beds in the Dakotas, Texas, Arkansas, Louisiana, Mississippi and Alabama. Peat is distributed in small beds throughout practically all of the United States and is continually forming in many marshes and on low-lying lands. 138. Coal Analyses. Two different coal analyses are in use, the simpler being known as the proximate analysis and the more exhaustive being called the ultimate analysis. Both are made and reported on a weight basis. The proximate analysis assumes coal to contain four different and separable things, which are called fixed carbon, volatile hydrocarbon or volatile matter or volatile, moisture and ash. Moisture is determined by maintaining a small quantity of finely ground coal at a temperature of about 220° F. for one hour. The material lost during this time is assumed to be moisture only and is reported as such. Coal from which the moisture has been driven in this way is called dry coal. Volatile matter is determined by heating a sample from which the moisture has been driven, or a fresh sample. The coal is maintained at a red to white heat with exclu- FUELS 321 sion of air until there is no further loss of weight. In the case of a previously dried sample the loss under these conditions is called volatile hydrocarbon. If the sample was not previously dried a separate moisture determina- tion is made on a similar sample and the weight of volatile is found by difference. Fixed carbon is found by combustion of a sample from which the moisture and volatile have been driven, the loss under these conditions being assumed to be entirely due to the combustion of carbon. Ash is the name given to the incombustible material left behind after determining the fixed carbon. The volatile hydrocarbons and the fixed carbon as determined in the proximate analysis are assumed to be the only combustible parts of the coal and their sum is called the combustible. Proximate analyses are reported in three different ways: On coal as received, on dry coal, and on combustible. Since the water content of a sample of coal received at any plant is largely a matter of the weather conditions during shipment, the best idea of the character of a coal can be obtained by excluding the consideration of its moisture content. It is generally best, therefore, to convert analyses to a dry coal basis, that is, recalculate the per- centages of volatile, fixed carbon and ash on the assumption that the analysis was made on the weight of coal which would result from drying the sample that was actually used. Ex- cessive moisture is, however, undesirable for steam-raising purposes, and the amount of moisture should therefore be determined in every case. Ash is also more or less a matter of accident in that the amount contained is largely determined by the care used in mining and subsequent cleaning of the coal. While it has a very appreciable effect upon the character of the material as a fuel it really has little connection with the combustible part of the fuel. For purposes of classifica- 322 STEAM POWER tion, therefore, the ash should also be eliminated and the analysis given on the basis of combustible. Sulphur is sometimes reported with a proximate analy- sis. In making such an analysis the greater part of the sulphur is really driven off with, and regarded as, part of the volatile, so that when the sulphur content is desired it must be determined by a separate analysis. The ultimate analysis attempts to separate the dry combustible into the various elements of which it is com- posed. The percentages of carbon, hydrogen, oxygen, nitrogen and sulphur are determined as well as the per- centage of ash in dry coal. Such analyses show the carbon contents of coal to vary from about 98 per cent in the graphitic anthracite through about 97 per cent in the semi-anthracite, 87 per cent in semi-bituminous, 80 per cent in bituminous and 74 per cent in lignites to as low as 61 per cent in peats. The corresponding figures for hydrogen run from about 1 per cent through a range in the neigh- borhood of 5 per cent for semi-bituminous to about 6 per cent in the case of peat. Oxygen varies from about 2 per cent or less in the case of good anthracite to as high as 33 per cent for peat; nitrogen generally forms about 1 per cent of the dry fuel and sulphur from 1 to 3 per cent. 139. Calorific Value of Coals. The calorific value of coals on a basis of combustible has been shown to vary approximately according to a smooth curve, but the local variations are so great that no generally applicable formula for calorific value has yet been proposed. The formula commonly used is based upon the ultimate analysis and is similar to that suggested as approximately applicable in the case of mixtures of combustibles. It is known as Dulong's formula, and is [62,0001 / n \ B.t.u. perlb. = 14,600C + or ( H-^-J +4000S, (100) [52,000 J V */ FUELS 323 in which the letters refer to the weight of the various ele- ments contained in one pound of dry coal. When an accurate knowledge of the calorific value of a fuel is desired it should be obtained by means of a fuel calorimeter. There are many varieties of this instrument, but practically all operate on the same general principle. A known weight of fuel is completely burned within a vessel and the heat liberated is absorbed by water or similar liquid. From measurements of liquid temperatures the heat absorbed by the liquid can be determined, and this with some additions for losses of various kinds must be the heat liberated by the fuel. For details see Chapter XX. 140. Purchase of Coal on Analysis. Until quite recently it was customary to buy coal from the lowest bidder pro- vided the material supplied could be made to give satis- factory results in the plant. Obviously the purchaser knew nothing regarding his purchase, and often bought quantities of ash and moisture at the price of combustible. Now, however, the larger power plants and many of the smaller are buying on the basis of analyses and calorific values as determined in calorimeters. A certain desirable standard analysis is set and cer- tain variations are allowed from it. Wide variations are penalized by deducting so many cents per ton for each variation of a certain degree, and, finally, outside limits are set for moisture and ash beyond which the fuel need not be accepted. In some cases limits are also set for sulphur. This is the logical method of purchasing coal in large quantities, and is sure to come into very general use as its advantages become known. 141. Petroleum. This material is obtained from drilled wells and has been found in many widely separated sections of the country. The oil wells of Pennsylvania and neigh- boring States, of Oklahoma, Texas and California have been the most productive and are hence the most widely known. Natural petroleum, as it occurs in the United States, is 324 STEAM POWER generally a dark, rather thick, oily liquid with a char- acteristic odor. It varies widely in composition so far as the compounds contained are concerned, but the variations in ultimate composition, specific gravity and calorific value are comparatively small. The ultimate analysis of crude oil generally shows about 83 to 85 per cent of carbon, 13 to 15 per cent of hydrogen and small quantities of oxygen, nitrogen and sulphur. The specific gravity generally lies between 0.80 and 0.90 and in most cases is nearer the upper figure. It is common practice to express the gravity in terms of the Beaume scale, an arbitrary scale developed for an instrument known as the Beaume hydrometer. This device is arranged to float in liquids and measures the gravity by the distance to which it sinks. Various corresponding values of the Beaume scale and specific gravity are given in Table XII for the region most used in connection with petroleum. TABLE XII Corresponding Beaume Readings and Specific Gravities Beaum6 Reading. Specific Gravity. Beaume' Reading. Specific Gravity. 20 0.9333 34 0.8536 22 0.9210 36 0.8433 24 0.9090 38 0.8333 26 0.8974 40 0.8235 28 0.8860 42 0.8139 30 0.8750 44 0.8045 32 0.8641 46 0.7954 The higher calorific value varies between 19,000 and 20,000 B.t.u. per pound and the lower value is generally 1000 to 1500 B.t.u. lower. Crude oil is sometimes used for fuel, but this is unde- sirable, for two reasons. First, the crude oil contains many highly volatile constituents which can be distilled FUELS 325 off and which have a high market value in the forms of gasoline and allied distillates. Second, the presence of these highly volatile constituents in the oil makes it more dangerous, as combustible vapors are given off in large quan- tities at low temperatures and the mixtures formed with the oxygen of the air are often highly explosive. As a consequence, the material generally sold as fuel oil is a residuum left after distilling off the more volatile constituents of the crude oil. It has practically the same properties as the crude, excepting that dangerous vapors are not given off at so low a temperature. PROBLEMS 1. A sample of coal gives the following proximate analysis: moisture, 5%; volatile, 4.25%; fixed carbon, 80.75%; and ash, 10%. Determine the percentage of combustible and the percentages of fixed carbon and of volatile in the combustible. 2. What variety of coal is indicated by the values obtained in Prob. 1? 3. The following results were obtained in making a proximate analysis of a sample of coal; moisture, 7%; fixed carbon, 56.7%; volatile, 24.3%; ash, 12%. Determine the percentage of com- bustible and the percentages of fixed carbon and of volatile in the combustible. What variety of coal is indicated by these values? 4. The ultimate analysis of a sample of dry coal gave the following results: carbon, 79.12%; hydrogen, 4.14%; oxygen, 1.84%; sulphur, 0.92% ; nitrogen, 0.74%; ash, 13.24%. Recal- culate these values for an ash-free coal. 5. Determine by means of Dulong's formula the upper and lower calorific values of the coal described in Prob. 4. 6. The ultimate analysis of a sample of crude petroleum from which all water was removed gave the following results: carbon, 85%; hydrogen, 13%; sulphur, 1.0%; oxygen, 0.25%; nitrogen, 0.12%; ash (sand and similar material), 0.63%. Determine the upper and lower calorific values by means of Dulong's formula. CHAPTER XVII STEAM BOILERS 142. Definitions and Classification. The term boiler is generally applied to the combination of a furnace in which fuel may be burned continuously and a closed vessel in which steam is generated from water by the heat liberated within the furnace. Boilers are classified in many different ways, the more important being given in the following schedule: Classification of Boilers (1) According to form "(a) Plain cylindrical, (6) Flue, (c) Tubular, (d) Sectional, etc. (2) According to location of J (a) Externally fired, and furnace [ (b) Internally fired. (3) According to use (4) According to direction of principal axis (a) Stationary, (6) Portable (as on trucks, or rollers), (c) Locomotive, (d) Marine. (a) Horizontal, (6) Inclined, (c) Vertical. (5) According to relative posi- f tions of water and hot I gases I 326 (a) Water tube, (b) Fire tube. STEAM BOILERS 327 Examples of boilers of the different types mentioned are given in subsequent paragraphs, 143. Functions of Parts. It has been shown that there are two essentially different parts in the apparatus commonly known as a steam boiler, the furnace and the boiling vessel. A simple form of boiler known as a horizontal, return tubu- lar boiler, or an H.R.T. boiler, is shown in Figs. 193 and 194 with the two essential parts and their components Pressure Regulator M •. • .'— j Steam Dome cr ": ~:~- Fig. 193.— Sectional Elevation of H.R.T. Boiler and Furnace. indicated. The furnace consists essentially of the combina- tion of grates, bridge wall, fire and ash doors, the ash pit and the space above the grates. It is the function of the furnace to so burn the fuel that the maximum amount of heat will be made available for absorption by the water within the boiling vessel. It is the function of the boiling vessel to transmit to the water within it the greatest possible quantity of the heat thus made available and to resist successfully the 328 STEAM POWER zm A tendency to rupture under the action of the high internal pressure, that is, the pressure of the steam. In the type of boiler shown the fuel is " fired " by hand, that is, it is spread on the grate by being thrown from a scoop shovel through the opened fire door. Air enters through both doors in regulated pro- portions and in such quantities as best to approximate complete combustion. The hot gases re- sulting from the com- bustion pass over the bridge wall, along the Fig. 194. lower part of the boiler Section through Furnace of H.R.T. Boiler, shell and then through the fire tubes, or flues, toward the front of the boiler as shown by arrows in the figure. From the front end of the tubes the products of combustion pass up through the smoke box to " breechings " or " flues," which carry them to the stack. Heat is received by the water within the vessel in two different ways: (1) The hot fuel bed on the grate radiates energy in the same way that the sun or any other glowing body radiates energy. Some of this energy traverses the space between fuel bed and boiler shell and ultimately passes through that shell to the water within. The rest of the radiated energy passes into the walls surrounding the furnace and heats them and the surrounding atmosphere. (2) The hot gases of combustion pass over the heating surface of the boiler, as shown, and transmit part of their STEAM BOILERS 329 heat to the water on the other side of those surfaces. The rest of the heat which they carry is either lost to the surround- ing walls or is carried up the stack by the gases which leave the boiler at a comparatively high temperature. This temperature ordinarily ranges from about 500° to 700° F. and in extreme cases goes even higher. 144. Furnaces and Combustion. In most forms of boiler the water within the boiler has practically the same temperature as the steam being generated, and this is generally from 320° to 400° F. Obviously the products of combustion cannot be cooled by the water to a tem- perature below that of the water, so that the gases leaving the boiler in an ideal case would have a comparatively high temperature. Practically, it is found undesirable to attempt to reduce the temperature of the gases to a value even approximating that of the water and, as indicated above, they are discharged at a temperature several hundred degrees higher. In order that the maximum amount of heat may be made available for the boiling vessel the prod- ucts of combustion must therefore leave the furnace with the highest possible temperature, and the ideal furnace would completely burn the cheapest fuel available in such a way as to give this highest possible temperature and not to generate smoke. Real furnaces fall far short of this ideal performance, for numerous reasons. The more important of these are given in the following paragraphs: (a) Incomplete Combustion of Carbon. In a real furnace the combustion of the carbon of the fuel may be incomplete in two senses; first, some of the carbon may remain entirely unoxidized and pass off with the ash, and second, some of the carbon may be burned to CO instead of to CO2. Imperfect combustion of the first kind can result from fuel falling through the openings in the grate before it has been ignited or when only partly burned, or it can result from failure to get air to some of the carbon in sufficient quantities to burn it completely before all of the surround- 330 STEAM POWER ing fuel has been converted into ash and the locality cooled down to such an extent as to allow the unburned carbon in its midst to cool below the temperature of ignition. Imperfect combustion of the second kind, resulting in the formation of CO, generally results from a lack of suffi- cient air or from imperfect mixing of air and fuel or of air, fuel and products of combustion. It can also be caused by chilling of air or gases rising from the fuel bed before com- bustion has been completed. In any coal-burning furnace there is a tendency toward the formation of CO within the bed of fuel, and this tendency is greater the greater the depth of the bed. Such CO burns to CO2 above the fuel bed when conditions are propitious, the necessary air either passing through the fuel or being admitted at a point above the fuel. It is obvious that if combustion of such gas is to occur in this way, the combustible gas and the air must be brought into intimate contact while they are still at a sufficiently high temperature. The percentage of CO2 in the flue gases is commonly taken as an indication of the character of combustion. With pure carbon and theoretical air supply, this should be 21 per cent, and with real coal something between 18 and 20 per cent. Practically, it is so difficult to bring the combustible material and the oxygen of the air into inti- mate contact that a large excess of air is always used. The excess coefficient in practice varies from about 1.1 to over 2 with averages of about 1.5 under ordinary good conditions. The value 1.5 corresponds roughly to about 13 per cent CO2 in the flue gases and an excess coefficient of 2 cor- responds roughly to 9 to 10 per cent CO2. Even with such excess coefficients as those indicated as averages, it is not at all uncommon to find a small quantity of CO in the flue gases. The fireman's task with respect to the combustion of carbon thus reduces itself to the use of such a quantity of air in such a way that the minimum loss results from excess air arid Unburned CO. STEAM BOILERS 331 (b) Incomplete Combustion of Hydrocarbons. The hydro- carbons which appear as volatile matter in the proximate analysis are practically all distilled from the fuel, as it is heated in the furnace before ignition in the same way as when making a proximate analysis. If they are to be completely burned they must be mixed with the requisite quantity of air after distillation and both the vapors and the air must be maintained at a sufficiently high temperature until combustion is complete. The air for the combustion of distilled volatile may all filter through the fuel bed or some of it may be admitted at a point above the level of the fuel bed. If the flame formed by burning hydrocarbons is allowed to come in contact with cold surfaces, as, for instance, the heating surfaces of the boiler, the gases are cooled below the temperature of ignition and combustion ceases. This results in the deposit of soot (unburned carbon) upon the heating surfaces of the boiler and in the carrying of soot and unburned hydrocarbons up the stack. The soot and some of these hydrocarbons form the unsightly smoke so familiarly associated with some stacks. Or, if the air supply is at a sufficiently high temperature, but is insufficient in quantity, the hydrocarbons are in- completely burned and smoke results. The formation of smoke can be conveniently studied by means of the ordinary kerosene lamp. Such a lamp operates by burning hydrocarbons of the same general character as those distilled from solid fuels. The hydro- carbons are drawn up by the wick in the form of liquids, are vaporized by heat near the top of the wick and then combine with oxygen from the atmosphere to give the luminous kerosene flame. If the flow of kerosene and the air supply are properly adjusted and if the temperature is high enough, the com- bustion results in the formation of invisible and practically odorless gases. If, however, the air supply be decreased 332 STEAM POWER or be greatly cooled, a very smoky and very odorous combus- tion ensues. The same result could be obtained by the use of too great a quantity of air, a condition often attained when the supply of kerosene in the bowl of the lamp is almost exhausted. The effect of a cold surface is easily seen by inserting a cold metallic or porcelain surface into the tip of the flame and then withdrawing it. It will be found covered with soot. (c) Advantages and Disadvantages of Excess Air. It has been shown that excess air is practically necessary in the real furnace in order to insure against a deficiency at any point, and it is thus advantageous in that it makes the combustion more nearly complete than would otherwise be the case. On the other hand, excess air represents just so much excess material to be heated at the expense of heat liberated by combustion and hence decreases the maximum temperature attained. A sufficiently great supply of excess air could so reduce the temperature that even if combus- tion were complete very little heat would be made available for absorption by the boiling vessel, because the temperature attained by the products of combustion would be too low. Excess air in large quantities may also result in cooling unburned gases before combustion to such an extent as to make the completion of combustion impossible. 145. Hand Firing. The commonest type of furnace is that shown in Figs. 193 and 194, and the commonest method of hand firing consists in spreading a layer of fuel as evenly as possible over the entire surface of the fuel bed as often as required to replace the fuel burned away. At such inter- vals as experience shows to be necessary the fire is cleaned, that is, the ashes are worked out from under the fuel by means of slice bars, so that practically nothing but live fuel resting on a thin layer of ash remains behind. This method is open to many serious objections; the more important are ; STEAM BOILERS 333 1. There is a gradual increase in thickness of fuel bed from the time of one cleaning until the time of the next. This gives a constantly changing set of requirements for the proper proportions of air entering below and above the fuel bed and a constantly changing resistance to flow of air through the bed, so that great skill is necessary if the best conditions are to be maintained throughout. 2. There is always a tendency for a fuel bed to burn faster at some points than at others, due to the accidental distribution of fuel, ash and air. Where " holes " are formed in this way large quantities of comparatively cold air can pass through with the consequences already enumer- ated. It takes considerable skill and watchfulness on the part of the fireman to prevent the formation and continued existence of such holes. 3. The firing door must be opened wide every time that fuel is to be fired, that is, at intervals varying from two or three minutes to fifteen or more, depending on load, character of fuel, etc. While the door is open large quanti- ties of cold air readily flow into the furnace and cool down all parts of it, and a proportionately smaller amount will ordinarily pass through the fuel bed. The result of this on the flue gases and operation of the boilers has already been considered, but there is another result of equal or greater importance. As a consequence of this action the volatile hydrocarbons distilled off from the freshly fired fuel, which are themselves at a comparatively low temperature, are surrounded on all sides by cooled walls and come in contact with cold air only. The chances of their burning completely are very slight, and a great part of these volatilized materials passes off unburned as invisible gas and as smoke. Ob- viously the greater the volatile content the greater the dif- ficulty, so that anthracite causes least trouble in this way, while most bituminous coals give heavy black smoke when burned under these conditions. The cooling down of the interior of the furnace during 334 STEAM POWER firing is accompanied by the covering of the fuel bed with cold fuel, so that, for the time being, very little radiant heat enters the boiling vessel, and the gases which come in contact with its surface are comparatively cool. The maintenance of a constant steam pressure under these con- ditions is practically impossible, but the difficulties can be partly overcome by very frequent firing of small quantities, so that the door is open a very short time and also that the layer of fuel is very thin and does not cut off much heat. 4. The cleaning of the fire necessitates keeping the fire door open for several minutes, with results of the same variety as those just enumerated. Summing up these difficulties, they divide themselves into two classes — those which can be almost or entirely eliminated by skill of a very high order and those which are inherent and cannot be eliminated by skill. It will also be observed that all should give more trouble with fuels high in volatile than with those of the anthracite variety, both as to incomplete combustion and to the formation of smoke. Several other methods of hand firing have been proposed, particularly for use with bituminous coals, and some of them have been successfully utilized in isolated instances. Nearly all depend upon covering only part of the fuel bed at one time and, by alternating the parts covered in this way, fresh fuel on one part of the bed is coked while air is heated by coming in contact with the uncovered incandescent part of the bed and is therefore in proper condition to burn more perfectly the volume of hydrocarbons being distilled off. These methods are all good, but they involve a great deal of careful work and a high degree of skill on the part of the fireman. Other methods of eliminating some of the difficulties depend upon modifications of the furnace and air supply. Most attempt to entirely surround the fuel and the gases given off with heavy masses of brick work and tile, so that STEAM BOILERS 335 enough heat will be stored during incandescent periods to tide over the periods of cooling. Some forms have combined with this idea a series of air ducts in the brick work so arranged that air on its way to the furnace passes through these ducts and is heated. In some cases the air supply is automatically controlled and more air is supplied above the fire during the period of distillation, or coking, as it is called, than during the following period, when the coked coal is brightly incandescent and little volatile matter is present. Fig. 195. In some hand-fired furnaces which are intended for use with bituminous coals that give a long flame the parts of the boiling vessel within range of the flames are covered with tiles. This prevents impingement of unburned gases upon cool surfaces and thus tends to prevent the formation of smoke and incomplete combustion. Carrying this principle to its logical conclusion results in the installation of the grate in a firebrick chamber in front of the boiler-setting proper, as shown in Fig. 195. Such a device is known as a Dutch oven and is often very efficient in totally or partially preventing the formation of 336 STEAM POWEE smoke. It does not, however, give as high an economy as might be expected, because a great part of the radiant heat of the fire does not reach the boiler surfaces and because the large external surface results in great radiation losses to atmosphere. Another interesting modification consists of reversing the direction of the draft, that is, the direction in which the air passes through the fuel bed. The type of furnace al- ready described is known as an updraft furnace, because the air passes upward in flowing through the bed. The modi- fied type here referred to is called a downdraft furnace, because the air flows downward in passing through the fuel. In downdraft furnaces the coal is fired on top of the grate as in other types, but the air is admitted above, flows downward toward what would normally be the ashpit, and from there on over the heating surfaces of the boiler. Fresh coal fired on top of the incandescent bed in such a furnace distills as in other types, but the volatiles are mixed with the entering air and are carried downward through the hot bed so that ideal conditions for combustion are more nearly attained. In some forms there is a second updraft grate beneath the downdraft grate. This second grate receives partly burned coals falling through from the upper grate and holds them until combustion is practically com- pleted. In downdraft furnaces the grate bars are generally made of pipes, and water, from the boiler or on its way to the boiler, is circulated through them. If this were not done the grates would quickly warp out of shape and ultimately burn away because of the high temperatures to which they are subjected. 146. Mechanical Grates. In order to overcome the dif- ficulties arising from opening the doors for the purposes of cleaning the fire, numerous so-called rocking, shaking, self-cleaning, or dumping grates have been developed. These are generally built up of grate bars which have a STEAM BOILERS 337 rough T or an inverted L section with the upper horizontal branch of the T or inverted L slightly rounded,- as shown in Fig. 196. These bars are arranged in groups with their longitudinal axes running across the grate, and they are so supported that they can be rocked about a point in the verti- cal leg of the T or L by means of levers located at the front of the boiler. By rocking the bars the lower part of the fuel bed which has been burned to ash can be dropped into the ash pit, while the upper part is sufficiently agitated to close up holes which may have formed, and this can all be done Fig. 196. with the doors closed. Or, if desired, part or all of the fuel bed can be dropped into the ash pit by a similar rocking motion. 147. Smoke and Its Prevention. An idea of the reasons for the formation of smoke will have been obtained from the preceding paragraphs. A reasonably skillful fireman should have little difficulty in burning anthracite coals in the simpler forms of furnaces without smoke, but it is almost impossible to commercially burn many of the varieties of bituminous coals in this way without the formation of excessive volumes of dense black smoke at intervals immediately following each firing. 338 STEAM POWER Aside from all aesthetic and sanitary considerations, smoke is undesirable because it represents poor furnace conditions and waste. The actual loss of carbon in visible smoke is generally almost negligible in comparison with the other losses in the form of unburned hydrocarbons, the lowered initial temperature, etc. All of these losses combined represent a waste of considerable magnitude. The proper method of smoke elimination is not the combustion or removal of smoke already formed, but it is the burning of fuels in such ways as not to form any appreciable quantity in the first place. To accomplish this end the following must be achieved: 1. Coal must be fired continuously and uniformly without the opening of doors which admit cold air to the furnace. 2. Volatiles must be distilled continuously and uni- formly and in such a place that they are given ample oppor- tunity to mix with proper proportions of air and to burn completely before coming in contact with cool surfaces. 3. The air supply must be properly controlled and tempered to meet the demands of the fuel both in and above the bed. 4. The fire bed must be worked continuously and uniformly so as to eliminate ashes as rapidly as formed and to maintain a bed of uniform depth and condition. Some of these necessary conditions can be attained by the use of the various forms of hand-fired furnaces already described but, even in the hands of skillful and industrious men, it is impossible to meet all of them. Mechanical stokers which more nearly approach the ideals set have therefore been developed and are widely used. 148. Mechanical Stokers. These mechanical devices are useful for two reasons — they eliminate a great deal of labor and they make possible the burning of many varieties of refractory fuels without the formation of excessive quanti- ties of smoke. STEAM BOILERS 339 Despite the good results which can be achieved by their use, mechanical stokers are not installed in small plants as often as might be expected. This is because good stokers are very expensive in comparison with hand-fired furnaces and, despite economy of fuel, do not generally show a finan- cial saving unless their use eliminates the services of several firemen. It is generally assumed that one man can care for water, coal and ashes for about 200 boiler horse-power or can handle coal only for about 500 boiler horse-power. Experi- ence has shown that one man can fire about 2000 to 5000 boiler horse-power when the boilers are equipped with good stokers and coal-handling apparatus. Financial calculations will not justify mechanical stokers in many of the smaller plants in which they are used. How- ever when improved conditions with respect to smoke and dirt and improved labor conditions are taken into account they are generally regarded as paying propositions. Mechanical stokers can be roughly divided into two types, those which duplicate hand spreading of fuel and are known as sprinkler stokers, and those which supply fuel at one or more points and work it progressively toward the ash end of the apparatus as it burns. The first type has not been widely installed, though it is possible that it may meet with more popular approval after further development. Stokers of the second type may be roughly divided into four classes, which are 1. Chain grates. 2. Inclined stokers or overfeed stokers. 3. Underfeed stokers. 4. Combinations of above. A chain grate, as made by the Illinois Stoker Company, is illustrated in Figs. 197, 198, 199, and 200. It consists 340 STEAM POWER O STEAM BOILERS 341 of a broad chain made up of a great number of small links and carried on toothed wheels and roller wheels supported in a frame which can be wheeled into position within the Fig. 198. — Sprocket and Links of Illinois Chain Grite. TOP VIEW OF CHAIN SHOWING DISTRIBUTION OF AIR SPACES BOTTOM VIEW OF CHAIN SHOWING ROLLERS FOR DRIVING-SPROCKET ENGAGEMENT Fig. 199. boiler setting. The general arrangement of the chain and rollers is shown in Fig. 197; details of the front or driving rollers and of the links are shown in Fig. 198; a top and bot- tom view of part of the chain is given in Fig. 199; and Fig, 200 is a perspective view of the frame showing the 342 STEAM POWER tracks on which it may be rolled into and out of the boiler setting. The chain is driven slowly in the direction indicated by the arrows in Fig. 197 by power applied, through worm gearing, to the shaft of the toothed wheels at the front of the stoker. Coal feeds automatically from the hopper by gravity and is carried into the combustion space by the moving chain, the thickness of the bed being controlled Fig. 200.— Framework of Illinois Chain Grate. by the height of the adjustable gate shown. As the fuel enters the furnace it passes under the coking arch, which spans the entire front part of the grate and which is main- tained at a high temperature by heat radiated from the in- candescent fuel nearer the inner end of the grate. The volatiles are distilled from the fresh coal by heat received from this arch and are heated and mixed with air at this point. The coked fuel is then carried on into the furnace and burned, the refuse being discharged at the bridge wall. If the thickness of bed and speed of chain travel are properly adjusted, all of the fuel can be coked before pass- ing out from under the arch and can be burned almost STEAM BOILERS 343 completely before reaching the bridge wall, so that prac- tically ashes only will be discharged. The apron shown at A in Fig. 197 is used to prevent the free passage of air to the part of the chain carrying 344 STEAM POWER practically nothing but ash, as this would result in excessive dilution of the products of combustion. A stoker of this type installed under a horizontal return- tubular boiler is shown in Fig. 201. In the illustration part of the side frame of the stoker is broken away in order to show the chain and its roller guides. The eccentric shown near the top of the front of the boiler drives the chain through an arm of adjustable length, which makes possible the control of the speed of chain travel. The early forms of chain grates were intended for use with natural draft, that is the space under the grate con- nected directly with the atmosphere and maintenance of a " draft " or " under-pressure " in the furnace permitted the external atmosphere to push through the fuel bed, this supplying the necessary oxygen for combustion. As the art developed a demand arose for stokers capable of burn- ing more fuel per square foot of grate area. This demand was met by developing stokers which could be used with " forced draft." These stokers were so arranged that air could be forced through the fuel bed by fans discharging into the space beneath the stoker. The better designs of forced draft, chain grates provide for differential air supply to different sections of the fuel bed. For this purpose the space between the upper and lower chains is divided into compartments by vertical partitions in planes at right angles to the travel of the grate. Provision is then made for controlling the air pressure in each box separately by means of dampers or equivalent. With such construction the air pressure applied to the under surface of the grate can be regulated by sections instead of for the entire grate and it becomes possible to grade the air supply to more nearly meet the requirements of the fuel in different stages of combustion. As an example of the usefulness of such an arrangement, consider the last section over which the material passes before being discharged from the end of the stoker. If STEAM BOILERS 345 the fuel is very nearly burned out before it reaches this section a very small amount of air will suffice for the com- pletion of combustion. Any greater amount would simply pass through without useful effect, would serve to lower the furnace temperature and to increase the excess coeffi- cient. On the other hand if the fuel still contains large amounts of unburned carbon when it reaches this section it is desirable to be able to supply a relatively large amount of air so as to more nearly approximate complete com- bustion of the carbon before the material is discharged from the grate. Development of the forced draft chain grate and of numerous small but important refinements in details of design has brought the chain grate into particular promi- nence in connection with the combustion of certain very poor grades of fuel. Thus very fine sizes of anthracite previously considered of no commercial value because of their small size and high ash and moisture content are successfully burned on chain grates particularly designed for such use. Some of the poorer varieties of bituminous fuel with high ash content of sux?h character as to form excessive amounts of clinker, are also being burned on such devices. An inclined stoker with front feed and a step grate, known as the Roney stoker, is shown in Figs. 202 and 203. The fuel is fed out of the hopper and onto the dead plate by means of the reciprocating pusher. From the dead plate it is pushed down upon the grate bars by the follow- ing fuel. These bars are rocked mechanically so that their tops alternately assume horizontal and inclined positions, and this action feeds the fuel downward until it is dis- charged onto the dumping grate. The material collect- ing on this grate is periodically dropped by hand into the ashpit. The fuel is coked while passing under the coking arch and the coked material is practically completely burned by the time it has traveled down the grate, The volatile^ 346 STEAM POWER are mixed under the coking arch with heated air which has passed through the grate and with heated air forced in above the fuel. An inclined stoker of the side-fees type with bar grates known as the Murphy stoker, is illustrated in Figs. 204, 205 and 206, This stoker is provided with two coal- Sheath Agitator Lock-Nut Connecting-Rod Fig. 202. — Details of Feed Mechanism, Roney Stoker. magazines or hoppers which are placed horizontally in the side walls of the boiler setting and feed fuel onto the inclined grate bars, Fig. 204, which carry it downward toward the lower point of the V formed by the grates. The grate bars, Fig. 206, are alternately fixed and mov- able, the movable bars being hung from above and their lower ends being moved up and down by power furnished by a small steam engine or other convenient source. STEAM BOILERS 347 m 3 348 STEAM POWER A toothed bar arranged for rotation by hand or by power is located at the bottom of the V and is used for grinding up ash and clinker which is too large to fall through into the ash pit. This bar is kept cool by making it hollow and connecting one end to the smoke flues or stack so that air is constantly drawn through it. iP£ ^ £/&■;, Fig. 204. — Transverse Section of the Murphy Stoker. The location of the coking arch and the method used for supplying warm air should be evident from the figures. A stoker of this variety is shown in place under a hori- zontal water-tube boiler in Fig. 207. An underfeed stoker made by the Combustion Engineer- ing Company is shown in Figs. 208 and 209. Coal is fed from the hopper onto the reciprocating bottom B by means of the reciprocating pusher P. The part B forms the bottom of a trough as shown in Fig. 209, and its reciprocating motion feeds the coal upward and out of this trough so that STEAM BOILERS 349 it spills over onto the inclined grate bars. The reciprocating motions are all obtained from the direct-acting steam cylin- der shown. The inclined grate bars are alternately fixed and mov- able, the movable bars sliding back and forth at right angles W^ZZ ZW A.-rL*.- ■»: J Fig. 205. — Longitudinal Section, Murphy Stoker. to the trough under the action of horizontal rocking bars R. This action gradually feeds the fuel downward and toward the side of the furnace, the refuse finally landing on the dumping trays shown. Air enters the duct below the trough through the adjustable gate G, controlled by crank C, and part of it passes out through holes H hear the top of the trough, 350 STEAM POWER Fig. 209. The remainder passes down through the hollow grate bars and into the heated air box from which it flows upward between the grate bars. It will be observed that the coal is fed onto the grate from below, so that all volatiles distilled off must pass up- ward through the incandescent fuel before entering the space above the fuel bed. Part of the air which is to burn Stationary Grate Bar Movable Grate Bar Fig. 206.— Grate Bars of Murphy Stoker. this volatile matter also passes through the fuel bed and the remainder flows over the incandescent fuel from the opening shown near the hopper in Fig. 208. The air and volatiles are thus raised to a high temperature and well mixed, and the operation is continuous and uniform, all tending to facilitate smokeless combustion. Another variety of underfeed stoker known as the Taylor stoker is shown in Fig. 210 (a), (b) and (c). This stoker is built up of alternated retorts and air STEAM BOILERS 351 boxes, the proper number to give the desired width of stoker being used. Coal is fed from the hoppers W into the retorts by the upper ram or plunger shown in Fig. 210 (b) and part of it is again pushed forward by 352 STEAM POWER Steam Cylinder Fig. 208.— Longitudinal Section of Type "E" Stoker Fig. 209.— Cross Section of Type "E" Stoker. STEAM BOILERS 353 354 STEAM POWER the lower ram or plunger. The stroke of the lower plunger can be regulated and in this way the relative quantities of coal pushed forward in the upper and lower parts of the retorts can be controlled. The coal spreads over the tuyere blocks which form the inclined tops of the air boxes and forms a comparatively even, inclined layer of fuel. Coking proceeds under the incandescent fuel which forms the upper surface of this layer, and the volatiles mix with air entering through the hot tuyeres and pass upward through the hot fuel above. In this stoker advantage is often taken of the fact that the draft (pressure of air) required with underfeed stokers is so great that it can be more economically attained by the use of a fan than by the use of a stack. The fan and the coal-feeding plungers are both connected to one engine and the speed of this engine is automatically controlled by the steam pressure within the boiler. As this pressure decreases the engine speeds up, thus delivering more coal and air and as the pressure increases the engine slows down with opposite results. By properly fixing the travel of the plungers initially, the best relative proportions of air and coal are set for the entire range of loads to be carried and the variation of both is thereafter in approximately the same proportions. A stoker of this type in position under a horizontal water-tube boiler is shown in Fig. 211. A double-ended arrangement of Taylor stokers as used under very large water-tube boilers is shown in Fig. 212. Powdered or pulverized coal burning equipments have been invented in great numbers and are successfully used in several of the industries. They are essentially stokers. Many attempts to use pulverized coal for firing boilers have been made from time to time but until quite recently the results were not considered satisfactory. Within the past few years a number of installations have been made in STEAM BOILERS 355 boiler plants and the results are claimed to be satisfactory in many instances. The coal after crushing to moderate size is dried by passing it through a drier heated by hot products of com- Fig. 211. — Taylor Stoker Under Horizontal Water-tube Boiler. bustion unless the moisture content " as received " is not considered too great for satisfactory pulverization and utiliz- ation. The material is then pulverized in a mill arranged to give very fine subdivision, pulverizing so that at least 60 to 90 per cent passes through a 200 mesh sieve being common practice. 356 STEAM POWER Fig. 212. — Double-ended Arrangement of Taylor Stoker under Sterling Type W. Boiler. STEAM BOILERS 357 The pulverized fuel is then transported to small hop- pers immediately adjacent to the individual boilers. From these hoppers it is fed into a stream of air in which it becomes suspended and by which it is carried through the burner into the furnace. The burner generally consists of a simple nozzle. The fuel burns in the form of a torch at the end of the nozzle, combining with the air which carried it in and generally with additional air admitted through controllable doors or openings in the walls of the furnace. The fine pulverization makes possible very rapid com- bustion of the individual particles and the flames have the appearance of an intensely hot gas flame or liquid fuel flame. Success depends on completing combustion while the particles of fuel are still suspended in space, that is before they have had time to settle or to impinge on any solid surface. The flame is easily brought to so high a temperature that the ash contained in the fuel is liquefied. Unless the installation is carefully planned and properly operated this fused ash is very apt to cause serious difficulties by building up stalagmites and stalactites in the furnace or by actually solidifying on the cold heating surfaces of the boiler. Many advantages are claimed for pulverized fuel firing in boiler plants. The principal claims are great flexibility with respect to character and quality of fuel burned, ease of operation so that one fireman can handle a large boiler installation, ease and simplicity of regulation so that high thermal efficiency can be obtained and maintained even under rapidly changing conditions, and elimination of banking losses as fuel is entirely excluded from the furnace when steam is not desired. However, the use of pulverized fuel entails a large investment in equipment for preparing it and the fixed charges on this equipment as well as the operating expenses chargeable to prepara- tion tend to balance gains resulting from the advantages 358 STEAM POWER enumerated above. The art is as yet so young that it is impossible to obtain sufficient data to make a true and complete comparison between pulverized and solid fuel firing. Oil firing is essentially a mechanical, rather than a manual process, and while oil burners are not ordinarily understood as belonging to the class of mechanical stokers, they have all the essential characteristics of such ap- paratus. To burn oil successfully under a boiler it must be finely atomized and mixed with the necessary quantity of air, and there must be sufficient open space within the furnace for the free development of the flame and the completion of combustion before impingement on cool surfaces. Oil-burning furnaces are generally given a rather large volume; considerable firebrick is used in such ways as to give incandescent walls and baffles to assist ignition and combustion, and all heating surfaces are arranged so that they are not in the direct path of the flame. The atomization of the oil is effected in two distinctly different ways. In some forms of burners it is brought about by mechanical means, the oil being pumped through a nozzle of some sort which is so shaped that the issuing jet breaks up into a great number of very small particles. In other forms, steam is used to break up the jet, the steam and oil entering the body of the burner separately and later coming into contact in such a way that the oil is literally torn apart by the steam. This form of burner has been more extensively used in the United States than has the former but present developments indicate a probable reversal in this respect. The mechanical type of burner has been developed to a very high degree during the past few years, and one form, at least, seems to possess marked advantages. Oil burning shares with the burning of powdered coal, STEAM BOILERS 350 the property of permitting very accurate regulation of the air supply to suit the quantity of fuel being burned. The excess coefficient may therefore be maintained at a low value and the initial temperature may be made cor- respondingly high. Part of the advantage thus gained over the commoner methods of coal firing is, however, counterbalanced by the quantity of steam used for heat- ing and pumping the oil and for atomizing in some forms of burners. Both oil burning and powdered-coal burning can be easily made to give smokeless combustion in properly designed furnaces and both yield readily to forcing. That is, the temporary consumption of excessive quantities of fuel to tide over short demands for excessive amounts of steam is comparatively easily effected if sufficient furnace volume is available. 149. Rate of Combustion. The rate at which coal is burned in a given furnace or on a certain grate is generally given in terms of pounds of coal fired per square foot of grate surface per hour and is referred to as the rate of combus- tion. The rate at which coal can be consumed is largely dependent on the intensity of draft available, that is, on the air pressure available for driving air through and over the bed of fuel. The higher the pressure available, the greater will be the quantity of air which can be sup- plied and the greater will be the quantity of coal that can be burned. If it were not for the cost of creating the draft, the only limit to increasing the rate of com- bustion would occur when the velocity of the air became so great that the fuel would be picked up from the grate and carried onward into the flues in a partly burned con- dition. Commercial drafts give pressure differences above and below the fuel bed which range from about 0.1 inch of water to as high as 8 ins. In stationary plants the pressures generally range from 0.1 to about 0.5 in cases 360 STE^M POWER where hand firing is employed, and are carried as high as 5 or more inehes of water with some forms of mechanical stokers. The best rate of combustion varies with the type and size of fuel, the type and size of furnace, the type and size of boiler, the draft and many other considerations. In ordinary power-plant practice with hand firing the rates of combustion commercially used generally fall within the following limits: with anthracite, 15 to 20 lbs. per square foot per hour; with semi-bituminous, 18 to 22 lbs.; and with bituminous, 24 to 32 lbs. In the case of stokers, these values may be doubled and even trebled if proper provisions are made. As practically all of the volatile is consumed above the grate, the fixed carbon content is practically the de- termining factor, since it is this constituent that is .burned on the grate. This explains the high rate possible with fuels with high volatile content. The most eco- nomical results are generally obtained when from 12 to 16 lbs. of fixed carbon are consumed per square foot of grate per hour. The figures given above do not represent limiting con- ditions. In torpedo-boat practice, where high-draft pres- sures are used (from 4 to 8 ins. of water), rates of from 50 to 120 lbs. are attained. On locomotives, which also use high-draft pressures, rates of combustion greatly in excess of stationary practice are generally used. The capacity of a given boiler, that is, its ability to generate steam, increases as the rate of combustion is increased, since more heat is thus made available. The economy of the combination, that is, pounds of steam generated per pound of coal fired, increases until some best rate of combustion for the fuel in question is reached, and thereafter decreases. The variation of economy is, however , not very great for a comparatively wide range of combustion on either side of the best rate. STEAM BOILERS 361 Curves giving approximate draft pressures required fee different rates of combustion when different kinds and sizes of fuel are hand fired are given in Fig. 213. The sizes referred to are explained in Tables XIII and XIV. Table XIII also shows the relative increase of ash content as the a? ° • / / / / t j/ 5 V 4 ^5^^ V V V ..^ I f^ ie — ^Ui &«»•-' 5 10 15 20 25 30 35 40 45 60 Pounds of Coal per Sq.. Ft. of Grate Surface per Hour Fig. 213. — Draft Required for Different Rates of Combustion with Different Sizes and Kinds of Fuel. size decreases, there being a tendency toward the concen- tration of the ash in the smaller sizes. 150. Strength and Safety of Boiler. Attention has already been called to the fact that the boiling vessel has to be designed with two different requirements in view: it must be mechanically strong to resist internal pressure and it must transmit the maximum amount of heat to the con- tained water. Spherical and cylindrical surfaces with the pressure act- ing on the inside of the curve are best adapted to resist such pressures, as they already have the shape which the pressure would tend to give them. Boilers are, therefore, 362 STEAM POWER TABLE XIII Sizes of Anthracite Coal (Sizes larger than pea coal generally too costly for power-plant use.) Name. Through Screen with Mesh. (Inclusive.) Over Screen with Mesh. (Inclusive.) Ash Content (Average). Run of mine Broken. unscreened unscreened 21 2 li 3 4 1 2 1 4 1 8 Egg Stove Chestnut 2| 2 li 3 4 1 2 1 4 6 10 13 Pea Buckwheat No. 1 Buckwheat No. 2 or rice. . . 15 17 18 TABLE XIV Sizes of Bituminous Coal (Considerable variation in commercial practice in £.~~~ng and sizing.) Name. Through Bars Spaced Apart. (Inches.) Over Bars Spaced Apart. (Inches.) Lump li Nut Slack li 4 4 constructed as far as possible of vessels having only spherical and cylindrical surfaces. "Flat surfaces which are poorly adapted to resist such pressures as act within a boiler must often be used despite their weakness. When incorporated in a boiler they are invariably " stayed," that is, braced by being fastened to other surfaces by stay bolts and other forms of fastenings. Examples will be given later. Most of the early designs of boilers and many of the modern types consist of large cylindrical vessels made by riveting together properly shaped steel plates. These shells are often traversed from end to end by flues or STEAM BOILERS 363 tubes for carrying hot gases and generally have flat ends more or less perfectly braced by these tubes and by long tie rods and other braces. Such boilers when in operation Fig. 214. — Lap Joint. Fig. 2 15.— Butt Strap Joint. are almost entirely filled with water and often hold many tons. Boilers of these types have been responsible for many disastrous boiler explosions, and this fact has led inventors to the development of models which should be less dangerous. It seems practically impossible to develop a commercial boiler which cannot be made to explode to a certain extent Fig. 216.— Riveted Plates of Boiler Shell. Fig. 217. if sufficiently mistreated and mishandled, but much can be done to minimize the danger. The great weakness of the older forms lies in the riveted joints, which can never be made as strong as the plates which they fasten together. Two types of joint are in use; they are known respectively as the lap joint and the butt strap joint. These are shown in Figs. 214, 215 and 216. So far as a circumferential seam, that is, one running around the cylinder as shown in Fig. 216, is concerned, the lap joint 364 STEAM POWER is perfectly satisfactory and is universally used. With longitudinal seams, however, this is not the case. A lap joint throws the joined edges out of a true cylindrical sur- face as shown in Fig. 217, and when the vessel is subjected to pressure there will be a tendency for the plates to assume a cylindrical contour as nearly as possible. This causes local bending of the plates on each side of the lines of rivets, and the continued repetition of this action ultimately causes failure. The conditions are often made still worse by calk- ing the joint on a line indicated by a in Fig. 217, that is, by hammering the metal at the inner surface of the edge of the outer plate into firmer contact with the outer sur- face of the inner plate for the purpose of making a tight joint. The butt-strap joint can obviously be made so that the joined plates more nearly form a true cylindrical sur- face. Other weaknesses of the older forms lie in the flat surfaces used; in constructions which render it possible for sediment to collect on heated surfaces and thus permit local over- heating of the plate; and, above all, in the very large quantity of water contained. The disastrous consequences of boiler explosions are generally due to the action of the hot water contained within the boiler and not to the steam contained at the time rupture occurs. The water within the boiler is under steam pressure and approximately at steam temperature. Removal of the pressure by rupture of the container would enable a great part of this water to flash suddenly into steam at the expense of its own heat, and this is exactly what occurs in the case of a boiler explosion. Local failure causes a sudden lowering of pressure, and this results in the formation of large volumes of steam which, blowing out through the initial fracture, tend to enlarge it, to move the boiler and surroundings, and, in general, to do all possible to further the rupture and make conditions worse. STEAM BOILEES 365 From the preceding discussion the requirements for maximum safety can be deduced. They are: 1. The smallest convenient diameter of cylindrical ves- sels, so as to decrease the total load on joints for any given steam pressure. 2. The elimination of the greatest possible number of riveted joints and the use of butt-strap longitudinal joints on all large-diameter, cylindrical vessels. 3. The substitution of curved surfaces for all flat stayed surfaces. 4. So shaping the boiler that the required extent of heating surface may be obtained without enclosing a great volume to be filled with hot water when the boiler is steaming. 5. So shaping the boiler that such water as is contained therein will be divided up into small masses contained within separate vessels connected in such a way that rapid flow of all water toward one point of failure is impossible. 6. So shaping the boiler that no riveted joints shall be in the paths of flames and that no sediment can collect on metal immediately over flames or exposed to very hot gases. 7. So shaping the boiler that it shall be free to expand and contract with changes of temperature, with the least resultant strain on the different parts. These various requirements are most nearly met in the different forms of water- tube boilers, some of which will be de- scribed in succeeding paragraphs. 151. Circulation in Boilers. If a flask of water, such as that shown in Fig. 218, */ G : 218 '~ i? ir( : u ~ . , . , . ,, . ,. . , ,' lation in a r lask. be heated in the manner indicated, the water will gradually acquire motion and follow paths such as those shown by the arrows in the illustration. The heated water will rise in the center of the mass and the cooler water will flow downward around the outer surface. 366 STEAM POWER Such motion is called circulation. Rapid circulation within a boiler is very desirable, since it brings the maximum quan- tity of water in contact with the heating surfaces in a given time and hence tends to increase the amount of heat taken from those surfaces. It also tends to sweep along any bubbles of steam or gas formed on such surfaces and to carry away any sediment which may have collected, thus pre- venting overheating of the surfaces. Circulation can be expedited by providing free and unrestricted paths for the water so as to guide it in the proper directions and by applying the most intense heat at the proper point along the path of the water. The tem- perature of the water which is subjected to the most intense mLbhki^ ~m~- ^=^iM ^= & Blo-w oil' Fig. 219. — Elementary Types of Boilers. heat is naturally raised and the water at that point becomes less dense than in other parts of the boiler. The formation of steam at such points also materially lessens the density. As a result of this lowering of density the heated water rises and the cooler water descends to take its place. The more rapid this exchange can be made, the more steam can be generated from a given amount of surface in a given time and hence, other things equal, the better the boiler. The elements of two common forms of boiler are shown in Fig. 219, the arrows indicating the direction of the cir- culation and its effect upon the delivery of steam and of sediment. 152. Types of Boilers. In a book of this scope it would be impossible to describe all the types of boilers at present STEAM BOILERS 367 in use. The more important varieties have therefore been chosen for description and illustration. Two types of internally fired, tubular boilers more accurately described as internally fired, upright or vertical, fire-tube boilers are shown in Fig, 220. The furnace is 368 STEAM POWER Tube Sheet Steam Space Exposed Tubes Water Level Water Column and Try Cocks Feed Water Connection Pressure Gauge Hand Holes (Closed by hand hole covers when in operation) Fig. 221.— Large Internally Fired Tubular Boiler. STEAM BOILERS 369 contained within the shell of the boiler and is almost com- pletely surrounded with water. The heat radiated from the hot fuel is thus almost entirely received by the water of the boiler. The hot gases, rising from the fuel bed, pass upward through the tubes and, after giving up part of their heat to the surrounding metal, enter the smoke box and pass directly to the stack. Boilers of the type shown in Fig. 220, (a) and (b) are called exposed-tube boilers, because the water level is carried below the tops of the tubes. The tubes, therefore, extend through the steam space and act as imperfect superheaters. Boilers of the type shown in Fig. 220, (c), in which the tubes do not enter the steam space, but are entirely covered by water, are called submerged-tube boilers. Upright tubular boilers of the types shown in Fig. 220 are built by a number of manufacturers in sizes ranging from about 4 boiler horse-power to about 50 boiler horse- power. They are self contained, require no setting of any kind, and are shipped completely erected. Such boilers are very often mounted on trucks or skids and used to generate steam for small hoisting and other forms of con- tractors' engines. They are also used on steam fire engines. The pressure carried in these small tubular boilers is generally under 100 lbs. per square inch, but they can be built for higher pressures if desired. In Fig. 221 is shown a larger type of internally fired tubular boiler as made by the Bigelow Company for station- ary use. These boilers are similar to those just described, but are made only in large sizes, in this case, in sizes ranging from 40 boiler horse-power to 200 boiler horse-power. The exposed tubes generally give a superheat of about 25° F. These large upright boilers can be built to operate with a pressure as high as 200 lbs. per square inch and because of the small area covered by even the largest sizes, they are particularly adapted to locations in which floor space is limited. 370 STEAM POWER The locomotive type of boiler is shown in Fig. 222. It is an internally fired, horizontal, tubular or fire-tube, boiler. Such boilers are seldom used for stationary purposes, but are universally used on steam locomotives and, in the smaller sizes, are often mounted on trucks or skids and used for semi-stationary purposes by contractors and others. Boilers of this type are built in sizes ranging from 10 boiler horse-power or less up to over 100 boiler horse-power for general power purposes, while those used on the largest locomotives generate over 2000 boiler horse-power. The Continental type of boiler, named from the Con- tinental Iron Works, is shown in Fig. 223. These boilers Handholes Fig. 222. — Locomotive Type of Boiler. may be described as internally fired, return tubular, with semi-external combustion chamber, this chamber being out- side of the boiler shell proper but being built as an integral part of the boiler and transportable therewith. Boilers cf this type are built in sizes ranging from about 75 boiler horse-power to 300 or more. The grates, furnace and ash spaces, and bridge wall are all carried within circular, corrugated flues, one flue being used in the smaller sizes and two in the larger. The corru- gations serve the double purpose of strengthening the flue and of exposing added heating surface to fire and hot gases. The steam pipe shown just below the steam connection at the top of the boiler is commonly used on boilers for the STEAM BOILERS 371 paaiog noipog dox 372 STEAM POWER purpose of preventing the escape of excessive quantities of moisture with the steam. These boilers are very compact in shape and are short for their capacity, but they contain a great volume of water. They possess the advantages of having a large steam space and a very extended liberating surface over which the steam separates from the water. Uptake Tubes Fire Doors Man Hole- Fig. 224. — Scotch Marine Type Boiler. The Scotch marine type of boiler is shown in Fig. 224. It has the same general construction as that just described excepting that the combustion chamber is entirely enclosed within the water space of the boiler. This chamber is built up of flat plates and is held against collapse by numer- ous stay bolts. Boilers of this type were until recently the standard for marine practice, but they are now being replaced in many instances by water-tube boilers of more recent design. STEAM BOILERS 373 Scotcn marine boilers are very economical in the use of fuel, are good steamers, and are absolutely self contained. They are built in numerous sizes, the smallest having shells with diameters of about 6 ft., while the largest diameter used is about 16 ft. The largest boilers have three and four corrugated furnaces. Two types of externally fired, return-tubular (or " H.R.T.") boilers are shown in Figs. 225 and 226. The Fig. 225.— Horizontal Return-tubular Boiler with " Full Flush Front." only essential differences in these two types are in the forms of setting and in the methods of suspending the boilers. The shell is generally rigidly supported at the furnace end and arrangements made to allow for movement of the other end with changes of temperature. These boilers can be built very cheaply and are therefore widely used when their limitations, do not prevent. It has been found inadvisable to build them in sizes larger than 200 boiler horse-power or for pressures higher than 150 lbs. per square inch, and they are generally used in smaller 374 STEAM POWER STEAM BOILERS 375 376 STEAM POWER Fig. 228.— Forged Header for Bab- cock & Wilcox Boiler. sizes and with lower pressures. These limitations are set by permissible thickness of metal immediately above the fire, experience having shown that the plates deteriorate rapidly at this point if made too thick. One form of Babcock & Wilcox water- tube boiler is shown in Fig. 227. This boiler is built up of sections consisting of several tubes joined at the ends by headers, and the sections are connected side by side at each end to a long horizontal drum. The ends of this drum are closed with " dished " heads, thus doing away with flat surfaces and the necessity for stays within the drum. A detail of the forged header is shown in Fig. 228. It may be regarded as a long box of rectangular section with opposite walls pierced by circular holes, which has been so distorted as to give it a wavy shape. The distortion brings the holes into such positions that the tubes when expanded into these holes are " staggered," that is, do not lie one above the other. The general principle in- volved in the arrangement of these sections or elements and the resulting circulation are shown in Fig. 229. The location of the feed- water inlet and other details are shown in Fig. 230. It will be observed that the feed water enters in such a direction and position that it is readily picked up by the current of water circulating in the boiler, carried toward the rear and down the rear header. During this travel it is heated by contact with the hot water in Fig. 229. — Elementary Babcock & Wilcox Boiler, Showing Circulation. STEAM BOILERS 377 Hand hole opposite end of tube, closed by hand hole cover when in operation. End of tube expanded into bole of header. Fig. 230. — Details of Babcock & Wilcox Boiler Construction, 378 STEAM POWER the boiler and most of its impurities are separated out and settle in the mud drum at the bottom of the rear header. The boiler is suspended by stirrups from beams carried by the brickwork as shown in Fig. 227, the tube sections STEAM BOILERS 379 simply hanging from the drum by the nipples at each end. The various parts of the structure are thus free to expand and contract independently as their temperatures change and are not bound in any way by the brick setting. The steam is collected from a perforated steam pipe near the top of the steam space. The baffle shown in Fig. 230 prevents the steam which rises from the front header from carrying the water bodily into the steam space and makes the greater part of the water surface in the drum act as separating surface. The scale which accumulates inside of the tube is removed by tools inserted through the hand holes in the front headers opposite the ends of the tubes. One of these hand holes ^nd its cover are shown in section in Fig. 230. Soot and dust which accumulate on the outer surfaces of the tubes are blown off periodically by a steam jet, the necessary nozzle and hose being in- serted through the tall and narrow side cleaning doors shown in Fig. 227 opposite each " pass." A section of the Heine water-tube boiler is shown in Fig. 231. This boiler con- sists of a slightly inclined drum with dished heads, two sheet-steel headers and nu- merous tubes connecting these headers. The shape of the header is shown in Fig. 232, which indicates the positions occupied by the tubes and the way in which the header is joined to the drum. The products of combustion are generally made to pass along the tubes by the longitudinal baffles shown, instead of across the tubes as in the boiler last described. o o o o o o o o o o o 0°0°0 0°0°0°0°OoO°000 o o o o o o o o o o o o°o°o°o°o°o°o°o°o°o°o°o o o o oooooooo O o o° O o o c o «• O o O o 0»0«OoO»0 0/T)VS, . . . (102) in which H.P. = Rated boiler horse-power; A = Internal sectional area in feet of circular or square chimney; H — Height above grate in feet. (b) Mechanical Draft. Fans can be so used as to force air into the ash pit, that is, to raise the pressure on the STEAM BOILERS 397 entering side of the fire. In such cases the equipment is said to give forced draft. Or fans may be installed at the discharge end of the flues and may "draw " the gases through the boiler by lowering the pressure within to a value below that of the external atmosphere. Such an instal- lation is said to give induced draft. Forced draft suffers from the disadvantage that the pressure within the furnace is greater than atmospheric and hot gases may therefore be blown out when the fire door is opened. On the other hand, the fan handles only cool air instead of hot products of combustion as in the case of induced draft and its useful life is therefore much longer. Forced draft is much more common than induced draft. Several arrangements giving balanced draft have been developed. With such apparatus a pressure equal to atmos- pheric is maintained above the fuel bed and no hot gases are blown out through the firing door. PROBLEMS 1. The equivalent evaporation of a boiler during a certain test was 3450 lbs. per hour. What boiler horse-power was. de- veloped? 2. A water-tube boiler with 5000 sq.ft. of heating surface and rated in the ordinary way gave an equivalent evaporation of 25,875 lbs. per hour. At what per cent of rating was the boiler operating? 3. A certain boiler produced 3500 lbs. of dry steam in one hour from feed water at a temperature of 50° F. The steam pressure was 200 lbs. per square inch gauge. What was the equivalent evaporation? 4. A boiler receiving water at a temperature of 250° F. con- verts it into superheated steam at a pressure of 210 lbs. per square inch gauge and a temperature of 580° F. The boiler produces 26,000 lbs. of steam per hour. What is the equivalent evaporation if the boiler is given credit for all the heat given the material passing through it? What boiler horse-power is developed? 5. A boiler produces 7.5 lbs. of dry steam per pound of coal fired. The feed- water temperature is 80° F. and the steam pres- 398 STEAM POWER sure is 125 lbs. per square inch absolute. What is the equivalent evaporation per pound of coal? 6. A boiler is supplied with coal which has a calorific value of 13,520 B.t.u. per pound. It produces 8 lbs. of dry saturated steam at a pressure of 150 lbs. per square inch gauge per pound of coal. The feed-water temperature is 70° F. What is the efficiency of the outfit? CHAPTER XVIII RECOVERY OF WASTE HEAT 160. Waste Heat in Steam Plant. There are two great heat wastes in the steam plant — the waste in the hot gases going up the stack and the waste in exhaust steam. The magnitude of the stack loss can best be appreciated by determining an approximate value for assumed conditions. For this purpose assume the fuel to be pure carbon, the excess coefficient 1.5, average atmospheric temperature 60° F., average stack temperature 600° F., and no moisture in the air. The specific heat of the flue gases may be taken as constant and equal to 0.24. With an excess coefficient of 1.5, the total weight of flue gas per pound of carbon burned would be about 18.4 lbs. and the heat carried up the stack figured above room temperature would be Stack loss = 18.4X0.24 (600-60). = 2380 B.t.u. per pound of C burned (approx.) With a calorific value of 14,600 B.t.u. per pound of carbon this loss would be equivalent to a little over 16 per cent of the total heat in the fuel. It would be more correct to use the temperature of the steam in the boiler instead of room temperature, because the lowest temperature theoretically attainable by gases passing through a boiler would be equal to that of the steam and water on the other side of the heating surface. Under ordinary conditions of operation, this method of figuring would give a theoretically avoidable stack loss equal to about 50 per cent of the figure obtained above. 399 400 STEAM POWER The magnitude of the exhaust loss can be similarly approximated. Assume for this purpose an engine receiving dry saturated steam at 115 lbs. absolute per square inch and exhausting it with a quality of 90 per cent at a pressure of 15 lbs. absolute per square inch. The heat above 32° in the entering steam is 1188.8 B.t.u. per pound and the heat exhausted per pound is 1053.7. The heat in the exhaust represents therefore about 89 per cent of all the heat supplied when calculations are made above a temperature of 32° F. If a feed-water tempera- ture of 60° be assumed and heat quantities be figured above that datum the results are practically the same. There are always numerous pieces of auxiliary apparatus in steam plants such as boiler-feed pumps, circulating pumps, vacuum pumps, etc. These are often steam driven and are generally very uneconomical in the use of heat, so that they throw away in their exhaust steam large quantities of heat originally transferred from fuel to water and steam in the boiler. 161. Utilization of Exhaust for Heating Buildings. It often happens that steam-power plants are located within or in the neighborhood of buildings requiring artificial heat during part of the year. In such cases the exhaust steam from main and auxiliary engines can generally be advan- tageously used for this purpose. Under particularly favor- able circumstances, the weight of steam required by the plant may equal approximately that required for heating, and the greater part of the exhaust could then be turned directly into the heating system. The engines in plants of this character may be regarded as reducing valves for the heating system, receiving steam at high pressure and reducing the pressure to the value best adapted to the heating system installed. If the com- paratively small losses arising from radiation from the engine, from friction and from the presence of hot water in the exhaust be neglected, all heat received by the engine RECOVERY OF WASTE HEAT 401 and not turned into useful mechanical energy is made use of in the heating system. The engine may therefore be very uneconomical in the use of steam and still not cause a waste of fuel, provided always that the heating system can absorb all heat exhausted.- Since the demands of a heating system vary from day to day and since there is generally no demand for heat during several months of each year, it follows that a high degree of skill is necessary in choosing the character of the apparatus installed. A compromise is generally made between the cheap and uneconomical engine allowable during the coldest months and the more expensive and more efficient engine desirable when no heating is to be done. There are other cases of somewhat similar character. In many industries use can be made of exhaust steam for the heating of evaporating pans, dye vats, kilns and other apparatus. Steam plants of an uneconomical character may be very economical financially in connection with such industries if all or nearly all of the heat in the exhaust can be utilized industrially. 162. Feed- water Heating. An examination of the steam table will show that the total heat above 32° F. per pound of saturated steam varies between 1180 and 1200 B.t.u. for such pressures as are commonly used in boilers. The aver- age temperature of water as it occurs on the surface of the earth is probably somewhere in the neighborhood of 60°, so that the heat above 32° per pound would roughly average 27 B.t.u. A boiler receiving water at 60° and converting it into steam at any of the ordinary pressures must therefore supply over 1100 B.t.u. per pound of water. This immediately suggests a use for heat in exhaust steam. Steam exhausted into very low vacuums has a temperature only 10° to 30° higher than the assumed average natural feed temperature, but steam exhausted at atmos- pheric pressure has a temperature of 212° F. and could therefore impart large quantities of heat to water at 60° F. 402 STEAM POWER Since the boiler must supply over 1100 B.t.u. per pound of steam made, raising the feed temperature about 11° or 12° should effect a saving of about 1 per cent in fuel con- sumption. By raising the temperature from 60° to 212° there should therefore result* a saving of approximately 13 to 14 per cent. Other advantages which would accrue from this pre- liminary heating of the feed water would be (1) the deposit, outside of the boiler, of a large amount of the solid matter carried by the water, (2) the use of fewer or smaller boilers, and (3) the reduction of the strains which occur in the metal of some designs when very cold feed water is used. Exhaust steam feed-water heaters are divided into two types, open and closed heaters. In open heaters the steam and feed water are brought into intimate contact in the form of jets, sheets and sprays within a vessel of appropriate size and shape. They are often called contact heaters. When the exhaust steam comes from reciprocating engines it always carries in suspension some of the oil used for lubricating the engine cylinders. If allowed to enter the heater, this oil would mix with the feed water and eventually reach the boilers, where it might cause serious damage by depositing upon heating surfaces exposed to the fire or to very hot gases. Such heaters are therefore always fitted with oil or grease extractors when used with reciprocating units. When receiving the exhaust from turbines, oil ex- tractors are not necessary, as no lubricant is used within the steam spaces of such units. Closed heaters consist of tubes or coils enclosed within a metal vessel. One medium passes through the tubes and the other over their outer surfaces. Such heaters are therefore often called non-contact heaters. As oil is a poor conductor of heat, the exhaust steam from reciprocating units should be passed through an oil extractor before entering a closed heater in order that the heating surfaces may be used to the best advantage. RECOVERY OF WASTE HEAT 403 Exhaust steam feed-water heaters are often divided into primary and secondary heaters. This distinction has nothing to do with structure, being based entirely on position and temperature. Thus there may be available exhaust steam at a pressure below atmospheric, as from condensing main units, and exhaust steam at atmospheric pressure from non- condensing auxiliaries. The lower pressure steam could be used to heat the feed water in a primary heater and the higher pressure steam could then raise its temperature still further in a second or secondary heater. The other great waste, that in the stack gases, can also be partly eliminated by using some of it to heat the feed water. As the highest steam temperature ordinarily avail- able in the exhaust system is about 212° F., and as the products of combustion leaving the boilers generally have temperatures in the neighborhood of 600° to 700° F., it is evident that on a basis of temperature the hot gases have a decided advantage as a heating medium. On the other hand, the specific heat of the hot gases is low, while exhaust steam can give up all of its latent heat with no change in temperature, so that on a basis of heat avail- able for transmission to the water, the steam has the advantage. The waste heat in the flue gases is used for feed-water heating in devices known as economizers, preheaters, flue gas beaters, etc. These devices are now built in two radi- cally different forms. In one form the heater is mechani- cally separate from the boiler and is generally located in the flue beyond the boiler damper. In the other the heater is constructed as part of the boiler but is so arranged that water fed into it can be heated by escaping gases before mixing with the water in the main circulating system of the boiler. The separate form of heater is commonly known as an economizer while the one which is essentially part of the boiler is called a preheater, a preheater section, an integral eoonomizer, etc, 404 STEAM POWER For many years economizers were all of the separate type and were made entirely of cast iron: They consisted of cast-iron tubes fastened in groups into cast-iron headers and so arranged in the flues that the tubes stood vertical. The flue gases passed over the external surfaces of the tubes and the water on its way to the boiler flowed through tubes and headers. The apparatus was generally arranged so that water entered at the end at which the gases left and left at the end at which the gases entered. This maintains the greatest available temperature difference between gas and water throughout the entire economizer and is known as a counterflow arrangement. Cast iron was used both because it is a cheap material and because it is highly resistant to corrosion. Practically all real fuels contain sulphur and some of this sulphur appears in the flue gases as sulphur dioxide. This gas dissolved in water forms sulphurous acid and when further oxidized yields sulphuric acid. If under any conditions the flue gases are cooled to the dew point while passing through the economizer a certain amount of acid or acidu- lated water is deposited on the external surfaces of the tubes and headers and even with cast-iron corrosion is fairly rapid if such deposits occur in large quantities. On the other hand, boiler-feed water often contains gases and other impurities which corrode steel rapidly if the water is heated in contact with it. Cast-iron tubes have been found to resist such corrosion to a much greater extent than any form of steel. The cast-iron economizer is therefore much safer against both internal and external corrosion than is an economizer built of steel. Economizers are generally connected into the system between the boiler-feed pump and the boiler so that the water within them is at a pressure at least slightly greater than that within the boiler. When boiler pressures were low little thought was given to this fact but as boiler pres- RECOVERY OF WASTE HEAT 405 sures were increased many engineers questioned the prac- tice of using cast-iron economizers under full boiler pressure. Opinions as to the limiting permissible pressure within cast-iron economizers varies, being placed by different engineers at all values between 150 and 250 pounds. Boiler pressures have now reached 300 pounds, and higher in the larger and more modern plants. Three solutions have been developed. These are: 1. The use of cast-iron economizers at a pressure lower than that in the boiler, a pump drawing hot water from the economizer and pumping it into the boiler. This involves the use of two sets of pumps, one set to force the water into the economizer at a pressure sufficiently high to prevent vaporization when the temperature of the water is raised and the other to raise the pressure from that in the econom- izer to the pressure of the boiler. 2. The use of steel economizers at full pressure. The use n£ such equipment involves operation of such character as to guard against the existence of conditions leading to internal and external corrosion. 3. The use of cast-iron and steel economizers in series, the cast-iron economizer being operated under moderate pressure and at the lower temperatures at which the tendency to corrode is generally greatest. This arrangement entails the use of pumps between cast-iron and steel economizers, the latter being operated under full boiler pressure or higher. At the present time the tendency seems to be toward the use of cast-iron economizers under full boiler pressure in the plants using moderate steam pressures and in the smaller plants while more of the high pressure plants and the larger stations are adopting steel economizers operating under full pressure. The mixed systems and the double pump systems are generally regarded as undesirably costly and complicated. It is generally conceded that external corrosion of steel economizers can be prevented by never permitting the metal 406 STEAM POWER to have too low a temperature when in contact with flue gases. The limiting temperature varies with the quantity of sulphur in the fuel but exact values are not yet available. A metal temperature in excess of about 125° F. seems to be safe for bituminous fuel with a sulphur content not in excess of 2 per cent while a temperature of 150 to 160° F. seems to be necessary with other fuels having a sulphur content of the order of 5 per cent. It is also generally conceded that under most conditions internal corrosion can be prevented by satisfactory degasi- fication of the water before entrance to the economizer. All water dissolves air to certain definite quantities deter- mined by temperature and pressure relations if the oppor- tunity is offered and the dissolved oxygen seems to be an active corroding agent when aerated water is heated in contact with steel. The laws governing such solution of air in water are such that the solubility becomes zero when the water is at the point of vaporization. Thus if water under atmospheric pressure is heated to 212° F. under such conditions that the dissolved air can escape and pass off with the vapor generated, complete degasi- fication can be effected. Or if water under any lower pres- sure is brought to the temperature of vaporization corre- sponding to that pressure under similar conditions, com- plete degasification can be brought about. This phenomenon at atmospheric pressure can be observed by raising a vessel of water to the boiling point. The bubbles which form first are bubbles of dissolved gas and it will be noticed that many of them escape from solution before any appreciable quantity of visible vapor, i.e., steam, is formed. Those plants which are using steel economizers success- fully are making provision for adequate degassing and for preventing the degassed water from again dissolving gases before entry to the economizer. In designing and operating any plant which is to use economizers it is necessary to strike some sort of balance RECOVERY OF WASTE HEAT 407 between the amount of feed-water heating which is to be done by the exhaust steam and the amount which is to be done by gases. With the cast-iron economizers it has been quite common practice to heat the feed with exhaust steam only to that temperature required to guard against external corrosion with steel economizers both in- ternal and external corrosion must be taken into account. The temperatures which have beeen used successfully vary between about 120 and 210° F. depending upon the character of the fuel, the character of the water, the kind of economizer, the arrangement of the plant and other considerations. Economizers heat the feed water to temperatures vary- ing from about 200° to 350° F. depending upon initial temperature, the extent of the economizers surface, initial gas temperature, and other variables. The temperature of the flue gases drops as the temperature of the water rises, the numerical ratio between the two depending upon relative quantities and specific heats. The temperature of the gases leaving economizers is so low under good operating condi- tions that induced draft of some sort is generally necessary. For this reason induced draft fans are practically always used when economizers are installed. If the best results are to be obtained from the use of economizers the heating surfaces must be kept clean. This applies both to the gas side and the water side of those surfaces. In the case of cast-iron economizers it was cus- tomary to furnish mechanically operated scrapers for remov- ing the soot deposited on the outside surface of the tubes but many recent installations have omitted such scrapers and substituted provision for cleaning these surfaces with steam lances or soot blowers of some sort. Internal sur- faces are cleaned periodically in the same way as the sur- faces of boilers, the scale being removed by washing if soft, and by mechanical chipping devices if hard. The great amount of heat carried away by the flue gases when economizers are used would lead one to assume that 408 STEAM POWER economizers should always be installed. This is, however far from the truth. Economizers are costly and their instal- lation involves the provision of space, and supporting structure, additional flues, amplified draft apparatus, addi- tional piping, etc. All this represents increased investment, bringing certain definite capital charges. Moreover, the use of economizers brings in certain operating and main- tenance expenses not incurred if they are not installed. The advisability of using economizers can be determined only after a proper balance of increased charges against the money value of the fuel saving to be expected. In general, the higher the cost of the fuel and the more nearly the plant runs throughout the year at full capacity, the better the chances for a net saving by the installation of economizers, PROBLEMS 1. Determine the heat lost in the chimney gases per pound of coal in a plant operating under the following conditions, and express the loss as a percentage of the heat value of the coal. The coal has a calorific value of 14,000 B.t.u. per pound; the temperature of the gases leaving the boiler is 570° F.; 20 lbs. of gas result from each pound of coal burned; the mean value of the specific heat of the gases is 0.245; and the temperature of the air entering the furnace is 75° F. 2. Determine the quantity of heat which could be obtained from the gases of Prob. 1 by using an economizer to reduce their temperature to 250° F. What percentage of the heat value of a pound of coal does this saving represent? 3. The boilers of a certain plant produce 100,000 pounds of steam per hour when the plant is operating at full load. The steam-driven auxiliaries consume 10% of this steam. Steam is generated at a pressure of 175 lbs. per square inch gauge, and is superheated 150° F. The main units operate condensing and the condensate leaves the condensers at a temperature of 75° F. The auxiliaries operate non-condensing and exhaust their steam at atmospheric pressure and with a quality of 92%. The coal used has a calorific value of 13,850 B.t.u. The boiler efficiency /Heat given water and steam\ . \ Heat in fuel supplied / °' RECOVERY OF WASTE HEAT 409 (a) Determine the amount of coal which would have to be burned per hour if the steam exhausted from the auxiliaries were thrown away and make-up water at a temperature of 50° F. were used in its place. The condensate from the condensers of the main unit is assumed to be returned to the boiler after being mixed with the make-up water. (6) Determine the amount of coal which would have to be burned per hour if the auxiliary exhaust were used to heat the con- densate from the main units in an open heater and if the operation of the plant were so perfect that no make-up water had to be added. CHAPTER XIX BOILER-FEED PUMPS AND OTHER AUXILIARIES 163. Boiler-feed Pumps. The pumps used for forcing the feed water into boilers may be of reciprocating or centrifugal construction and may be driven by reciprocating steam cylinders, by small steam turbines or by electric motors. Steam-driven pumps are very wasteful, often using over 100 lbs. of steam per horse-power hour. It would therefore seem more economical to use motor-driven pumps in electric- power stations, as the large power units will generate electric power with a consumption of from 10 to 25 lbs. of steam per horse-power hour and the motor efficiency will generally be over 80 per cent. There is, however, another point which must be considered. The exhaust steam from small engines operating boiler-feed pumps can be used for heating the feed water as described in the last chapter, and thus the poor economy of these units is of little significance; practically all heat exhausted can be returned to the boiler in the boiler feed if desirable. As a result of this considera- tion, coupled with others of less importance, nearly all boiler-feed pumps and other similar auxiliaries are steam driven unless there are so many that there would be more exhaust steam than could be absorbed by the feed water. There is at present a marked tendency toward the use of turbine-driven, centrifugal pumps for boiler feeding, in place of those driven by reciprocating steam units. The turbine type has several advantages, the more important being : 410 BOILER-FEED PUMPS AND OTHER AUXILIARIES 411 (1) No oil in exhaust steam, so that latter is well adapted to use in all forms of feed- water heaters; (2) Higher speed because of continuous flow of water and continuous rotation of mechanical parts, thus making possible great decrease in size for a given amount of work, and (3) Better pump characteristics for this sort of work. The Duplex Steam Pump. The great majority of re- ciprocating steam pumps used for boiler-feed purposes are of the duplex pattern, one design of which is shown in Figs. Steam Valv.e Ciest Fig. 240. — Duplex Steam Pump. 240 and 241. Two steam cylinders are arranged side by side, their piston rods extending into similarly arranged water cylinders and Carrying water plungers or pistons as shown in Fig. 241. As there is no rotating shaft in a pump of this kind, the steam valves cannot be operated by eccen- trics as is common with steam engines. For the purpose of operating these valves, bell cranks, pivoted near the center of length of the pump, are provided. These are arranged so that the long arm of one bell crank engages a collar on the piston rod of one steam cylinder and the short arm operates the valve gear of the other steam cylinder. The motion of the valve of one cylinder is therefore derived 412 STEAM POWER from the piston motion of the other cylinder. The steam pistons are practically 180° out of phase, one moving out while the other moves in. Practically no expansion of the steam is obtained in the cylinders of pumps of this type. They operate on the rectangular cycle described in an earlier chapter and are correspondingly wasteful in their use of steam. Slide Valves Steam End Fig. 241. Water End -Duplex Steam Pump. A turbine-driven, centrifugal boiler-feed pump is shown in section in Fig. 242. The turbine is a multistage arrange- ment of the impulse type, having one Curtiss wheel at the high pressure end. The pump is a three-stage device, the first stage discharging to the suction of the second stage and the second stage discharging to the suction of the third stage. By multistaging in this way any desired boiler- feed pressure can be obtained with moderate rotative speed and diameter, [To face page J t 12. Fig. 242. — Steam Turbine-driven Centrifugal Boiler Feed Pump. [To face page 412.] BOILEE FEED PUMPS AND OTHER AUXILIARIES 413 164. The Steam Injector. On stc\m locomotives and in other portable steam plants, as well as in many small stationary plants, a device known as a steam injector is used, instead of a pump, for forcing feed water into the boiler. A simple form of steam injector is shown semi-diagram- matically in Fig. 243. 414 STEAM POWER Steam from the boiler flows through the steam nozzle and expands from boiler pressure to a very low pressure, thus acquiring a high velocity at the expense of the heat energy which it brings from the boiler. At the end of the nczzle it mixes with water and imparts to that water some of its kinetic energy, so that the mixture moves into the small end of the delivery tube with a high velocity. By the time it has reached that point, practically all the steam has been condensed, and, as the sectional area of the delivery tube increases, the velocity of the liquid decreases with a corresponding increase in pressure according to Bernoulli's theorem. In properly designed apparatus, the resultant pressure is great enough to force the mixture of water and condensed steam into the boiler against boiler pressure. The space at the end of the steam nozzle is maintained at a low temperature by the feed water flowing through it and the pressure of the steam is therefore very low at this point, being less than atmospheric in most cases. Atmos- pheric pressure is therefore able to force water up the suction pipe if the " lift " is not too great, and when once started such a device can therefore " raise " its own water as well as deliver it against pressure. It is interesting to note that the efficiency of this appa- ratus is almost 100 per cent on a heat basis. All heat not radiated from the apparatus is returned to the boiler in the mixture of condensed steam and feed water and, as the external surface is very small, very little heat is lost by radiation. 165. Separators. Two kinds of separators are used ki steam plants: (a) the oil separators already referred to for separating oil from exhaust steam, and (6) steam separators, which separate water from steam. As it is impossible entirely to prevent radiation from steam pipes, it follows that condensation will occur in any pipe line which carries saturated steam. Water is also formed in the cylinders of reciprocating engines not supplied BOILER-FEED PUMPS AND OTHER AUXILIARIES 415 SieyeS Jacket of Insulating Material to Decrease Radiation Loss. Water Storage / Chamber Fig. 244. — Steam Separator 416 STEAM POWEE with very highly superheated steam, and much of it is generally present in the exhaust of the high and intermediate cylinders of multiple-expansion engines. A small amount of water can be passed through the cylinder of a reciprocating engine without mechanical damage, but it probably causes a loss of heat by clinging to the walls and assisting in the heat interchanges which always occur. Large quantities of water are apt to cause mechanical damage, as water is inelastic, and if more of it is trapped in a cylinder end than can be con- tained in the clearance, something must give way when the piston reaches the end of its stroke. It is customary to separate as much as possible of the water of condensation before admitting steam to the cylinder. The separators used are built in many different shapes and types, but practically all depend upon two principles. These are: (1) Water is much more dense than steam, and if a stream of a mixture of water and steam be made to travel in a curve, the water will therefore collect at the outside of the curve, and (2) Water brought into violent contact with metallic surfaces " wets " them and has a tendency to adhere thereto. In steam separators the stream of mixture is therefore made to change its direction of flow suddenly and to impinge upon baffles in such a way that the greater part of the liquid is caught and drained off. One form of separator is shown in Fig. 244. The mixture impinges on seives in the first part of its passage through the separator, part of the water passing through the open- ings and draining to the reservoir at the bottom of the device. Ridges and troughs catch all water separated and guide it to drains leading to the reservoir so that no water which is once deposited is again picked up by steam. Another form of separator is illustrated in Fig. 245. The steam impinges upon the inverted V-shaped casting BOILER-FEED PUMPS AND OTHER AUXILIARIES 417 and water caught on the projecting ridges drains toward the sides and then downward into the receiver, while the steam passes on as shown. 166. Steam Traps. In the separators just described, there is a constant accumulation of water which must be drained off periodically if the entire device is not to fill up and become inoperative. Similarly there is a constant accumulation of liquid in steam jackets, in receivers of multi-expansion engines and in low points in steam lines. CO (*) Fig. 245. — A Steam Separator. 167. Steam Piping. There is a great deal of piping of various kinds in all steam plants and the financial success or failure of a plant often depends upon this apparently insignificant item. It is beyond the limits of a book of this scope to consider the many different forms of piping and the many different ways in which apparatus may be connected. This is a study in itself and one of great importance. It should be noted, however, that all of the following points must be kept in view when designing and installing piping and that that installation which most nearly meets all these requirements may be regarded as the best. (1) The various lines should conduct the materials flow- 418 STEAM POWER ing through them with the minimum loss of pressure and with the minimum loss (or gain) of heat. (2) The pipe lines should be so constructed as to mak'3 failure of a dangerous sort, from expansion and contraction, water hammer and such, most unlikely if not impossible. (3) All connections should be so made that the careless manipulation of valves cannot cause an accident. (4) The number of flange and screw connections and the number of valves and fittings should be reduced to the minimum, as they are often sources of weakness and are always costly. (5) The entire layout should be so arranged that inter- ruption of service because of pipe, or valve, failure is (as nearly as possible) impossible. (6) The cost of the system should be as small as it can be made, consistent with the other requirements. It is almost unnecessary to say that all of these desirable ends are never attained in any plant. A compromise must always be made in order to bring the cost within reasonable limits, but most of the recent installations show a tendency toward better design in this part of the plant and a con- sideration of reliability and safety far in excess of what was formerly customary. CHAPTER XX PERFORMANCE OF STEAM POWER EQUIPMENT 168. Meaning of Performance. The term performance is used in a general sense in engineering and refers to the extent to which, or the way in which, a material or piece of apparatus or a structure performs, that is, the extent to which it meets expectations or the way in which it compares with other similar things. The performance of steam power equipment of different sorts is measured in different ways, depending on the purpose in view. Thus the performance of a given engine with respect to a theoretical engine might be measured by comparing the quantity of steam used by the real engine when doing a certain amount of useful work, with the quantity of steam which would be required if the theoretical incomplete- expansion cycle could really be attained and utilized. Or in other cases the performance might merely be expressed in pounds of steam used per delivered or brake horse-power hour or per indicated horse-power for comparison with other engines of like or unlike characteristics. Or, the efficiency of a boiler might be determined by dividing the heat put into water and steam (i.e., useful out- put or result) by the heat in the fuel fired (i.e., input). The numerical value thus obtained would be a measure of the performance of the boiler with respect to the extent to which it utilized thermal energy. Or, the engineer might be interested in a complete plant generating electrical energy and he might desire to compare it with other plants doing the same thing and particularly to compare it with respect to its thermal efficiency. He might determine the overall efficiency by dividing the 419 420 STEAM POWER thermal equivalent of the electrical output by the heat value of the fuel supplied to produce that output, or he might determine the number of thermal units supplied in fuel for each kilowatt-hour produced. Either expression would serve as a measure of the performance of the plant in this respect. Or an engineer might want to compare the performance of one boiler room with another with respect to the total of all costs. For such purposes he might use the total cost (including fuel, water, supplies, labor and all main- tenance charges) per thousand pounds of steam produced. Such a figure would serve as a measure of performance for such a purpose. It is impossible to consider in a book of this kind all or even many of the different measures of performance used in connection with steam power equipment. Attention will be limited to some of the most common which all fall under the heading of efficiencies. Reduced to its simplest terms the function of a steam power plant consisting of steam boilers, prime movers and all associated apparatus, equipment and buildings is to produce a certain amount of useful mechanical energy as a result of expending a greater quantity of heat which is liberated by burning fuel. The extent to which it succeeds in a thermal sense is measured by the overall thermal effi- ciency of the plant, that is the quotient obtained by dividing the heat equivalent of the useful mechanical energy by the heat supplied in fuel. This thermal efficiency is not, by itself, a complete measure of the performance, since high efficiency does not necessarily mean lowest cost of power. This fact should never be lost sight of. The overall thermal efficiency of the plant is a com- posite made up of a number of other thermal efficiencies, principally those of the boilers and those of the prime movers. High boiler efficiency combined with high prime mover efficiencv does not necessarilv mean hi.2:h overall PERFORMANCE OF STEAM POWER EQUIPMENT 421 efficiency of the plant because the overall efficiency of the plant is also influenced by the ways in which auxiliary equipment is arranged and operated. However, with other things equal, or nearly so, the thermal efficiencies of boilers and prime movers will determine the over-all performance. These facts lead to frequent determinations of such thermal efficiencies for the purpose of guiding manufac- turers of such equipment, designers and builders of power plants and operators of power plants. It is the purpose of this chapter to consider briefly how such thermal effi- ciencies are determined. In order to be logical the boiler will be considered first and the prime mover second. 169. Determination of Boiler Performance. The thermal performance of the boiler is most commonly expressed as overall thermal efficiency. This means the ratio of heat put into water and steam in the boiler to the heat in the fuel supplied to the boiler furnace. Quite obviously some means must be available for determining not only the total quantity of fuel supplied in a given time but also the heat value, or calorific value, of that fuel. The quantity of fuel supplied is determined by weighing the fuel if it is solid and by means of appropriate meters if it be liquid or gas- eous. Weighing is also sometimes used with liquid fuels. The calorific value of the fuel is determined by testing carefully taken samples in fuel calorimeters which will be described in a later paragraph. With the heat supplied determined in this way it is still necessary to measure the useful output. This is deter- mined from (a) the quantity of steam leaving the boiler and (6) the heat added to it in the boiler. The quantity of steam produced is usually determined by measuring the water supplied the boiler, either by weighing or by means of carefully calibrated meters and then carefully guarding against loss through blowoff connections or other openings into the boiler. The temperature of the feed water is measured by means 422 STEAM POWER of an accurate thermometer as it enters the boiler to obtain a measure of the heat supplied in the feed water. The pressure and temperature or quality of the steam leaving are also determined. If the boiler is producing super- heated steam, pressure and temperature give all necessary data for determining the heat above 32° F. If the boiler is producing saturated steam, pressure and quality give the necessary information. Temperature of superheated steam is determined by means of a thermometer immersed in liquid contained in a well projecting into the interior of the steam pipe. Quality of saturated steam is determined by means of steam calorim- eters which are described in a later paragraph. 170. Fuel Calorimeters. The calorific value of fuels is determined by means of instruments known as fuel calorim- eters, which were briefly referred to in Chapter XVI. The fuel calorimeter is a device in which a known quantity of fuel can be completely burned under such conditions that all the heat liberated can be measured. The bomb calorimeter is the standard instrument for determining the calorific value of solid fuels such as coal and it is also sometimes used with liquid fuels. A vertical section through one make of bomb calorimeter is shown in Fig. 246. In this figure the letter a designates the bomb. This is a heavy walled vessel, generally made of steel and lined with a non-corrosive material such as nickel. The non- corrosive lining is used to prevent attack by acids formed during combustion of the fuel. Within the bomb is supported a small shallow vessel known as the crucible. This crucible receives the sample of fuel which is to be burned in the calorimeter. The sample is finely ground and sometimes dried before insertion in the crucible and an amount weighing about one gram is ordinarily used. When making a determination the sample is placed PERFORMANCE OF STEAM POWER EQUIPMENT 423 in the crucible, the crucible is placed on its support in the lower half of the bomb, the upper half of the bomb is put in place on the lower half and the two halves are locked together so as to form a gas-tight joint. The necessary oxygen for supporting combustion is insured by filling the bomb with oxygen under a pressure of several hundred pounds. Combustion is started elec- trically and it is then self con- tinuing until all of the com- bustible is consumed. Two different methods are used for initiating combustion. In one, a fine iron wire partly immersed in the coal in the crucible is heated by passing current through it until it ignites and burns and thus ignites the coal. In the other a platinum wire similarly located is heated in the same way. It ignites the coal but is not itself consumed. When the bomb has been Fig. 246.- prepared for a determination by the insertion of the sample and the necessary oxygen it is placed in the vessel b, figure 246, which is then filled with a known weight of water. A thermometer, designated by c, in the figure is provided for reading the temperature of the water and a motor-driven stirrer designated by d, is arranged to circulate the water and thus maintain it at a uniform temperature. The bomb, the water around it, the vessel b, the ther- mometer and the stirrer really form the calorimeter and could be used as such. However, it is customary to supply another vessel with double bottom and double walls to serve -A Bomb Calorimeter. 424 STEAM POWER as a thermal shield for the calorimeter proper. In the type shown in Fig. 246 this double- walled vessel is desig- nated by e and is filled with water. It is obvious that if a sample of fuel is burned in a device such as that shown, the heat which is liberated will heat the contents of the bomb, the bomb itself, the water in which it is immersed, the container or vessel b and parts of the thermometer and the stirrer. If no heat were used in any other way it would be possible to determine the total quan- tity liberated during combustion from the respective weights of the materials heated and the temperature rise as deter- mined by the thermometer c. In fact, however, the interchange of heat between the calorimeter and its surroundings and certain other phenom- ena make it necessary to add complications to what would otherwise be a simple procedure. The interchange of heat between the calorimeter and its surroundings is spoken of as " radiation " and a " radi- ation correction " is determined during each test. The determination is made by observing the rate of change of temperature of the water as indicated by the thermometer c after assembling the apparatus and while it is attempting to reach the same temperature as that of its surroundings. When the rate of change has been established by several readings at regular intervals the charge is ignited. Tem- perature readings are continued at certain definite intervals while the temperature of the water rises to a maximum due to the heat liberated within the bomb and then begins to drop again at a regular rate due to loss of heat to the sur- roundings. A radiation correction is calculated from the respective rates of change of the temperature before combustion and after the attainment of maximum temperature. By means of this correction, the temperature rise resulting from combustion (as read on the thermometer) is corrected to what it would have been had it been possible to thermally PERFOEMANCE OF STEAM POWER EQUIPMENT 425 isolate the calorimeter. With this corrected temperature rise it is then possible to calculate the total quantity of heat liberated within the bomb. Instead of using the respective weights of all of the different materials of the calorimeter in this calculation it is customary to determine, once for all, what is known as the " water equivalent " of the calorimeter. The water equivalent is that weight of water which would experience the same increase in temperature as does the real calorim- eter with any given supply of heat. This water equiv- alent in pounds, multiplied by the temperature rise in Fahrenheit degrees (after correction for radiation) would then give directly the British thermal units liberated within the bomb. There are several different methods of determining the water equivalent. That which is generally considered the most satisfactory is based upon the combustion within the bomb of a known weight of material of known calorific value. When this is done, and the necessary radiation correction is made, part of the known amount of heat liberated will be accounted for by the observed temperature rise of the known weight of water in the calorimeter. The remainder is accounted for by the same temperature rise of the " water equivalent " of the rest of the calorimeter. The water equivalent can therefore be calculated with the data available. In preceding paragraphs reference has been made to the " heat liberated within the bomb." This wording was used because the heat liberated within the bomb is not necessarily the same as the heat resulting from combustion of the fuel within the bomb. For example, if an iron wire is used for igniting the fuel and if this iron wire burns as it generally does, its combustion supplies heat in addition to that supplied by the fuel. In any case, such additional heat supplies must be evaluated and deducted from the total absorbed by the 426 STEAM POWER calorimeter in order to obtain the heat liberated by the combustion of the fuel. The bomb calorimeter, when properly constructed and used, is the most accurate instrument available for deter- mining the calorific value of solid fuels. It is, however, an instrument which requires very careful manipulation and it does not lend itself to rapid determinations. For such reasons many other types have been produced and some of them are capable of giving results which are sufficiently accurate for many commercial purposes. An entirely different type of calorimeter known as the Junker calorimeter is most commonly used for determin- ing the calorific values of gaseous fuels. It can also be used with many liquid fuels. Such a calorimeter arranged for use with gaseous fuel is shown in Fig. 247. The fuel gas enters the system through a very accurate meter indi- cated by a in the figure. From the meter it passes to the gasometer b which serves to maintain constant gas pressure. From the gasometer it passes through the tube c to a burner which projects into the center of the large cylinder d which is the calorimeter proper. The gas is burned within the cylinder d by means of air admitted through adjustable openings in the bottom of the cylinder. The products of combustion first pass upward to the top of the cylinder. They then flow down- ward in tubes of small diameter which pass through a space between the cylinder d and an inner cylindrical wall parallel to it. They then pass out of the instrument through the nozzle e which contains an adjustable damper for con- trolling the rate of flow. Water enters the instrument through the tube / being delivered to a small box or head tank g. An overflow pipe h serves to maintain a constant level in this head tank. From the head tank the water flows down the pipe i and into the calorimeter vessel through the regulating cock j. Within the calorimeter vessel it flows upward in the space PERFORMANCE OF STEAM POWER EQUIPMENT 427 between the two concentric cylinders previously mentioned, entirely surrounding the small diameter pipes which carry the products of combustion downward. The water finally leaves through the discharge tube k which spills it into a graduated measuring cylinder I when a run is being made. Fig. 247. — Junker Calorimeter. A thermometer m gives the temperature of the water entering the calorimeter and another one n gives the tem- perature of the water leaving. With the quantity of water and its entering and leaving temperatures determined, the quantity of heat liberated is calculated easily. The quantity of fuel used is measured by the meter a and it is therefore possible to calculate the heat value per unit of volume or weight. The instrument is commonly adjusted so that the prod- ucts of combustion leave the nozzle e at the same temper- 428 STEAM POWER ature as that with which the gas and air enter the instru- ment. Under such conditions the heat value determined is very nearly the higher heat value and it is commonly so regarded. 171. Steam Calorimeters. The devices known as steam calorimeters are used for determining the quality of satu- rated steam and, occasionally, the degrees superheat of superheated steam. They do not measure heat as do the fuel calorimeters. Since they are not heat measures the term calorimeter is really a misnomer. They are really quality meters. There are many different varieties of steam calorimeters, some very crude and some very exact instruments. Only the three most common types will be considered here. These are the Separating Calorimeter, the Throttling Calorim- eter and the Universal Calorimeter. The Separating Calorimeter is, as its name implies, a device in which quality is determined by separating out the water and determining its proportion of the total. A section through one form of separating calorimeter is shown in Fig. 248. Steam enters the device at the top passes through a small separating device indicated by a. The water which is carried in suspension separates out at this point and gravi- tates to the bottom of the inner vessel or cylinder indicated by b in the figure. The steam which has given up its sus- pended moisture passes through the small openings indi- cated, flows through the jacket space c and out at the bottom of the calorimeter through a tube d which leads it beneath the surface of water contained in the vessel e. The opening through which the steam leaves the jacket of the calorim- eter is made of small diameter to maintain the pressure in the jacket at practically the same as that within the calorimeter. The jacket temperature is thus maintained. A water glass or gauge glass, /, is connected into the water collecting space of the calorimeter and serves to indi- PEEFORMANCE OF STEAM POWEE EQUIPMENT 429 cate the height of the accumulating water. A slide, g, which can be moved up and down on this glass indicates the weight of water within the calorimeter by reference to a scale, h. The instrument is used by first allowing a sample of the steam under test to flow through it continuously and Fig. 248. — Separating Calorimeter. to run to waste until the entire instrument has been raised to working temperature. Water which collects in b during this period is drained down to a convenient point by bleeding through the cock i. The weight of water in the container e is determined by weighing or the height is noted by reference to the scale on the neck of the vessel. The height of water in the glass / is noted and then the dis- 430 STEAM POWER charge tube d is dropped as quickly as possible into the container e. The instrument then continues in operation until enough steam has passed through to give reasonable accuracy to the determination. The flow of steam through the instru- ment is then discontinued, the tube e is withdrawn from the container and the quantities of water separated in the calorimeter and of steam condensed in the container e are determined. If the effect of radiation and other possible sources of error is neglected the quality of the sample is then W * = T ^— , (103) W+w in which x = quality of sample expressed as a decimal fraction; W = dry steam leaving calorimeter and condensed in container, measured in pounds; w = water collected in calorimeter, measured in pounds. It is customary to minimize heat loss from the instru- ment and from the pipe bringing the sample to the calorim- eter by covering all parts with hair felt which is an excel- lent thermal insulator. For exact work it is, however, customary to determine the heat lost by " radiation " and to correct for this amount. This is conveniently done by expressing this heat loss in terms of the amount of water which will collect in the calorimeter in a given time and then using the formula for quality in the following form : x =w+^> (104) in which x, W and w have the same significance as before, and R = weight in pounds of that part of total liquid PERFORMANCE OF STEAM POWER EQUIPMENT 431 separated in calorimeter which must have condensed fco supply the heat lost from the instrument and connections. An improved form of separating calorimeter is shown in section in Fig. 249. The improvement consists in the use of a calibrated nozzle at the bottom of the jacket and a gauge graduated to read flow of material in a given time. When this in- strument is properly cali- brated and used it is capable of giving accurate results but the calibration should be checked at intervals by condensing the discharged steam. The separating calori- meter can be used satis- factorily to determine the moisture in steam over the entire commercial range of quality. However, with very small moisture con- tents it is necessary to work very carefully and to apply several corrections to ob- tain accurate results. For this reason many engineers prefer to use the throttling calorimeter under such con- ditions. A throttling calorimeter is shown in section in Fig. 250. Steam enters through a small nozzle a and leaves through a much larger opening b from which it is carried to any Fig. 249. — Separating Calorimeter. 432 STEAM POWER convenient point for disposal. The nozzle at a has so small an opening that steam cannot flow into the interior of the calorimeter fast enough to build up any great pres- sure. Under ordinary conditions of use the pressure within the calorimeter is equal to atmospheric or a fraction of an Fig. 250.— Throttling Calorimeter. inch of mercury above atmospheric. The steam flowing through the nozzle a therefore expands from the pressure in the pipe leading to the calorimeter to the much lower pressure within the calorimeter, a process analogous to that occurring in the nozzles of steam turbines as described in Chapter XIII. There is an important difference between the two cases which will appear in later paragraphs. If the calorimeter is assumed to have been in use for some time, so that all metal parts are heated up to working temperature and if loss of heat to surroundings be assumed to be zero, all of the energy in the steam entering through PERFORMANCE OF STEAM POWER EQUIPMENT 433 the nozzle must leave with the steam issuing from the large opening at the bottom of the calorimeter for the simple reason that there is no other way in which it can escape. This is the principle on which the calorimeter works and it can be put in the form of an equation thus, energy entering = energy leaving. The energy in the entering steam consists of two parts, that determined by its pres- sure and quality and determinable from the steam tables and that due to the velocity with which it flows in the small pipe leading to the calorimeter. This velocity is so small that it can be neglected without sensible error and it is always so neglected. The energy entering the calorimeter is therefore taken as that determined by the state of the steam and is equal to q-\-xr per pound of wet saturated steam as given in equation (15) page 34. The energy in the steam leaving the calorimeter is similarly that due to the state of the steam and that due to its velocity of efflux from the calorimeter. This velocity is always small under normal conditions of use so that it can be neglected. The energy determined by the state of the steam is therefore taken as that in the steam leaving and as that which must be equal to the heat above 32° F. in the entering steam. It follows from these facts that the fundamental equation of the throttling calorimeter be- comes Heat above 32° F. in entering steam = heat above 32° F. in leaving steam if heafc loss by " radiation " is neglected. If the steam tables are consulted it will be discovered that for saturated steam the total heat above 32° F. at a pressure near atmospheric is always lower than that for a higher pressure. In fact, if the higher pressure is 50 pounds or more as is common in power plant work the difference will be comparatively great. It is therefore true that if there is not too much water in the high-pressure steam flowing to the calorimeter the low-pressure steam leaving it will have to be superheated to contain the necessary 434 STEAM POWER heat above 32° F. This is the only condition under which 'this form of calorimeter can be used. Under such condi- tions, the fundamental equation of the calorimeter becomes [xr+ql^ly+Cpmils-Qh, (105) in which subscript 1 denotes values for entering steam, sub- script 2 denotes values for leaving steam, / s is temperature of the superheated steam in the calorimeter and t v is the temperature of vaporization corresponding to the pressure in the calorimeter. From this equation, the quality x musL be [k+C pm (t s —tv)]2 — Qi n (106) The calorimeter is used by allowing steam to blow through it and noting the pressure of the sample outside the calorimeter, the pressure inside the calorimeter, the barometric pressure and the temperature of the steam inside the calorimeter as indicated by the thermometer immersed in liquid in the thermometer well projecting into the interior of the calorimeter. The pressure within the calorimeter is expressed in pounds absolute and the cor- responding temperature of vaporization is determined from the steam tables. If this is less than the temperature indicated by the thermometer the steam in the calorimeter is superheated and the quality of the entering steam may be determined. It is only necessary to convert the pressure of entering steam into pounds absolute, take the necessary values from the steam tables and substitute in equation (106) above. If the steam entering the calorimeter is superheated instead of wet there will be no indication of such a fact until equation (106) is solved. The value obtained for x will then be greater than 1. When this happens it is pos- sible to determine the degree of superheat, if desired. For this purpose take from the steam table the total heat of PERFORMANCE OF STEAM POWER EQUIPMENT 435 dry saturated steam for the pressure at which the steam enters the condenser and subtract it from the total heat of the steam leaving the condenser, i.e., the right-hand member of eq. (105). The difference is the amount of heat contained in the entering steam by virtue of its super- heat and must be equal to [C vm (t s — t v )]\. The value of t v is given in the steam tables but C pm and t s are unknown. A solution must therefore be found by trial and error. That is, some value of / s is assumed, the corresponding value of C pm is obtained from Fig. 14 and both are substituted in the expression last given and the result compared with the amount of heat to be accounted for. If the difference is greater than permissible, another initial assumption is made, and this is continued until the desired degree of accuracy is obtained. All calculations outlined above for determining the quality or initial superheat by means of observations made with a throttling calorimeter can be eliminated by using the Mollier Chart, previously given as Fig. 154. It has been shown that values are determined merely by assuming the heat above 32° F. to be the same for entering and leaving steam and then solving for quality of entering steam after obtaining all other values by observation and from the steam tables. The Mollier Chart is so drawn that any change occurring with constant heat above 32° F. is represented by a vertical line. If the condition of the low pressure steam in the calorim- meter (and leaving the calorimeter) is spotted on this dia- gram it will be found to be somewhere near the middle of height (about 15 pounds) and to the left of the saturation curve. A line dropped vertically from that point to the curve for the initial steam pressure will show at its inter- section with that curve the initial condition of the steam. It is obvious that if the relative locations are such that this vertical line crosses the saturation curve the initial condition must have been that of wet steam, whereas if 436 STEAM POWER it does not cross, the initial condition must have been that of superheated steam. The difference in the behavior of steam flowing through a turbine nozzle and that of steam flowing through a throt- tling calorimeter nozzle can now be brought out. In the chapter on Steam Turbines it was shown that that which happens in a turbine nozzle is pictured by a horizontal line drawn from left to right on the Mollier Chart. Inspec- tion of the chart will show that this means a change which will reduce superheat or increase wetness. And yet in the case of a similar nozzle in a throttling calorimeter we seem to meet a different condition. That which occurs in the nozzle is substantially the same in both cases. It is in what happens after the steam leaves the nozzle that the difference occurs. In both cases expansion within the nozzle is essentially isentropic and represented by a horizontal line drawn from left to right on the diagram. The steam pressure drops, the heat above 32° F. decreases and the difference appears as kinetic energy of motion. In the case of a turbine nozzle the turbine wheel removes this kinetic or velocity energy from the steam and thus leaves it with the smaller heat above 32° F. with which it flowed out of the nozzle. In the case of the calorimeter the kinetic energy is con- verted back into heat as the jet of steam is brought to rest by bombarding its surroundings. The heat above 32° F. thus becomes the same at the end of the whole process as it was at the beginning merely because the kinetic energy developed in the nozzle is not removed as it is in the case of the turbine. It is obvious that a complete representation of what happens in the calorimeter is more complicated than the simple vertical line used above to obtain a solution. However, it is not necessary to study the intermediate steps of the process as initial and end conditions give all that is necessary for the determination of initial quality or superheat. PERFORMANCE OF STEAM POWER EQUIPMENT 437 The throttling calorimeter has certain definite limita- tions. It will work with any initial superheat but if the initial quality is too low the steam will not be superheated in the calorimeter and the initial quality cannot be deter- mined. This can be shown by means of the Mollier Chart. Assume, for instance that the intitial steam pressure is 50 lbs. absolute, that the initial quality is 95 per cent and that the pressure within the calorimeter is 15 lbs. abso- lute. Locate the intersection of the 50 lb. and 95 per cent line and run vertically upward on the chart to intersection with the 15 lb. line. The chart indicates that the steam will still contain moisture when expanded into the calorim- eter. In fact, the chart indicates that the initial quality will have to be about 97.5 per cent to just dry the steam in the calorimeter. It is not safe to use a throttling calorimeter with less than about 10° superheat in the calorimeter so that for 50 lbs. absolute steam pressure, the initial quality would have to be about 98 per cent or higher to make the throttling calorimeter usable. The percentage of water at which the throttling calorim- eter becomes inoperative increases with the initial pres- sure so that at 200 lbs. absolute a quality as low as 95 per cent can be determined with this instrument. The Universal Calorimeter is a combination of a throt- tling calorimeter and a separating calorimeter. It is so arranged that the steam passes first through the throttling calorimeter. If it is superheated in this part of the appa- ratus nothing further is required. If it is not superheated by the throttling calorimeter, the separating calorimeter removes the remaining water. The combination thus makes it possible to determine the quality of a sample with any initial quality. 172. Determination of Engine Performance. Engines are generally purchased on the basis of their thermal per- formance, expressed as pounds of steam of a given character 438 STEAM POWER which will be required per unit of output. The unit of output is sometimes taken as the indicated horse-power and sometimes as the delivered horse-power. The latter is the more logical since the delivered power is the useful output of the engine. The steam supplied can be determined easily when the engine exhausts into a surface condenser because under such circumstances all the steam exhausted is condensed to liquid water and can be weighed. When an engine is not fitted with a surface condenser the accurate determi- nation of the steam consumption is more difficult. The most common method is to isolate the necessary number of boilers and to supply all steam delivered by them to the engine under test. The feed water going to these toilers is then weighed, all losses through blowoff valves and other connections, safety valves, steam separators, etc., are deducted and the remainder is taken as the steam supplied the engine. Occasionally steam meters are used but such devices are not yet developed to the point where they can be regarded as at all comparable with tank and scales for accuracy. The character of steam delivered can be determined by pressure gage and thermometer, or steam calorimeter, at the engine. The indicated horse-power can be determined by means of the steam engine indicator already described in Chapter VIII. The determination of the delivered or brake horse- power is not always as simple a matter as might be desired. There are numerous ways available but many of them are cumbersome, or inaccurate or difficult of application, or have some other drawback. Some of the more common are considered in the following paragraphs. 173. Determination of Delivered Horse-power. The devices used to determine delivered horse-power are all classifiable as dynamometers. There are two distinctly PERFORMANCE OF STEAM POWER EQUIPMENT 439 different types known respectively as absorption dynamom- eters and transmission dynamometers. The absorption dynamometer absorbs the energy which it measures as is indicated by its name. The transmission dynamometer merely measures the energy as it passes or is transmitted without absorbing any of it, except a negligi- ble amount in its own friction. One of the most common forms of absorption dynamom- eter is a device known as the Prony brake. It is probable that the term brake horse-power comes from the use of such Fig. 251.— Prony Brake. a brake for determining the delivered horse-power of engines many years ago. One type of Prony brake as applied to a steam engine is shown semi-diagrammatically in Fig. 251. In this figure the wheel represents either the fly wheel of the engine, a belt wheel or the engine shaft or a special wheel fastened to the shaft for the purposes of the test. A strap a made of steel or leather or any other convenient material surrounds the wheel and is spaced away from its face at intervals by means of blocks of wood or other appro- priate material. A clamping device of some convenient form such as that shown at b is used to pull the blocks against the face of the wheel with any desired force. If the wheel is rotated and such a band without the long arm shown in the figure is clamped to it, the band will 440 STEAM POWER rotate with the wheel. If now the band is held stationary- while the wheel is rotated the friction of the blocks on the face of the wheel will offer some resistance to the rotation of the wheel. In fact, power will be expended in overcoming this friction and the energy " consumed " will appear as heat. By clamping more or less tightly more or less power can be absorbed in this way. This part of the Prony brake is therefore a power absorber. The frictional resistance to motion is equivalent to a tangential force applied at the rim of the wheel in such a direction as to resist its motion. If the wheel moves it must exert a tangential force equal and opposite to this resisting force. The energy supplied by the wheel (and absorbed by the brake) is equal to the product of this force by the distance through which it travels. That is Energy in foot pounds per minute = FX2tRu, . (107) in which F = tangential force in pounds; R = radius of wheel in feet: n = revolutions per minute. And the power supplied by the wheel would be . , Fx2irRn - v Power mb.p. = — . . . . (108) 1 33,000 It thus appears that if the tangential force could be measured the horse-power output could be determined. The tan- gential force is measured indirectly by measuring the reaction at the knife edge c at the end of the arm which holds the band or brake against rotation. It is obvious that the reaction at the knife edge which may be called F' must be as much smaller than the tangential force F required to hold the F f F ~~ R F'D- = FR PERFORMANCE OF STEAM POWER EQUIPMENT 441 brake stationary if applied at the surface of the wheel, as the distance D is greater than the radius R. That is and The product F'D may therefore be substituted in eq. (108) for FR giving i v. , i_ i i F'X2TrDn ,_ N Horse-power absorbed by brake = — , . (109) which is the equation for the Prony brake. The force or reaction F' is measured by means of scales as shown. It is the total reading of the scales when the wheel is revolving, less that part of the reading which is due to the weight of the brake arm and the weight of the pedestal on which the brake arm acts. The Prony brake cannot be used to absorb great quan- tities of energy. It can be designed to care for several hundred horse-power but becomes cumbersome when built to absorb more than about 100 horse-power. There are many other forms of brake and many of them can be used to absorb large quantities of energy. A brake known as the Alden brake is shown diagrammatically in Fig. 252. The disk a is keyed to the shaft and rotated by the engine under test. The spaces b are filled with water under any desired pressure and press flexible diaphragms c against the rotating disk. The friction between these diaphragms and the disk absorbs the power generated by the engine. The casing d to which the diaphragms are fastened is free to rotate and is prevented from doing so by weights suspended from an arm or by resting the end of the arm 442 STEAM POWER on scales as previously described. The formula is the same as for the Prony brake. A brake of this type can be made up with numerous Fig. 252.— Alden Brake. rotating disks and can thus be made to absorb large amounts of power. One of the most common forms of absorption dynamom- eters is a direct connected electric generator, that is, a gener- ator which is coupled to the shaft of the engine under test. If all losses in the generator are added to its electrical out- put the resultant sum must be the input to the generator and the output from the engine. The use of generators in this way makes possible the convenient testing of large engines and turbines which are used to drive direct con- nected generators. Transmission dynamometers merely pass the energy along and measure it while doing so instead of absorbing it as in the case of absorption dynamometers. Many transmission dynamometers have been developed but very few have been widely used for measuring the output of steam engines. One has, however, achieved some prominence in this use because it happens to offer about the only available means of determining the output of a marine engine or marine turbine in place in the ship and driving a propeller. This form is called a torsion dynamometer because its operation depends on the torsion or twist in a long shaft delivering power. Several forms are available but all PERFORMANCE OF STEAM POWER EQUIPMENT 443 merely indicate or record the relative twist between points on the shaft surface at a known distance apart along the length of the shaft. The horse-power transmitted is then calculated by formulas involving the shaft diameter, physical constants of the shaft material, angle of twist, revolutions and length of shaft under test. TABLES 446 STEAM POWER rO d en d 0) o ft s I 1 o Q p O Pn o Eh o CO w 1—1 H P^ W Ph o Pm H W H O > 3 PQ si « -p > P-3 MM >>.pW a; ^h-S d W w 03 „ & os >£ ■< o,^ ^• B TJ ^coco^ g O TO 03 HH fc ^Id^I o3 " p aj a> o ~ -S a;- 73 ° OS Pres. Abs. Lbs. Sq.ip a Pm S^ «£8 CO I?" O (MOCi CO l> (N Q5 TtUO^^t^ C005t^iOrt< CO i-i o M & o a 03 CM 001OCOC0"* COTHCOOilO CO OrhiiOCOlM NCOMNN CO r^ r^ CO ^ rH OOCO-^NH rH OOIOOOOCO t^l>l>l>l>» d NHH 1-5 t-I l-H i-i rH r-i rH 'B < 0.0000 0.0536 0.0913 0.1151 0.1313 0.1454 0.1559 0.1662 0.1747 0.1816 0.1883 13 O H g < 2.1832 2.0944 2.0358 2.0007 1.9776 1.9678 1.9435 1.9296 1.9183 1.9095 1.9008 to 3 P < § P3 H W H w 05 2 I < w -(J d e8 ►H* f-1 . K.-H += 03 * d H * rH OCCO CO CO »0 U0 to O OOOiOiO OS OS OS C5 C5 O H 1073.4 1058.3 1047.3 1040.0 1035 . 1 1030.6 1027.2 1023.9 1021.1 1018.8 1016.5 O) 03 OO^H !>Ttl CM rH O OS 00 O OOOOSOS OSOSOS0000 d t^i^dood cQo6coi>rH CNjTticOcOt> OOOOOSOSO "a o 1073.4 1085.4 1094.3 1100.1 1104.0 1107.5 1110.2 1112.8 1115.0 1116.7 1118.4 Wo H « Oil CiOiC*i-iG> iOHOOt)! CO IQt-CSO© HtNNWW r-i iH iH iH tH rH rH w Ph P CO to W « P4 CO CO d a 0.0886 0.2472 0.4893 0.741 0.975 1.235 1.467 1.736 1.992 2.219 2.467 CO v L - -§o|t 1 0.1804 0.50+ 1.00- 1.50+ 2.00- 2.50 + 3.00- 3.50 + 4.00 + 4.50 + 5.00 + SATURATED STEAM TABLE 447 t-ie© i-ie*eort. t-I iH tH tH 1 *" ■ iq ONOOOJO 1-1 t-i iH t-I N iHNCOmiO CO OscOt- CO 00 00 "OlM COCO O COCO (N i-Hcooo OS t^ 1>CM os go co r^ oo NNHCO CO t^ CM CO O t-h CO CO OS CO CO CO 00O CO CO!> 1-1 OSl> COHrH HCONWO0 CO LO"* rH CO lONOOO CO CO CO coco NhOOtJ( CO CO CO i—i 00 00 O -* go loco cm o LO lo »o »o lo >oooco OS co co cm CO 1> COLO ^1 T^H Tt 1 Tfl i> CO rH t-H "*rH rHrH CO CM rH O OS -* rH "* rH CO OOrHCONO oo go r^ coco COCO CO CO CO b- OS GO 00 GO (M -* O OS -* CONOrHOO HHNcqiM t-h OscocOCM J>1>- t^ lOCO rJH to cot- GO CM- i-H LO i-l 0 OS CM OS CO CO Os CO co-* -* rH lo CO CO CO CO CO OOOOO OOOOO OOOO oo OOOOO OOOOO -* O GO-* CM LO 00 -* i— l CO Jh- i-h 00 CO tH OS OS CO CO 00 L.8285 L.8161 L.8053 L.7S68 L.7874 l^l> rH rH OS (N COO r^r^coco iCOS CO-* to to rH -* o oo OS-* O cO(M rH rH rH COCO Ot-itH OS CO 00 tH o COCO CM CM CM t-h t-i N03 01H05 CC*OCOH -* CO CM CO COOS CM LO GO O CM "* t^OSOCM i-i «* tO £~ l> CO CO CO CO CO 00 O O O i-( CO c©l> J^J> iHHMN CM CM CO CO COrH -* r^r^l>L^J> tH tH -* LO LO OS 1>- -* CO Tf OrH CM-* Os GO oooco coco -* t-h os 00 00 OSO CM LOGO CM COCO CO LO "* CO CO OS Os OS OS Oi i> CM CO-* O WINHHH OS OS OS OS OS t^rH CM OS OOO OS OS OS OS GO b- CO OS OS GO GO rH CM Osi>- lO OS OS CO GO GO oooooo oooo CO CO O COCO GO OOGOt^L^- GO 00 00 00 00 CO O CO t^ CO 00 00 CM O O CM CO CM OS CO CD1> 00 O MNHiOO 1034 1021 1012 1005 1000 OS OS GO 00 00 OS OS OS OS Os OSCO-*i-H OS Os OS Os Os os NiOMHO CO CO cO cO CO OS OS OS OS OS co colo co oq LO LO LO LO LO OS OS OS Os Os 69.8 94.0 109.4 120.9 130.1 OSt^OOCMrH i> -*' O CO i i-I CO -* LOLOCO T— 1 T— 1 T-H 1— 1 1— 1 i>OS COLO LOOSCO b- cOcOt^t^ oo o'.t-I 00 00 TH lO lO Tfl l— * •*r^ oco cd CO GO OS OS OS GOCOOOi-H-* 00 i-H CO CO 00 osoooo rH CM CM CM CM -* o o»o »o N>OOHH OSLOOrH -*1> O t-i CM 00 00 OS o LO LO LO LO CO CQ1OMH00 GOrH lOO CM CO iCCON lo l>OS00»O OO •CO CM i-< CO-* rJH o CO CO CO CO-* CM CM CM •!>. GO GO OS 1> t-H LO OS OsOOO i-H CM 00 C7SO CJCMCMCqCM CO oo * CO CO CO CO CO CO CO CO CO CO tHCMCO"* no CO l> CO oso 448 STEAM POWER ! rH^£ ^ CO t- COOS O (NCNCNCNCO «th«oooo CO CO CO CO ^* WtH»00O t* rH tH •># IO m >6 CO co oor-c75^ NhcOhN CO CO 00 rH OS (M Os O rr rH O tOCO 00 »0 tC IO T* -o CO 00 CO 1> CO t^t^ 00 O0 OS CO CO CO CO CO CO CO Ol>C0 (OOi*NH OSOOOrH CO Tfl TtH tH tH 1 ooooo ooooo OOOOO t w. P 13 1 g < CO b- 00 Cii-H HOOOCi I>1> t^t- CO GO rH COCO rH CO OSTtH OCO OS 00 00 00 I>- co CO CO CO cO rH CO t^ CO rH (NOO^rMGO l> CO CO CO to CO co CD CO CO 00 s p l> oo oo oo s X rH tOO tOO t^ t^t^l> CO 00 00 00 00 00 CO t* oo CO OS coco ooo to CO CD CO to to oo oo oo oo oo CD CO CQ rH O CO rH OS l> tO tO tO Th TfH TH 00 00 00 00 00 p o H (ONOO^H tO rH 1> IO CO CO CO CO CO to H CQ O OSb- CO to tO tJ< tH rH H/i OS OS OS OS OS NON>0« tH rJH CO CO CO OS OS OS OS OS rH OS t^ tO CO COCO CO CO CO OS OS OS OS OS IS § ^ CO t^ 00 00 00 COCO COOS rH rH O 00 tO rH OM-^ CO 00 CO CO OS CO CO CO CO CO CO CO CO COCO CO CO OS CO HHt^O CO TjH TH -HH IO 00OS co CD cO cO cO OrHCOCOCO - 242.2 244.4 246.4 248.4 250.3 HOOIMCO TjH 1> rH "$ 1> to to CO CO CO CO CO CO CQ CO CO rH OOtOO OCO lOOOrH l>t^t^t^00 COCO CO CO CO e 00 02 B « * oa Q B. tst-oooo OS rH CO tO rHrHCOCOCO l> OS rH CO tO CO CO CO CO CO SATURATED STEAM TABLE 449 (N^WOOO to to 10 uaco oo t> t> t^ i> cotoocdcd cd »o 10 io »o 10 10 10 10 •*" rj< rdi rfl ^ tjh o CNOOOOOO "*Ol>COcO 00O»OO*O CCHOOH C0iO00 OC01>i-iCO 0>00^fl5 tJHhiOh® -t^cO CO O »0 iO "* ■* CO CO CO 0 t^OrH CO"* WNON "* •O to CO CO CO T^ ^H T^l ^^ CO CDGOON^t* CO CO t^ t^ t^- ooooo OOOOO ooooo OOOOO OOOOO 6549 6519 6490 6460 6432 fflO»OHN O 00 lOCO o "* CO COCO CO cO CO cO CO CO *OCOCO 00 "* OOCOO I> "* rH 05 CO co i-i 05t>- io "*"*COCOCO 00 00 00 00 00 CONOOCN CO CO CO *0"* CO oo oooooo 00 HOGNtO HH OOO oooooooooo l>OSC0 ON-*NO) r^iococ^o rH OOOCO"* CO 00 CO OcM lO iO *C iO CO CM I> t^t>.oooo Cq 00 o oooo oo oo o ■ 00 00O O OOHHN oo oo oooooo (NCOCO"*"* OOOO 00 00 00 rH iO iO CO CO oooooooooo 05INION OOOOO 00I>»OCO© 00 "*rHl> CO 00*OtH00 COiOGOOiM oo oo ooo o . ooo rH rH rHrH (N CO COCO COCO rH CO"* CO!> o o o o CO CO CO CO CO CO CO CO CO CO CO CO CO CO CO CO CO CO CO lO CO CO CO CO CO CO CO ^ ^ ^ lOiOOOcO «OcOI>l>l> 450 STEAM POWER Pres. Abs. Lbs. Sq.in. a IOOI0010 OHHNN © IO© lOO CO CO ^ ^ to >o o to O IO tot© CO C- t» rail ONOCDM CO rfH 00 CO 00 NO00N1O (M h O M (M IOC0HHH r« CO CO i-h © ©^COtOCO NCOtONO 05 00NOCO xtn tHH CO CO CO CO COCO CO CO CO CO CO CO CO o « & H O a < hoOOtHO OS O CO to 00 HHOCD00 l-H t-HtH ©© ©co toco © O^NHiO oon ©©to ©©©©© ©T-H ©T-H 00 G0CONC0 © "-tf rH CO CO CO ©©©©© "2 13 "2 s CO OS OS OS 00 00 to to to to to nnn © co ON TjH H © OONNNO iO to to to to ^ © tO©N © co t-h as © © © © iO iO IO to to to to g s p .< « H W H w CO B M ffl I H W ►H* * CO 00 OS © t-H co tf ©r>oo © ©CO CO CO HHH(N(M 00 00 00 00 00 CO CO CO CO CO 00 00 00 00 '00 CO coco coco 00 00 00 00 00 ©N ©CO © OiOONi* rH 00 © tO ^ COON»ON O © OS OS OS OOQONNN ©N lOCO © © oo oo oo oo GO tOCO rH © NNNN © Is O H 3 NiOOONN co ©© "*co O00 0CNN oo qonn n oo oo oo oo oo CO ©N lOCO t^ co co © © 00 00 00 00 00 rH 00©TtHC0 CO to »o to to 00 00 oooooo a; O iO O CO iO ©n ©^co © ©C0NC0 0©©i-ht-h CO COCO CO CO 00 tH t-jh> © h co co coco CO CO CO CO CO CO to 00 © CO CO CO CO TjH rtH CO CO CO CO CO Is O H s (MOOOCOfO © ©CO 00 ^ ©lO©TjH OS 1> 00 GO OSO oo oo oo oo os tH rH CO CO CO © © © © © tH •* to to to © © © © © ft . - TJH 00 t-H CO'* xH CO tH 00 to © © © tooo i-H tJH 00 tH tHH CO CO CO ^ ^ CO CO CO CO CO n © co to oo xHH IO IO tO iO CO CO CO CO CO i-H CO © 00 © co © © © t> CO CO CO CO CO H K tn « cd 5S 3 a IOOIOOIO OHHNN ©tooio© CO CO tJ< ^) to >oo too to to co co I> t- Pres- sure. Lbs. a & a CO CO CO CO CO © lodod OS OS © O t-h CO CO CO CO CO IO © iO © to t-h CO CO COCO . CO CO CO coco © tO©tO© «HH TjH LO «0 © SATUBATED STEAM TABLE 451 OlOOiOO COOOOOIO NWtHWOON COOO CO CD O O i-< COCDOrtiOS lOtONCO IC *0 rti rti CO O »OrH^ CO OS HCOHCOH NhhOO OOOOO OiOGOCM CO CM CD 1> T-H 00 CO 00 OOOOO OOOOO CO O 00 CO CO ^N05N>0 IO »0 Th tJH "* LO »0 iO iO »0 COOS J> CM OS O OS . - io »o -* Ttn co *~ ^ ."^^ iocoi> co io TJH -*«# -* »0 O 00 i> loeoH os CO CO CD CO *0 NOHQCNOO ^ CM ON 00 CO t> t^ t> CO IO Tt< GO 00 OS O CM CO CO O O OOcOiO CO IO Tt< TJH ^ Tt< GO GO GO GO GO COHCMWOO CM i-h GO CD CO OS 00 00NNCO-* COO'tf NOS iO GO O CM ■*' tF -^ iO »0 to CO CO CO CO CO »OcM CM G0»O, COCO'* rt^ iO Tt*O0MNH »OH COCONNOO OS OS Os Os Os i-i -<* CO 00 OS T-iiOGOCO MIONC5H NNNNGO CO CO CO CO CO HN^MOOCD •<* ■"*'<*■<* iO CD N oiaoioo oo oo o cs o ooooooo tOOOOOOO NMtJUOOON CO CO CO CO CO CO CO CO CO CO • • • • • IO IO lOtOlO 00 CO COOO oo os o t-1 t-1 t-1 t-1 i-i CMCMCO*** * * 452 STEAM POWER PROPERTIES OF ONE POUND OF SUPERHEATED STEAM [Condensed from Marks and Davis's Steam Tables and Diageams, 1909, by permission of the publishers, Longmans, Green & Co.] Sp.V. = specific volume in cu.ft.; AQ = B.t.u. total heat above 32° F.; A<£ = total entropy above 32° F. Absolute Pressure. Lbs. Sq.in. Sat. Temp. °F. Degrees of Superheat. 50 100 150 200 250 300 15 (213) 50 (281) 100 (327.8) 110 (334.8) 120 (341.3) 130 (347.4) 140 (353.1) 150 (358.5) Sp.V. AQ A(f> Sp.V. AQ A Sp.V. AQ Acj> Sp.V. AQ A<}> Sp.V. AQ A0 Sp. V. AQ Acj> Sp.V. AQ A0 Sp.V. AQ A 26.27 1150.7 1.7549 8.51 1173.6 1 . 6581 4.43 1186.3 1.6020 4.05 1188.0 1 . 5942 3.73 1189.6 1.5873 3.45 1191.0 1.5807 3.22 1192.2 1.5747 3.01 1193.4 1.5692 28.40 1174.2 1.7886 9.19 1198.8 1.6909 4.79 1213 1.6358 4.38 1215.9 1 . 6282 4.04 1217.9 1.6216 3.74 1219.7 1.6153 3.49 1221.4 1.6096 3.27 1223.0 1 . 6043 30.46 1197.6 1.8199 9.84 1223.4 1.7211 5.14 1239.7 1.6658 4.70 1242.0 1 . 6583 4.33 1244.1 1.6517 4.02 1246.1 1.6453 3.75 1248.0 1.6395 3.51 1249.6 1.6343 32.50 1221.0 1.8492 10.48 1247.7 1.7491 5.47 1264.7 1 . 6933 5.01 1267.1 1.6857 4.62 1269.3 1.6789 4.28 1271.4 1.6724 4.00 1273.3 1 . 6666 3.75 1275.1 1.6612 34.53 1244.4 1.8768 11.11 1271.8 1.7755 5.80 1289.4 1.7188 5.31 1291.9 1.7110 4.89 1294.1 1 . 7041 4.54 1296.2 1.6976 4.24 1298.2 1.6916 3.97 1300.0 1 . 6862 36.56 1267.7 1.9029 11.74 1295.8 1.8002 6.12 1313.6 1.7428 5.61 1316.2 1.7350 5.17 1318.4 1 . 7280 4. 80 1320.6 1.7213 4.48 1322.6 1.7152 4.19 1324.5 1 . 7097 38.58 1291.1 1.9276 12.36 1319.7 1.8237 6.44 1337.8 1.7656 5.90 1340.4 1.7576 5.44 1342.7 1.7505 5.05 1344.9 1.7437 4.71 1346.9 1.7376 4.41 1348.8 1 . 7320 SUPERHEATED STEAM TABLE 453 PROPERTIES OF ONE POUND OF SUPERHEATED STEAM {Continued) Absolute Pressure. Lbs. Sq.in. Sat. Temp. o R Degrees of Superheat. 50 100 150 200 250 300 160 (363.6) 170 (368 . 5) 180 (373.1) 190 (377.6) 200 (381.9) 300 (417.5) 500 (467.3) Sp. V. AQ A(f> Sp. V. AQ A Sp. V. AQ A0 Sp. V. AQ A0 Sp. V. AQ A0 Sp. V. AQ A Sp. V. AQ A(f> 2.83 1194.5 1 . 5693 2. 1195.4 1 . 5590 2.53 1196.4 1.5543 2.41 1197.3 1.5498 2.29 1198.1 1.5456 1.55 1204.1 1.5129 0.93 1210.0 1.470 3.07 1224.5 1 . 5993 2.91 1225.9 1 . 5947 2.75 1227.2 1.5904 2.62 1228.6 1.5862 2.49 1229.8 1.5823 1.69 1240.3 1 . 5530 1.03 1256 1.519 3.30 1251.3 1 . 6292 3.12 1252 1 . 6246 2.96 1254.3 1.6201 2.81 1255.7 1.6159 2.68 1257.1 1.6120 1.83 1268.2 1.5824 1.11 1285 1.548 3.53 1276 1.6561 3.34 1278 . 4 1.6513 3.16 1279.9 1.6468 3.00 1281.3 1.6425 2.86 1282.6 1.6385 1.96 1294.0 1.6082 1.22 1311 1.573 3.74 1301.7 1.6810 3.54 1303.3 1 . 6762 3.35 1304 1.6716 3.19 1306.3 1.6627 3.04 1307.7 1 . 6632 2.09 1319.3 1 . 6323 1.31 1337 1.597 3.95 1326.2 1.7043 3.73 1327.9 1.6994 3.54 1329.5 1.6948 3.37 1330.9 1 . 6904 3.21 1332.4 1 . 6862 2.21 1344.3 1.6550 1.39 1362 1.619 4.15 1350.6 1 . 7266 3.92 1352.3 1.7217 3.72 1353.9 1.7169 3.55 1355.5 1.7124 3.38 1357.0 1.7082 2.33 1369.2 1.6765 1.47 1388 1.640 INDEX PAGE Absolute pressures 41 Absolute temperature scale 12 Action of steam, in cylinder 24 on impulse blades of steam turbine 237-239 Adiabatic expansion 58 Advance angle 167 Advantages of condensing 261, 262 Advantages, relative, of contact and non-contact condensers . . . 293 ,294 Air, excess, combustion 305 Advantages and disadvantages of 332 Analogy, hydraulic 26 Analyses of coal (see fuels) 320-322 Purchase of coal on analysis 323 Angle of advance 167 Ash in coal 321 Atmospheric line on indicator diagram 119 Atoms 296 Avogadro's Law 300 Babcock & Wilcox superheater 391 " water-tube boiler 375-377 Balanced slide valves 184 Barometer, conversion of readings from inches mercury to pounds per square inch 265, 266 Barometric Condenser 271-278 Baume scale to express gravity 303 Bearings Ill, 112 Bilgram diagram 168-182 Angularity of connecting rod 179 Diagram for both cylinder ends 177 Exhaust and compression 175-177 Indicator diagram from 180-183 Piston positions ■. .'. 177-182 Blades, impulse, action of steam on, in impulse turbine 237-239 455 456 INDEX PAGE Boiler-feed pumps and other auxiliaries 410-418 Boiler, generation of steam in 38, 39 Boilers, steam 326-398 Circulation in 365, 366 Classification according to — (1) form; (2) location of furnace; (3) use; (4) direc- tion of principal axis; (5) relative positions of water and hot gases . 326, 327 Draft apparatus 394-397 Chimneys or stacks 394-396 Mechanical draft 396, 397 Effects of soot and scale 388, 389 Efficiencies 387, 388 Functions of parts 327-329 Furnaces and combustion 329-332 Hand firing 332-336 Mechanical grates . . 336, 337 Mechanical stokers 338-359 Rate of combustion 359-381 Rating 382-386 Boiler horse-power 384 Equivalent evaporation 385 Scale 389, 390 Prevention of 390 Smoke and its prevention 337, 338 Strength and safety 361-365 Superheaters — Built in 390, 391 Separately fired 390, 391 Babcock & Wilcox 391 Foster 393, 394 Heine 391, 392 Types of boilers 366-382 Babcock & ,Wilcox, water-tube 375-377 Continental :-. 370-372 Externally fired, return tubular 373-382 Heine water-tube 378-380 Internally fired, tubular 367-373 Locomotive 370 Scotch marine 372, 373 Sterling water-tube 380-382 Vertical fire-tube 367-370 Wicks vertical water-tube. 382, 383' INDEX 457 PAGE British Thermal unit s, 13 Buildings, heating of, by exhaust steam 400, 401 Built-in superheaters 390, 391 Calorific value of coals — Dulong's formula 322, 323 Fuel Calorimeter 323 Calorific value of petroleum oils 324, 325 Calorimeter, fuel 323, 422-428 Calorimeter, steam 428-438 Carbon, combustion of 298 CO, combustion to 298-301 C0 2 , combustion to 301, 302 CO and C0 2 , conditions determining formation of . . . 303-305 CO to C0 2 , combustion of 302, 303 complete combustion of 298 flue gases from combustion of 305, 306 Card factors and conventional diagram 125-128 Centigrade scale 10, 11, 12 Chain grate stokers 339-344 Chart— Mollier, for steam 230 temperature-entropy, for steam 62-65 Chimneys or stacks 394-396 Circulation in boilers 365, 366 Classification of boilers, according to — (1) form; (2) location of furnace; (3) use; (4) direction of principal axis; (5) relative position of water and hot gases . 326, 327 Classification of steam engines 92, 93 Clearance — steam engine — mechanical and volumetric 84, 85 Clearance volume determined from diagram 131, 132 Closed and open feed-water heaters 402-408 Coal-fuels 318-320 Analyses of — proximate and ultimate 320-322 Purchase of, on analysis 323 Coefficient, excess in combustion 305 Combined indicator diagrams 155-158 Combined type turbine 251 Combustion and furnaces; steam boilers 329-332 Combustion 296-316 Definitions — Compounds, elements, heat or calorific value, atoms, molecules, etc 296-298 458 INDEX PAGE Combustion — Combustion of — Carbon 298 Hydrocarbons 308, 309 Hydrogen 306-308 Calorific value of 309 Mixtures 309, 310 Sulphur 309 Combustion to — CO 298-301 C0 2 301, 302 CO to C0 2 302, 303 Conditions determining formation of CO and C0 2 303-305 Excess Air and excess coefficient 305 Flue gases from combustion of carbon 305, 306 Rate of, in boiler furnaces 359-361 Temperature of combustion 312-314 Theoretical temperature 313 Commercial fuels — solid, liquid and gaseous 317, 318 Complete expansion cycle 55-58, 72 Complete T^-chart for steam 68-70 Compound engine 149-151 Compounding 141-158 Combined indicator diagram 155-158 Compounding 144-149 Cylinder ratios 151-153 Gain by expansion 141-144 Indicator diagrams and mean pressures 153-155 The compound engine 149-151 Compounds — combustion 296-298 Compression and exhaust — Bilgram diagram 175-177 Condensation, cylinder, methods of decreasing 89-92 Condensation, initial 81, 82 determination of 86-89 Condensers and related apparatus 261-295 Advantages of condensing 261-262 Conversion of readings from inches of mercury to lbs. per square inch 265, 266 Cooling towers 294, 295 Measurement of vacuum 262, 265 Principle of 266-268 Types of — Contact 268-278 INDEX 459 PAGE Condensers (continued) — Contact — Barometric 271-278 Jet, Parallel flow 269-271 Siphon 276 Westinghouse — Leblanc 276-278 Non-contact 278-282 Surface 278-281 Two-pass or double flow 280-281 Relative advantages 293, 294 Water required by contact condensers 291, 292 Water required by non-contact condensers 292, 293 Condensing, advantages of 261, 262 Condensing plants 23 Conditions determining formation of CO and C0 2 303-305 Connecting rod 109, 110 Angularity of 179 Conservation of Energy, law of 2 Conservation of Matter, law of 1 Constant-quality lines on T^-chart 66, 67 Constant volume lines, on 7V-chart 68 Constant speed governing 218, 219 Contact condensers 268-278 Continental type boiler 370-372 Conventional diagram and card factors 125-128 Conversion of barometric readings, from inches mercury to pounds per square inch 265, 266 Cooling towers 294, 295 Corliss and other high-efficiency engines 196-215 Locomobile type 210-215 Non-detaching Corliss gears 201-205 Poppet valves 205-208 Trip-cut off Corliss 196-201 Uniflow engine 208-212 Corliss engine, trip-cut-off 196-201 Corliss gears, non-detaching 201-205 Crank end of engines 98 Cross-head and guides 107, 108 Cushion steam and cylinder feed 85, 86 Cut-off governing , 215 Cut-off ratio 128 Cycle, area on T<£-chart representative of work 73 • ^_ Complete expansion 55-58, 72 460 INDEX PAGE Cycle, incomplete expansion 58-60, 74, 75 Modifications for wet and superheated steam 73, 74 Of events in simple steam power plant 22 Theoretical, of steam turbine 228-231 Cycles, desirability of various, in engines 55 Cylinder, action of steam in 24 Condensation, methods of decreasing 89-92 Efficiency 139 Feed and cushion steam 85, 86 Ratios . 151-153 Cylinder and steam chest 101, 102 Decreasing cylinder condensation 89-92 De Laval impulse turbine 239-241 Delivered H.p., determination of 438-443 Density, specific, of dry saturated steam 38 Description and method of operation of D-slide valve 159-165 Design of nozale, steam turbine 231-237 Determination of clearance volume from diagram 131, 132 Determination of I.h.p 120-124 Determination of delivered H.p 438-443 Developed horse-power 137 Developed thermal efficiency 138 D-slide valve 159-195 Angle of advance 167 Angularity of connecting-rod 179 Bilgram diagram 169 Description and method of operation 159-165 Diagram for both cylinder ends 177 Exhaust and compression 175 Exhaust lap , 168 Indicator diagram from Bilgram diagram 180 Lead 166, 167 Limitations of D-slide valve 183-185 Piston positions 177 Reversing engines 185-187 Steam lap — outside lap 165, 166 Valve setting 187-195 D-slide valve engine, simple ■ 96-98 Diagram, Bilgram, for both cylinder ends 177 Bilgram, indicator diagram from 180-183 Indicator 24 Indicator and mean pressures for compound engines. 153-155 INDEX 461 PAGE Diagram (continued) — Indicator — combined 155-158 Indicator, conventional and card factors 125-128 water rate 86, 132-136 Diagrams from real engine 192, 193 Double acting engines 55 Double-flow condenser 270 Downdraft furnace 315 Draft apparatus 394-397 Chimneys or stacks 394-396 Mechanical draft 396, 397 Dry-air pump 274 Dry-saturated steam, total heat of 33 Specific density of 38 Specific volume of 36-38 Dry-vacuum pump 274 Dulong's formula— combustion 322, 323 Duplex steam pump 411, 412 Eccentric 160-165 Effective pressure, mean, methods of varying 215 Efficiency 52, 53 Cylinder 139 Developed thermal 138 Effect of temperature range on 75 Indicated thermal 138 Mechanical and thermal 137-140 Of boilers 387, 388 Relative 139 Elements — combustion 296 Energy- Conservation of energy, law of 2 Heat 2 Mechanical 2 Units of 3 Engine — Application of theory for an ideal to a real 54, 55 Compound, triple, quadruple, quintuple 148 Receiver type 149 Tandem and cross-compound 151 Woolf type 149 Desirability of various cycles 55 462 INDEX PAGE Engine (continued) — Double acting 55 Efficiency 52, 53 Heat quantities involved 50-52 Ideal steam 43-60 Operation of 42, 45 Operation of the real steam 77-80 Reversing 185-187 Steam — Classification — (1) On basis of rotative speed; (2) Ratio of stroke to diameter; (3) Valve gear; (4) Position of longitudinal axis; (5) Num- ber of cylinders; (6) Cylinder arrange- ment; (7) Use 92, 93 Clearance, volumetric and mechanical 84, 85 Crosshead and guides 107, 108 Cushion steam and cylinder feed 85, 86 Diagram water rate 86 Cylinder and steam chest 101, 102 Determination of initial condensation 86-89 Initial condensation 81 Losses, in real installations 80-84 Methods of decreasing cylinder condensation 89-92 Nomenclature 98 Principal parts 98-114 Bearings Ill, 112 Connecting rod 109, 110 Crosshead and guides 107, 108 Cylinder and steam chest 101, 102 Flywheels 112, 113 Frame 99, 100 Piston 102-106 Piston rod and tail rod 106, 107 Shaft 110,111 Re-evaporation in 83, 84 Rotative and piston speed 93-96 Simple D-slide valve 96-98 Throttling or wire-drawing . 82, 83 Work done by 46-60 Engines, Corliss and other high efficiency 185-187 Locomobile type 213-215 Non-detaching Corliss gears. 201-205 INDEX 463 PAGE Engines (continued) — Poppet valves 205-208 trip-cut-off Corliss 196-201 Uniflow 208-212 Entropy diagram 61-71 of liquid, vaporization, and dry saturated steam .... 61-63 7Vchart for steam 62-65 Complete 68-70 Constant quality lines 66, 67 Diagram for a real engine 136 Heat from 68 Quality from 65-68 Saturation curve 63 Superheating lines 63, 64 Volume from 68 Water line . 63 Entropy, diagrams of steam cycles 72-76 Equivalent evaporation, boilers 361 Excess air — combustion 286 Advantages and disadvantages 311 Excess coefficient 286 Exhaust and compression — Bilgram diagram 175-177 Exhaust lap 166, 168 Exhaust steam, utilization of, for heating buildings 400, 401 Expansion, adiabatic 58 Cycle, the complete 55-58. 72 the incomplete 58-60 Gain by, in compounding 151 -144 Ratio of, apparent and real 128-130 External latent heat of vaporization 31 Externally fired, return tubular boiler 373-382 Fahrenheit scale 11, 12 Feed-water heating 401 Open and closed heaters 401-408 Firing boilers by hand 332-336 Fixed carbon in coal 321 Flue gases from combustion of carbon 305, 306 Flywheel 97, 98, 112, 113 Regulation 216, 217 Foot-pound, definition 3 Forward stroke of engines 98 Foster superheater 393, 394 464 INDEX PAGE Frames of engines 99, 100 Front end of engines 98 Fuel calorimeter 323, 422-428 Fuels. 317-325 Commercial — Solid, liquid, gaseous 317, 318 Coal 318-320 Analyses 320-322 Calorific value of — Dulong's formula 322, 323 Fuel Calorimeter 323 Petroleum 323-325 Baume scale to express gravity 324 Calorific values 324, 325 Purchase of coal on analysis 323 Functions of boiler parts 327-329 Furnaces — and combustion 329-332 — Updraft and downdraf t 336 Gases, and vapors, steam 27 — Flue, from combustion of carbon 305, 306 Gaseous fuels 317, 318 Gauge pressure 39-41 Gearing and staging — turbines 241-248 Gears, Corliss, non-detaching 201-205 Generation of steam in real steam boiler 38, 39 Generation of steam or water vapor 28 Governing — throttle and cut-off 218 Coefficient of regulation 219 Constant speed 218, 219 Governor 97, 98 Regulation 217, 218 Governors — Pendulum. 220 Rites inertia 221-223 Shaft 220, 221 Grates, mechanical 336, 337 Gridiron valve '. 185 Guides and crosshead 107, 108 Hand firing — steam boilers 332-336 Heat 9 Absorption, reversal of process. 38 INDEX 465 PAGE Heat (continued) — Energy 2 Unit of 13 From T^-chart 68 Latent, of vaporization 30, 32 Internal and external 30, 31 Of liquid, q or h 31, 32 Of superheat 34, 35 Quantities in rectangular cycle 50-52 Quantity of 16 Specific 14 Total, of dry saturated steam 33 Of superheated steam 36 Of wet steam 33, 34 Value of elements and compounds 296-298 Heat, waste — in steam plant 399, 400 Feed-water heating 401 Open and closed heaters 402-408 Utilization of exhaust for heating buildings 400, 401 Heaters, feedwater, open and closed 401-408 Heine superheater 391, 392 Heine water-tube boiler 378-380 Horizontal, return, tubular boiler 373-376 Horse-power 17 Delivered, determination of 438-443 Developed 137 Hour, definition 18 Of steam boilers 384 Hydraulic analogy 26 Hydrocarbons, combustion of 308, 309 Calorific value of 309 Hydrogen, combustion of 306-308 I.h.p. — determination of 120-124 Impulse steam turbine 224-228 De Laval type 239-241 Inclined stokers 345-348 Incomplete expansion cycle 58-60, 74, 75 Indicated thermal efficiency 138 Indicator 115 Indicator diagram 24, 115-140 Atmospheric line 119 Conventional and card factors 125-129 466 INDEX PAGE Indicator diagram (continued) — Cut-off ratio 128 Determination of clearance volume from diagram. 131, 132 Determination of I.h.p 120-124 Diagram factor or card factor 126-129 From Bilgram diagram 180-183 Mean effective pressure 122 Planimeter 123 Ratio of expansion 128-131 Reducing mechanism 118 Scale of spring 118 Indicator diagrams and mean pressures for compound engines . 153-155 Combined 155-158 Indicator diagrams from real engine 192, 193 Inertia governor, Rites 221-223 Initial condensation 81, 82 Determination of 86-89 Injector, steam 413, 414 Inside lap, negative 168 Internal latent heat of vaporization 30 Internally fired, tubular boilers 367-373 Jet condensers 269-278 Joule's equivalent 14 Joule, the 3 Kinetic mechanical energy 8 Lap angle 166 Lap, steam 165, 166 Negative inside 168 Outside and exhaust 166, 168 Latent heat of vaporization 30, 32 Internal and external 30, 31 Lead 166, 167 Leblanc — Westinghouse condenser 276-278 Liquid fuels 317, 318 Liquid, heat of , q or h 31, 32 Entropy of 61 Limitations of D-slide valve 183-185 Balanced slide valves . 184 Gridiron valve 185 Piston valve 184 Riding cut-off valves 185 INDEX 467 PAGE Locomobile type of high efficiency engines 213-215 Locomotive type boiler 370 Low-pressure turbines 257-258 Matter 1 Law of conservation of matter 1 Units of matter 3 Mean effective pressure 122 Methods of varying 218 Mean pressures and indicator diagrams for compound engines . 153-155 Meaning of performance 419-422 Measurement of temperature 10 Measurement of vacuums 262-255 Mechanical and thermal efficiencies 137-140 Mechanical clearance, steam engine 84, 85 Mechanical draft 396, 397 Mechanical energy 2, 3, 7 Potential and kinetic 7, 8 Mechanical grates. 336, 337 Mechanical stokers 338-359 Mercury readings, conversion to pounds per square inch 265, 266 Mercury thermometers 10-12 Method of operation and description of D-slide valve 159-165 Mixtures, combustion of 309, 310 Moisture in coal 320 Molecular activity 9 Molecules 297 Natural draft, chimneys 395 Negative inside lap 168 Non-condensing plants 23 Non-contact condensers 278-282 Surface (Wheeler) 278-281 Non-detaching Corliss gears • 201-205 Nozzle design, steam turbine 231-237 Oil firing 358, 359 Open and closed feed-water heaters 402-408 Operation of simplified steam engine 45, 46 Operation of real steam engine 77-80 Outside steam lap 166 Outstroke of engine 98 468 INDEX PAGE Parallel-flow condenser 269-271 Parson's type turbine 251 Pendulum governors .....' 220 Performance, meaning of 419-422 Petroleum 323-325 Baume scale to express gravity of 324 Calorific values 324, 325 Piping, steam 417, 418 Piston, engine 102-106 Piston positions for Bilgram diagram 177-182 Piston rod and tail rod 106, 107 Piston speeds of steam engines 93-96 Piston valve 184 Planimeter 123 Plant, steam power 20 Plants, condensing, non-condensing 23 Poppet valves 205-208 Positions of piston for Bilgram diagram 177-182 Potential mechanical energy 7 Powdered coal stokers 354-359 Power and work 17 Power, unit of, horse power 17 Pressure, absolute , . 41 Gauge. . 39-41 Mean effective 122 Methods of varying 215 Pressures, mean, and indicator diagrams for compound engines. 153-155 Prevention of smoke 337, 338 Prime-mover 20 Principal parts of engines 98-114 Principal of condenser 266-268 Properties of steam 27 Proximate analysis of coal 320 Pump, dry air or dry vacuum 274 Vacuum " 282-288 Pumps, boiler feed 410-412 Purchase of coal on analysis 323 Quality from TV-chart 65-68 Constant, lines 66, 67 Quantity of heat 16 Rate, diagram water 86, 132-136 Rate of combustion in boiler furnaces 3£.9-361 INDEX 469 PAGE Rating of steam boiler 382-386 Ratio, cut-off 128 Ratio of expansion — apparent and real 128-130 Ratios, cylinder 151-153 Reaction type turbine 248-251 Receiver engine 149 Recovery of waste heat 399-409 Reducing mechanism 118 Re-evaporation 83 Regulation 216-223 Coefficient of governor 219 Constant speed governing 218, 219 Governors — Pendulum 220. Rites inertia 221-223 Shaft 220, 221 Kinds — flywheel and governor 216-218 Methods of varying mean effective pressure — Throt- tling and cut-off 218 Relative advantages of contact and non-contact condensers.. . 293, 294 Relative efficiency 139 Return tubular boilers, horizontal 326, 327, 373-382 Reversal of process of heat absorption 38 Reversing engines 185-187 Riding cut-off valve 185 Rites inertia governor 221-223 Rotative speeds of steam engines 93-96 Safety and strength of boilers 361-365 Saturated steam ; dry, specific volume of 36-38 Saturated vapor 31 Saturation curve, temperature entropy chart fcr steam 63 for compound engine cards 156 Scale 389, 390 Prevention of 390 Scale of spring, indicator 118 Scotch marine type boiler 372, 373 Separately fired superheaters 390, 391 Separators 413-417 Setting, valve 187-195 Shaft governors 220, 221 Shaft of engine 110, 111 Simple D-slide valve engine 96-98 470 INDEX PAGE Siphon condensers 266 Slide valves 184 Balanced 184 Gridiron valve 185 Piston valve 184 Riding cut-off valve 185 Smoke and its prevention 337, 338 Solid fuels 317, 318 Soot and scale, effects of, in boilers 389, 390 Specific density of dry, saturated steam 38 Specific heat 14 Specific volume of dry saturated steam 36-38 Speeds, rotative and piston, of steam engines 93-96 Spring, scale of, indicator 118 Stacks or chimneys 394-396 Staging and gearing, steam turbines 241-248 Steam, action in cylinder „ 24 Action of, on impulse blades of turbine 237-239 Boiler, generation of steam in 38, 39 Calorimeter 428-438 Consumption of steam turbines 252-257 Cushion, and cylinder feed 85, 86 Diagram water rate 86 Cycles, 7>-diagrams of, 72-76 Steam engine the ideal 43-60 Bearings Ill, 112 Classification 92, 93 Connecting rod 109, 110 Crosshead and guides 107, 108 Cylinder and steam chest 101, 102 Determination of initial condensation 86-89 Flywheel and governor 97, 98 Flywheels 112, 113 Frame 99, 100 Losses in real installations 80-84 The real 77-114 Initial condensation 81 Re-evaporation 83, 84 Throttling 82 Wire-drawing 82, 83 Methods of decreasing cylinder condensation . . . 89-92 Nomenclature of 98 Operation of 77-SO INDEX 471 PAGE Steam engine (continued) — Piston 102-106 Piston rod and tail rod 106, 107 Principal parts 98-114 Rotative and piston speeds 93-96 Simple D-slide valve 96-98 Steam, entropy of dry saturated 61, 62 Generation of 28-39 Heat of superheat 34-35 lap, D-slide valve 165, 166 Modification of TV>-chart for wet and superheated .... 73, 74 Properties of 27 Specific density of dry saturated 38 Specific volume of dry saturated 36-38 Temperature-entropy chart for 62-71 7>-chart complete 68-70 Total heat of dry saturated 33 Total heat of wet 33, 34 Vapors and gases 27 Wet, effect of 53 Steam injector 413, 414 Steam piping 417, 418 Steam power plant 20-22 Steam trap 417 Steam turbine (see Turbine) . Steam turbo-generators 258-260 Stephenson link gear 186 Sterling water-tube boiler 380-382 Stokers, mechanical 338-359 Chain grate 339 Inclined, overfeed 345-348 Powdered coal 354 Sprinkler 339 Underfeed 348-354 Strength and safety of boilers 361-365 Sulphur, combustion of 309 Sulphur in coal 301 Superheat, heat of 34, 35 total heat of 36 Superheaters — Built in 390, 391 Separately fired 390, 391 BabcQck and Wilcox 391 472 INDEX PAGE Superheaters (continued) — Foster 393, 394 Heine 391, 292 Superheating 31 Lines, on temperature-entropy chart for steam. . 63, 64 Surface condensers 278-282 Tail rod and piston rod of engine 106, 107 Temperature 9 Measurement of 10 Pressure relations 29 Temperature of combustion 312-314 Temperature rise 314 Theoretical 313 Temperature-entropy chart for steam 62-71 Complete chart 68-70 Heat from 68 Quality from 65-68 Volume from 68 T^-diagram for a real engine 136 T^-diagrams of steam cycles 72-76 Complete expansion cycle 72 Area of cycle representative of work 73 Modifications for wet and superheated steam . 73 Temperature range, effect on efficiency 75 Temperatures of vaporization 29 Theoretical cycle of steam turbine 228-231 Thermal and mechanical efficiency 137-140 Developed thermal efficiency 138 Indicated thermal efficiency 138 Thermometers, mercury 10-12 Throttle governing 215 Throttling or wire-drawing 82 Towers, cooling 294, 295 Traps, steam 417 Trip-cut-off Corliss engine 196-201 Triple expansion 148 Tubular boiler, horizontal return 326, 327 Turbine, steam , 224-260 Action of steam on impulse blades 237-239 Combined type 251 De Laval impulse type 239-241 Gearing and staging 241-248 INDEX 473 PAGE Turbine (continued) — Impulse 224-228 Low-pressure turbine 257-258 Nozzle design 231-237 Parson's type 251 Reaction type 248-251 Steam consumption 252-257 Steam turbo-generators 258-260 Theoretical cycle 228-231 Types of boilers 366-382 Babcock & Wilcox, water-tube 375-377 Continental 370-372 Externally fired, return tubular 373-382 Heine water-tube 378-380 Internally fired, tubular 367-373 Locomotive 370 Scotch marine 372, 373 Sterling water-tube 380-382 Wicks vertical water-tube 382, 383 Types of condensers — Contact 268-278 Barometric 271-278 Jet, parallel flow type 269-271 Siphon .' 276 Westinghouse-Leblanc 276-278 Non-contact — Surface 278-281 Two-pass or double flow 280-281 Ultimate analysis of coal 320, 322 Underfeed stokers 348-354 Uniflow engine 208-212 Unit of heat energy 13 Units of matter, energy and work 3 Updraft furnace 315 Utilization of exhaust steam for heating buildings 400, 401 Vacuum 262-265 Measurement of 262-265 Pump 282-288 Valve, D-slide — (see D-slide valve) 159-195 Setting. 187-195 CONTINUOUS AND ALTERNATING CURRENT MACHIN- ERY PROBLEMS. 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