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AIR COMPRESSION AND TRANSMISSION

McGraw-Hill Book Company

Publisters of Books for

Electrical World The Engineering and Mining Journal
Engineering Record Engineering News

Railway Age Gazette American Machinist
Signal Engineer American Engineer
Electric Railway Journal Coal Age
Metallurgical and Chemical Engineering

AIR COMPRESSION

AND

TRANSMISSION

H. J. THORKELSON

ASSOCIATE PROFESSOR STEAM AND GAS ENGINEERING
UNIVERSITY OF WISCONSIN

McGRAW-HILL BOOK COMPANY
239 WEST 39TH STREET, NEW YORK
6 BOUVERIE STREET, LONDON, E. C.
1913

*,* ;
- PC
eater),

CopyniGnt, 1913, BY THE

McGraw-Hitt Boox Company

18 MAY 14 lh

MoOlurg.

rchitecture. /F/7A

PREFACE

This text is designed to present in logical order the fundamental

_ principles dealing with the subject of the compression of air, and its

transmission through ducts or pipes, together with such examples as
will serve to illustrate their application.

It is hoped that the presentation will make clear the methods to
be followed in calculations dealing with air at various pressures,
and that students and engineers will be better able to appreciate the
advantages and limitations of the various systems of securing the
pressures desired, and of using air as a means of transmitting energy,
or for securing certain results which cannot be obtained under
normal atmospheric pressure.

The material offered consists of notes used for a number of years
by the author in his classes. He wishes to acknowledge his indebt-
edness to the many excellent texts on the subject which have been
published, notably those of Richards, Saunders, Hiscox, Harris
and Peele.

The fundamental formule are to be found in most texts on Thermo-
dynamics for Engineers. The author is particularly indebted to
Prof. C. H. Peabody’s text on this subject, and to the lectures of
F. W. O’Neill, F. D. Longacre, H. deB. Parsons and F. W. Towl,
given at Columbia University in a course on Applied Thermo-
dynamics of Air Compressors. The material on turbo-compressors
is taken from articles on this subject in recent numbers of the
Engineering Magazine by Franz zur Nedden. Permission to use
this material has been courteously granted by these authors and their
publishers.

The author is also indebted to the editors of Compressed Air
Magazine andto various manufacturers for cuts, and to his col-
leagues, particularly Prof. A. G. Christie and Mr. W. C. Rowse, for
assistance in preparing this text.

| Healers
April, 1913.

259994

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‘ f 7 os

INTRODUCTION

HISTORICAL ACCOUNT OF MECHANICAL USES OF AIR

The earliest writings describing mechanical uses of air are found
in a book entitled ‘‘Pneumatics”’ by Hero of Alexandria, published
about 200 B. C. An English translation of this by Bennet Wood-
croft indicates.a very complete knowledge of many mechanical
devices possessed by the Ancients, and shows various pumps,
Hero’s steam turbine and many remarkable uses of air as a means of
transmitting energy.

Fic. 1.—Hero’s device for opening temple doors.

One of the most interesting illustrations is a device for opening
temple doors by fire on an altar illustrated in Fig. 1. The altar,
E is hollow and a tube F passes through the altar and is attached
to a leather bag, K. Beneath this a small weight LZ is suspended
which is connected to the bag and to the pivots of the temple doors
as shown. Weights LZ and B are so proportioned as to normally

Vii

vil INTRODUCTION

close the doors. When a fire is lighted on the altar, the bag K
will expand under the pressure of the heated air below the altar and,
in doing so lift the weight L. Weight B then falls and causes the
doors to open. If the fire is extinguished, the air under the altar
will cool, contract, and the bag K will take the position indicated
and weight B will be dropped, causing the doors to close.

A somewhat similar device is also described in the same text
using heated air under a similar altar to force water from one cham-
ber into a pail, thus counterbalancing the weight and causing the
temple doors to open. Many automatons are also described, in
“which air is made use of to produce musical notes and cause water to
flow from vessels or objects when certain changes in
the mechanism take place.

A-very interesting experiment, illustrated in most
texts on physics, called ‘‘Hero’s Fountain” and in-
vented by the author of the book ‘‘ Pneumatics,”
consists of two globes, Fig 2, M and N, and a brass
dish, D. The dish D communicates with the lower
part of the globe N by a tube B and another tube A
connects the two globes. A third tube passes through
the dish D to the lower. part of the globe M@. This
tube having been taken out, the globe M is partially
filled with water, the tube is then replaced and water
is poured into the dish. The water flows through
Fic. 2.—Hero’s tHe tube B into the lower globe and expels the air

fountain, | Which is forced into the upper globe. The air being

compressed acts upon the water and makes it jet out
as shown. If it were not for the resistance of the atmosphere and
friction, the liquid would rise to a height above the water in the dish
equal to the difference of the level of the water between the two
globes.

Although a knowledge of this wonderful method of transmitting
energy has been known for centuries, it is only within comparatively
recent years that it has been used to any considerable extent in
practical work. Its modern use dates from the construction of the
Mt. Cenis tunnel completed in 1871. The work on this tunnel,
which is about 8 miles long, had progressed very slowly from 1857 to
1861, the tunnel headings having been drilled by hand labor with
an average advance in each of the two headings of about 1 1/2 ft.
per day. Machine drills driven by compressed air were introduced
and the speed rose to 4 3/4 ft. per day, and later when dynamite was

INTRODUCTION 1X

introduced, the cut was increased to 6 ft. per day. Sommeiller
deserves the honor for solving, in this work, many of the initial
problems of compressed-air production and use. The type of com-
pressor used is illustrated in Fig. 3. A natural supply of water
was used for compressing the air. The water was conducted in a
sluice A through a valve C, compressing the air which is in D and
forcing it into the reservoir H. When this was done the valve C
was automatically closed and the water in D allowed to escape and
be replaced by a new supply of air at atmospheric pressure. This

Fic. 3.—Sommeiller’s compressor.

was in turn compressed and forced into the reservoir E, giving not
a continuous but an intermittent flow of air. This compressor
furnished air at 80-lb. pressure, but only gave an efficiency of 50
per cent., that is, only one-half of the available energy was turned
into useful work.

Although the value of compressed air for machine drills for tunnel
work was clearly demonstrated in the building of Mt. Cenis tunnel,
it was some time before this was applied to mining work. One of
the earliest tests for mining work was made at the Calumet and
Hecla copper mine in Michigan in 1878, and the advantages in lower
costs and higher speeds were so clearly demonstrated that its use
for this work has since become almost universal.

One of the most important of modern applications of compressed
air is to be found in the braking of trains. George Westinghouse,
in 1869, designed his first “straight air brake,’ which was later
changed to the “automatic” type of air brake. This apparatus has

x INTRODUCTION

been improved and perfected to such an extent that its operation is
truly marvelous and its application world wide.

Railroad men were among the first to appreciate the uses of com-
pressed air in shop and structural work, and its application to
manufacture and other allied arts has since become so universal
that a mere recital of the modern applications of compressed air
would become tedious. One of the largest manufacturers of air
compressors has recently published a partial list of various ap-
plications of compressed air for which they have furnished compres-
sors. This list includes over sixty different industries, with a great
many different uses of compressed air in each.

While compressed air has many advantages over other systems
of transmitting energy, it has also certain disadvantages and limita-
tions which should be clearly understood. In order to appreciate
these, it is necessary to study in detail the nature and characteristics
of air and the fundamental principles governing its generation,
distribution, and application.

CONTENTS

PREFACE.

INTRODUCTION

CHAPTER I

CHARACTERISTICS OF AIR .

Air—Vapor in air—Free acs aye air—Effect af pressure on aiite
perature.

CHAPTER II

FUNDAMENTAL DEFINITIONS

Reeth Hine Dewersel cai perakiressA bsolute Renimeracire
—B.t.u.—Effects of heat—Energy in air—Specific heat—Specific heat
at constant pressure—Specific heat at constant volume—Real specific
heat—Apparent specific heat.

CHAPTER III

CHARACTERISTIC AND ENERGY EQUATIONS FOR AIR .

Boyle’s law—Law of Charles—Characteristic equation for perfects gases
—Numerical value of R—Weight of air—Relation between specific
heats—Work of isothermal change—Exponential change—Work of
adiabatic change—Relations between P, v and T for adiabatic and
exponential change—Computation of intrinsic energy.

CHAPTER] LV

GRAPHICAL DIAGRAMS

AIR

Construction of isothermal curves—Construction of exponential
curves—Heat added or taken away for isothermal change—Heat added
or taken away for exponential change—Difference between isothermal
and adiabatic compression—Temperatures due to adiabatic com-
pression—Work done by a compressor—Exponential compression—
Isothermal compression.

CHAPTER V

AT PRESSURES BELOW THE ATMOSPHERE : Ae
Venturi vacuum pump—Sprengle air Bee ueacurines vacuums—
Condenser pumps—Wheeler combined pump—Size of water and air
pumps—Steam cylinder size—U. S. Navy air pumps—Edwards air
pump—lIndustrial uses of vacuums—Salt evaporating effects—Con-
centration of liquids—Evaporation of cane juice—Vacuum cleaners—
Syphon.
xl

vii

xll

CONTENTS

CHAPTER VI

Arr AT Low PRESSURES

Uses of air at low Botan eer Me arise sais for ion ea eet ey ce
forges—Air for cupolas—Air for ventilation—Fans or blowers—Classi-
fication—Definitions—Measurement of draft—Fan efficiency—Flow of
gas through an orifice—Loss of head due to friction in ducts—Usual
velocity in ducts—Notation of symbols—Pipe losses—Rotary blowing
machines—Blower pressures and capacities—Power for rotary blowers
—Mechanics of the fan—Effect of outlet on capacity—Work required to
move a volume of gas—Design°of fans—Description of fans—Centrif-
ugal fans—Fan blast or steel plate machine—Housing—Cone wheel
fans—Turbine blast or “‘Sirocco” fan.

CHAPTER VII

PISTON COMPRESSORS

Action of piston Nes ernie tt: eid oF pictow compressor—
Effect of clearance—Methods of reducing clearance—Suction line—
Compression line—Wet and dry compression—Actual compression—
Cards for air compressors.

CHAPTER VIII

EFFICIENCIES AND ENERGY COMPENSATION

Volumetric Efficiency—Apparent volumetric Brrcienmt ernie dave
metric efficiency—Cylinder efficiency—Efficiency of compression—
Mechanical efficiency—Net efficiency—Blower efficiency—Energy com-
pensation—Hydraulic compensator—Lever compensator—Weight com-
pensator—Straight line compressor—Duplex compressor,

CHAPTER Ix

MULTI-STAGE COMPRESSION .

Advantage of multi-stage Compr eceionen Dress Teed ford various
stages—Intercoolers—Types of intercoolers—Cooling surface and
capacity—Intercooler pressure—Effect of clearance on volumetric
efficiency.

CHAPTER X

DETAILS OF PISTON AIR COMPRESSORS

Classification of valves—Mechanical tee ince valves pete
Effect of changing discharge pressure—Automatic valves—Valve area—
Forms of Poppet valves—Piston-inlet valves—Semi-mechanical valves
—Regulators, unloading devices, etc.—Belt regulator—Westinghouse
governor—Norwalk regulator—Combined governor and regulator—
Nordberg governor—Unloading devices—Clearance unloader.

CHAPTER XI

TURBO-COMPRESSORS .

Design of turbo- eerarecenre = alone blower The Rae blower—

113

CONTENTS

Cooling turbo-compressors—Cooling devices—Expansion of casing—
Runners—Balancing axial thrust—Balance by counter-position—
Balancing by diminishing back area—Balancing by balancing piston—
Stuffing-boxes—Coupling compressors—Rateau multiplicator—Mixing
blower.

CHAPTER XII

HypDRAULIC COMPRESSION OF AIR Aerie et
Trompe—Frizell’s renvecccm Baloche end Ncrahniass compressor—
Arthur compressor—Taylor compressor—-Taylor compressor at Magog,
Quebec—Taylor compressor at Ainsworth, B. C.—Taylor compressor
at Victoria Mine, Mich.—Phenomena of hydraulic air compression—
Losses of hydraulic compression.

CHAPTER XUHI

EFFECT OF ALTITUDE AND COMPRESSOR TESTS .
Effect of altitude on capacity—Effect of mlutuder on arent
between altitude and volume—Compressor tests—Mode of conducting
the tests—Results of the tests—Tests of plant No. 1—Test of plant
No. 2—Test of plant No. 3—Test No. 4—Summary.

CHAPTER XIV

RECEIVERS. MEASUREMENT AND TRANSMISSION OF COMPRESSED AIR
Receivers—Measurement of air and gases—Standards of measure-
ments—Volumetric meters—Velocity meters—St. John’s meter—
Venturi meter—Thomas meter—Meter comparisons—Pipe lines—
Dresser coupler—Hammon coupler—Pipe-line formule—Reheating—
Stoves.

Cin b LE Rea,

THE SELECTION AND CARE OF AIR COMPRESSORS
Available power—Valve gear—Size and type of eorneeeore Cone
pressed air explosions—Lubricating compressors—Cleaning valves—
Inlet connections.

APPENDIX A—ComMoN LOGARITHMS

APPENDIX B—-NAPERIAN LOGARITHMS

APPENDIX C—HyYGROMETRY .

INDEX

xlil

129

140

159

179

184
188
IQI

201

254

>; oF

So
74 .

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a a le vie
Py St, Sar i

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Lane

AIR COMPRESSION AND TRANSMISSION

CHAPTER I

CHARACTERISTICS OF AIR

Air.—Air is a mechanical mixture of several gases, principally
oxygen and nitrogen, its average composition by volume being as
follows:

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INGUCOUSAVADOLE eri eete ek Die ni cet ane Ween eh Okina 0.84
Ar pOuiceacit CAS merger i ee Nh oars Geeks 0.04

By weight it contains about 77 parts of nitrogen to 23 parts of
oxygen. There may be, in addition to the above, local impurities
in the atmosphere, the principal ones being ammonia and sul-
phuretted hydrogen.

The carbonic acid gas arises principally from the respiration
of animals and the processes of combustion, but, notwithstanding
the enormous continual production of this gas, the composition
of the atmosphere does not vary, for plants in the process of growth
decompose the carbonic acid, assimilate the carbon, and restore
to the atmosphere the oxygen, which is being continually consumed
in the processes of respiration and combustion.

Vapor in Air.—The vapor of water is always present in the atmos-
phere and when the air contains as much of this vapor as it possibly
can, it is said to be saturated.

The amount of vapor present in the air when saturated will vary
with the temperature, as shown in Table I, taken from the Smith-
sonian Institution Reports.

2 AIR COMPRESSION AND TRANSMISSION

TABLE I.—GRAINS VAPOR IN 1 CU. FT. OF AIR SATURATED WITH MOISTURE
(7,000 Grains=1 lb. Avoir.)

es 4 | I | 2 3 4 5 6 7 8 9
10 0.481 | 0.505 | 0.529 | 0.554 | 0.582 | 0.610] 0.639] 0.671} 0.704) 0.739
nme) 0.7767 (02816) 1028504] OF S060) 0.0408 OnOS Sir. 0321.07 Om teal cio meter tou:
20 Te235e) Le2O4 a) Las S on Le 416 fd 4 Ose eee S 5 0 mn Os ele OO 7imeLeye7 oi eos
30 T.035' || 2.022 192.1130) 26104 152527900102), 200\e2, 457920550127 040|n2n7A0
40 2.849 | 2.955 | 3.064 | 3.177 | 3.294 | 3.414) 3.539] 3.667) 3.800) 3.936
50 4.076 | 4.222 | 4.372 | 4.526 | 4.685 | 4.849] 5.018] 5.191] 5.370] 5.555
60 52745 | 5. O4L, |) O20A2) 46. C408 On5OS nis On7O27. O00 7 624 Tl 7A COTE 76
70 7.980 | 8.240\| 8.508 | 8.782 129.066) |.0.356) 02 6551) 0.002) 50.277) 10.007
80 10.934 |LE275 |Lt 026 111. O87 12.350) (12) 730) tan 27103).520/Ta Os TANenG
90 I4.790 |15.234 |15.6809 |16.155 |16.634 |17.124\17.626|18.142/18.671|19.212

This shows the number of grains of vapor present in each cubic
foot of air when the air is saturated. As there are 7,000 gr. in a

0 20 40 60 80 100,
Temperature, Degrees Fahrenheit

Fic. 4.—Water present in saturated air.

pound avoirdupois, these weights can be easily converted into
pounds. This data is shown graphically in Fig 4.

Free Air.—Free air is air at the pressure and temperature of the
atmosphere. This is a term used extensively in texts on compressed
air and in rating the capacity of a compressor.

It can be shown experimentally (Boyle’s Law) that if a cubic

CHARACTERISTICS OF AIR 3

foot of free air at sea-level (14.7 lb.) is compressed to a pressure
of 44.1 lb. by the gauge, or 4 atmospheres absolute, and allowed
to cool to the temperature of the atmosphere, the compressed air
will occupy one-fourth its original volume; if compressed to a pressure
of 58.8 lb. gauge, or 5 atmospheres absolute, and allowed to cool to
the temperature of the atmosphere, it will occupy one-fifth of its
original volume.

Figure 4 shows the weight of moisture which may be held in 1
cu. ft. of air at different temperatures, if saturated, and is true
no matter what pressure the air may be under.

It is evident then, that if a volume of free saturated air be com-
pressed into a smaller space and kept at the same temperature,
part of the vapor it originally contained must be precipitated for
the reason that 1 cu. ft. of air at a certain temperature can
only hold a definite weight of vapor when saturated, whether
compressed or at atmospheric pressure.

If saturated air is compressed to 5 atmospheres, or 73.5 lb. per
square inch absolute, and allowed to cool to atmospheric tempera-
ture, its volume will be reduced to one-fifth of its original volume
and 1 cu. ft. of compressed air will contain the moisture content
Obes CU eit, Of iree alr.

TABLE II—POUNDS OF WATER PRECIPITATED PER CUBIC FOOT OF COM-
PRESSED AIR AFTER COMPRESSION AND COOLING OF SATURATED
FREE AIR (Pressures)

Temp. Ga. 29.4 4AM foSOnO pLO2-07 147.0) 30725 1735 20] 2205.0
of Abs. 44.1 CO om (aay pO LOL giigole 2: 7A0u 7 | 2210.9
air Atm. 3 4 5 8 Er 26 5 I51

fo) .OOOI | .0002| .0003} .0005] .0007| .0017| .0035| .Oo105
IO .0002 | .0003] .0004| .0008] .ooII| .0027| .0055| .o165
20 .0004 | .0005] .0007) .0013| .0018) .0045) .0090] .0270
30 .0006 | .0008) .COII] .0020} .0028) .0070] .0140]} .0420
40 .0008 | .0012) .0016} .0028} .0040} .0100| .0200] .o600
50 .0012 | .0017} .0023| .0041| .0058] .0145| .0290] .0870
60 0016 | .0025| .0033] .0057| .0082] .0205| .o410] .1230
70 10023 de. CO84) 0045), .0070| JO1I4! .02351 30570" 22770
80 .003I | .0048| .0062] .o109] .0156] .0390] .0780] .2340
go SOOA2 she: GOOS4 =. O0GA |) .OT43! (O2TIh 20527. 1055) 23305

(For a further discussion on moisture in the air see Appendix C.)

4 AIR COMPRESSION AND TRANSMISSION

When reduced to the temperature of the atmosphere, the moisture
held in suspension per cubic foot of the compressed air cannot
exceed the moisture held in suspension per cubic foot of the free air,
and in consequence the remaining moisture will be precipitated.
This will represent for each cubic foot of compressed air a weight
of water equivalent to four times the weight held in suspension.

These weights have been calculated for various temperatures
and pressures as shown in Table IT.

Dry Air.—Air is said to be “dry” when water evaporates and
moist objects dry rapidly, and the air is ‘““moist”’ when they do not
dry rapidly and when the least lowering of temperature brings
about deposits of moisture. The terms are therefore relative ones,
but the expression “dry air,’”’? when used with reference to com-
pressed air, is usually understood as air containing less than half
the amount of moisture that is contained in “‘saturated”’ air.

Effect of Pressure on Temperature.—It is well known that air
can be made to expand by the application of heat. The altar
trick of the Egyptians illustrates this as does the modern hot-air
engine. Before friction matches came into general use, fire was
often produced by means of an air plunger-pump, which consisted
of a cast-iron barrel weighing several pounds with a bore about
3/8 in. in diameter, in which a steel piston fitted rather tightly.
The end of the piston had a small cavity for receiving a piece of
punk, and by pushing the barrel down on the piston the air in the
barrel was compressed and its temperature rose high enough to
ignite the punk. If heat is generated by compressing air, it is
natural to expect that if compressed air be allowed to expand
the temperature of the air will fall. This is exactly what does
happen in compressed-air motors, and if the compressed air contains
much moisture, the temperature may fall so low that this moisture
is frozen and collects as a frost in the exhaust pipe. Frost may even
collect to such an extent as to clog the exhaust pipe and stop the
motor. ‘The methods used to overcome this obstacle will be dis-
cussed later in detail.

The principal characteristics of air to be considered in discussing
its mechanical uses are: pressure, temperature, volume, weight
and humidity.

The relations existing between the temperature and humidity
have been considered, but before considering the other character-
istics mentioned, it is necessary to state clearly certain fundamental
_ definitions.

CHAPTER II

FUNDAMENTAL DEFINITIONS

Work.—Work is a force overcoming resistance, and is measured
in foot-pounds. A force of 10 lb. exerted for a distance of 4 ft.
represents 104 or 4o ft.-lb. of work, or a 1oo-lb. weight lifted
3 ft. represents 300 ft.-lb. of work, etc.

Energy.—Energy is the ability to do work and may be measured
in the same units. Energy may exist in a number of forms, for
example: a water-fall, heat, light, electricity, the wind, etc. The
source of all energy is the sun, but unfortunately this energy, as it
reaches the earth, is not in the most suitable form for all of the
work of the world. It is the province of the engineer to change
available energy into the desired form with as few losses and as few
changes as possible. Most of the energy required in commercial
enterprises is supplied by coal, but the burning of coal represents
the use of energy from the sun which reached the earth ages ago.
It is only a question of time when other sources of energy than
coal will have to be provided in greater abundance than at present
and the attention of scientific investigators is being called to the
importance of a more direct method of getting energy from the sun
and of using other available forms of energy with fewer changes
than are now necessary.

Heat.—Heat may be defined as a form of energy, without at
this time going into any discussion regarding its characteristics or
effects.

Power.—Power is defined as the rate of doing work. The engi-
neers’ unit being the horse-power, or 33,000 ft.-lb. of work per
minute.

Temperature.—Temperature is an indication of the direction
in which heat will flow if it has an opportunity to do so. That
is, heat will naturally flow from a hot to a cold body. Temperature
does not represent the heat energy that a substance contains.

Absolute Temperature.—Temperature in engineering work is

6 AIR COMPRESSION AND TRANSMISSION

usually measured on the Fahrenheit thermometer in which the
freezing-point of water at atmospheric pressure is 32° and the boil-
ing-point is 212°. As the temperature falls, the vibration of the
molecules becomes less rapid and the energy contained in any
substance decreases. That point at which the vibration of the
molecules ceases is called absolute zero. From a study of the
property of gases, it is evident that this point is about 460° below
the zero-point of the Fahrenheit scale. The absolute tempera-
ture, then, is the sum of the temperature Fahrenheit and 460°.

B.t.u.—Amounts of heat are measured in British thermal units.
A B.t.u. is the amount of heat required to raise the temperature of
t lb. of water from 63° to 64° F. The mechanical equivalent of this
is about 778 ft.-lb. This is usually represented by the letter J, and
its reciprocal, or as by A.

Effects of Heat.—If heat is applied to a substance, many of its
characteristics may change. Its pressure may change, temperature
may change; its volume, conductivity, elasticity, etc., may also
be affected by the application of heat. However, all these effects
may be classified into two groups: internal changes and external
changes. This may be represented by an equation as follows:
Heat applied = internal energy changes + external energy changes.

The internal changes may be represented in part by changes
of temperature which mean an increase in the velocity with which
the molecules vibrate back and forth. This energy expressed in
heat units may be represented by S. If a substance is of such a
nature that expansion takes place when heat is applied, then the
molecules must be separated farther apart against whatever mutual
attraction exists between them. This also represents an internal
application of energy and can be represented by the letter L. }

If external work is done, as in the expansion of any substance
at constant pressure, this work can be measured by the product of
the pressure in pounds per square foot and the change in volume
measured in cubic feet. The product in foot-pounds can be repre-
sented by the letter W, and its heat equivalent as AW.

If the heat supplied in B.t.u. is represented by Q, the equation
Q=S+L+4AW, in which each term is measured in B.t.u. may be
considered as a fundamental energy equation.

Energy in Air.—Air may be treated as a perfect gas or as a sub- -
stance in which there is no mutual attraction existing between the
molecules. In this case L=O, and the fundamental equation

FUNDAMENTAL DEFINITIONS © 7

when applied to air becomes Q=S+AW. That is, if heat is applied
to air the effect of that heat (Q) will be either to increase its tem-
perature (S) or to cause the air to expand and do work (AW), or
both. However, of the heat energy given to the air, the only por-
tion that can be stored up in the air itself (internal energy) will be
that portion which is used in increasing the temperature of the air.
In other words, the internal energy of air depends upon its temperature
alone.

This statement is very important and should be thoroughly
appreciated by the engineer working with compressed air.

At first thought, it does not seem possible that there is no more
energy in the air (internal or intrinsic energy) if at atmospheric
pressure than if the air is compressed and at the same temperature
as the atmosphere. This, however, is the case, as shown by the
above equations.

Although 1 lb. weight of air at a pressure of 1,000 Ib. per square
inch at the temperature of the atmosphere has no more internal
energy than 1 lb. of air at atmospheric pressure and temperature,
still the energy contained in the air under pressure is available for
use, while that under atmospheric pressure is not, for in the first
case the compressed air may expand, suffer a loss of pressure and
also of temperature, cool to a point below the temperature of the
atmosphere, and in that way give up a portion of its internal energy.
The greater the fall of pressure during expansion, the greater the
fall of temperature, and hence the greater the amount of internal
energy available for use.

Some engineers are under the impression that the energy used in
compressing air is actually stored up in the air. This, however, is
far from true, the internal energy in compressed air depends on its
temperature alone, and that part of this internal energy that may
be available for use will depend upon the fall of pressure and hence
of temperature that is permissible.

Specific Heat.—The specific heat of a substance in English units
is the amount of-heat required to increase the temperature of 1 |b.
of the substance by one degree, and is usually represented by C.

Specific Heat at Constant Pressure.—The specific heat at con-
stant pressure (Cp) is the amount of heat required to increase the
temperature of 1 lb. of the substance one degree F. the pressure
remaining constant.

Specific Heat at Constant Volume.—The specific heat at constant
volume Cy is the amount of heat required to increase the tempera-

8 AIR COMPRESSION AND TRANSMISSION

ture of 1 lb. of the substance one degree, the volume remaining
constant. .

As external work is done during a change at constant pressure,
it is quite clear that the former specific heat is greater than the
latterthatusy-C p. eC:

Real Specific Heat.—The real specific heat of a substance is the
amount of heat required to merely increase the temperature of 1 lb.
of the substance, one degree F. that is, this excludes any energy
that may be used in doing external or other work.

Apparent Specific Heat.—The apparent specific heat is the amount
of heat required to increase the temperature of 1 lb. of the substance
one degree F. including heat used in doing external or other work at
the same time. This is, therefore, usually greater than the real
specific heat. From the fundamental equation Q=S-++AW, it is ap-
parent that, if all the heat applied is to be used in raising tempera-
ture, AW =O and Q=S. This condition can only exist if there is no
change in volume.

For a perfect gas, and hence for air, the real specific heat is
equal to the specific heat at constant volume, that is, Co.

These specific heats are measured in heat units, but may be ex-
pressed in foot-pounds by multiplying by the mechanical equivalent
of a heat unit or 778 ft.-lb., or J. When this is done, the specific
heat is represented by K, that is,

AED and JC p=K »p.

TABLE III.—C, FOR AIR AT VARIOUS PRESSURES AND TEMPERATURES

Pressures in atmospheres and pounds per square inch
absolute
Temperatures
Fahrenheit

I 10 20 40 70 100
LAND eet A TLDs 294 lb. 588 lb. | 1,029 lb. | 1,470 lb.
212° On2202 0. 2389 0. 2408 0.2446 Oresr2 0.2583
Bo ONz275 O.2419 0.2465 O12 512 0.2773 0. 2986
—58° ON238O0umO. 2455 3. 0.25720 (One 755m mons TOMImOnd laa
— 148° 07238010) FO. 2585" 1. °O,, 2004 mt O es 00 MMO eA OT ie ce meee
— 238° O72424 Cl OAS TOS Ul GORSO4S Arr ee aia ie eee tent
—274° 02240791 AOL 4147 Spree. ie meee aeaegp noid eee ae sire ee

FUNDAMENTAL DEFINITIONS 9

The specific heat of air at constant pressure, (C p) is usually taken
as 0.2375 B.t.u., and the specific heat at constant volume, (Cy) as

; : G
0.1689 B.t.u. The ratio of these two specific heats, Ce. Or 1.405;
Vv

is frequently used as shown later in compressed-air calculations.

As a matter of fact, the specific heat of air at constant pressure
is not the same under all conditions, as it increases with increasing
pressures or decreasing temperatures, as shown in Table III given by
Prof. Linde.

CHAPTER III

CHARACTERISTIC AND ENERGY EQUATIONS FOR AIR

The principal characteristics of air to be studied are its pressure,
volume and temperature. Pressure may be measured in pounds
per square inch or pounds per square foot, and in the equations that
follow, absolute pressure is used. This is the sum of the pressure
of the atmosphere and the pressure shown by the gage. When
measured in pounds per square inch, the pressure will be indicated
by p. When measured in pounds per square foot, it will be indicated
by P. Volumes are measured in cubic inches or cubic feet, usually
thelatter. If the volume of a1 lb. weight in cubic feet is considered,
it will be represented by v. If the total volume of any weight in
cubic feet is considered, it will be represented by V. Temperature
may be measured on the Fahrenheit scale, and when so done is indi-
cated by ¢. If absolute temperature is considered, it will be repre-
sented by 7. TF-=t-+-460.

Boyle’s Law.— Boyle’s law for perfect gases, which was determined
by experiment, states that the product of the pressure and volume of
a perfect gas is a constant if the temperature is constant. ‘That is,
PV1= pod2= p3ds, etc., if the temperature is constant, or PiVi=
PoVeo, etc, Or P3,=P oto.

Law of Charles.—The law of Charles states that the volume varies

inversely as the temperature, if the pressure is kept constant. Or
Vie Ce : ;
m =7, etc., if the pressure is constant.
Leads

Characteristic Equation for Perfect Gases.—Combining these two
Piri uN Povo
be T > :

a constant is obtained. This constant will be indicated

equations, which were determined by experiment, or

Bal ie
Lael oe
as R and the above equation as Pyv,:=RT}.

Numerical Value of R.—The value of this constant (R ao 7) will
1

vary with the gas considered. For example: As 1 lb. weight of oxygen
at 32° F. and 14.7 lb. absolute pressure occupies 12.21 cu. ft., the
| 10

CHARACTERISTIC EQUATIONS FOR AIR 1a

14.7 X144 X12.21
reais eA
In the same way, as the volume of 1 lb. weight of air at this tempera-
ture and pressure is 12.39 cu. ft., the value of this constant for air is
14.7 X144 X12.39 |
460+ 32 mice
1V1 _ Pode

Weight of Air.—In the equation, “Pin i meeepierne dey Keven a,
this may be multiplied by the number of pounds of air, giving the
equation SA aA E ss, 3, in which w represents the number of
pounds of air. This equation is of great assistance in determining
the weight of air in a certain receiver. For example: If a receiver
3 {t. in diameter by ro ft. high, and having, therefore, a volume of
70.68 cu. ft., contains air at a temperature of 70° F. and a gauge pres-
sure of 143.3 lb., or 158 lb. absolute, the weight of air contained in
the réceiver if the pressure of the atmosphere is 14.7 lb. per square
Pr vie tso 144 70.08
53.321 §3-3X(460+70) 9°91”

Relation between Specific Heats.—If 1 lb. of air contained in a
vertical cylinder having a weighted piston above is heated, the
temperature of the air will increase and, as the pressure is kept con-
stant, the heat absorbed, may be expressed as the product of the
amount of heat required to cause a change of one deg. F. and the
actual change in temperature measured in degrees F. or as
CONC intB tiesony ke (haa) eit expressed in mechanical
units; where 7; and J, are the initial and final temperatures
respectively.

S, the energy required to cause the increase in temperature alone,
is Cy(T2—T}) in heat units; K,(T2—71), if expressed in foot-pounds.

W, the external work, must be P:(v2—21), where Pj represents the
pressure on the piston in pounds per square foot; v1 and v2 the initial
and final volumes of the air respectively in cubic feet, but as Pu=
RT this may be expressed as R(T2—T7)); that is, the heat energy ap-
plied, or 77830=K.(T2—T1)+R(T2—T;), in foot-pounds. But the
heat energy applied may also be represented by the expression,
K »(T2—T71), hence:

K »(T2—T3) =K,(T,—-T,)+R(T2—-T1)
K p=K.+R, or R=K p—Ky=778(Cp—Cv) =
778(0.2375 —0.1689) = 53.3
If expansion of air takes place in a perfectly non-conducting

value of this constant for oxygen will be

inch will be

12 AIR COMPRESSION AND TRANSMISSION

cylinder, Q of the equation Q=S+AW must equal O, and hence
W= — or W = —7785; that is, during an adiabatic expansion of air,

as it is called, the temperature falls and the amount of energy used
in doing work during the expansion will be given by the expression
K,(T1—T>2) in foot-pounds.

As no heat energy is given to the substance nor taken from it
during this change, all the work that is done must be done at the
expense of the internal energy.

The principal relations known to exist between Kp and Kv for
air are as follows:

K» he, eee

K,
K p=KytR=Kot53-3
K p—Kv=53-3

Work of Isothermal Change.—From the characteristic equation
for air, it is evident that if the temperature is a constant, PV =

?P 1

PV= RPV, = PVo

Fic. 5.—Isothermal change of air.

constant, or PiVi=P2V2=P3V3, etc. That is, if a certain volume
of air is compressed to one-half its original volume isothermally,
its pressure will be doubled, while if its volume is reduced to one-
fourth its original volume, its pressure will be quadrupled, etc.
This relation follows the path shown by the isothermal curve of
the chart, Fig. 5, the equation being that of an equilateral hyperbola.

In order to find the work done during an isothermal compression,
it is necessary to find the area under the compression curve drawn |
on a PV diagram or plane, this area representing the work done

CHARACTERISTIC EQUATIONS FOR AIR 13

expressed in foot-pounds. ‘The fundamental expression for any area
on a pressure—volume diagram is given by the expression,

Area=work = { PdV

with the isothermal curve PV =P1V,, hence pan, and
V2 Ve
w= { Pav ea | dV
V V
1 Vi
Ve.
hence, W =P1V, loge =
Vi

where P, represents the maximum pressure in pounds per square
foot and V; the minimum total volume of the compressed air in
cubic feet, V2 the total volume in cubic feet occupied by the air
before compression.

This also represents the work done by any number of pounds
of air expanding at constant temperature from an initial volume
of V; cu. ft. to a final volume of V2 cu. ft. If 1 Ib. of air expands
in this way the work done will be represented by the equation

P21 loge = or as P\vi= RT), this may be written RT; loge .
As the relation P1v,;=P2v2 applies to this isothermal change,
| Py
2 Ps
To illustrate the application of this formula, suppose it is required
to find the work done as 1 lb. of air expands at constant temperature

of 120° F. from a pressure of 150 lb. per square inch absolute to
30 lb. per square inch absolute the work done will be:

: ae
the equation may also be written Pv; loge si or RT, loge

ISOX144
53.3X(120+460) loge peat

It is evident that the ratios of pressures in pounds per square inch
is the same as the ratios of pressures in pounds per square foot.

53-3 X 580 Xloge 5 =53-3 X 580 X 1.6094, or 49,800 ft.-lb.

A table of logarithms to the base e, or hyperbolic or Naperian
logarithms as they are called, is given in Appendix B. This is
for numbers from 1 toro. If it is desired to obtain other logarithms
as for the number 0.12, this is equal to the loge of 1.2 minus the loge

14 AIR COMPRESSION AND TRANSMISSION

of ro, or the loge of ee In the same way the loge of 25 is the same

as the loge of (2.5 X10), or the loge of ro plus the loge of 2.5.

Work of Exponential Change.—When the equation of the com-
pression line is unknown, it may be represented by the equation
PV”=a constant, where z is the unknown exponent (Fig. 6).

F 1

Pyns PV, "= PV"

Fic. 6.—Exponential change of air.

PV"=P,V," and par Substituting in the formula for
work
Ve V2 V2
: d
w= frav,we | Davie gen | poner | V-"dV
Vi 1 1,
n artes Ln
pas ee Mee as Pi Va Pa Vere
P2Vo—PiVi_ PiVi-—PoVo2

I—n N—TI

P, represents the maximum pressure and P, the minimum pressure
in pounds per square foot, and V; the minimum total volume of
the compressed air in cubic feet and V2 its volume in cubic feet
before compression.

The above expression for work also represents the number of
foot-pounds of work that will be done by any number of pounds of
air expanding according to a change of pressure and volume repre-
sented by Pai =LfsV 5".

If 1 lb. of air expands in this way the work done will be represented
by the equation
ya P ate _ ROT 2).

n—I N—I

As Pyvy" =P vq", Pove=P 01 (°:)

CHARACTERISTIC EQUATIONS FOR AIR 15

From this it is evident that the equation may also be written:

Pv (*) ‘| Piven es) =|
SLY mes he > or as Tle
nN—-TI V2 n—I ) 1
As an illustration suppose it is required to find the work done
as 1 lb. of air expands according to the equation Pyv\)*=Po2""4
from an initial pressure of 180 lb. per square inch absolute and a

- volume of 1.2 cu. ft. to a final pressure of 15 lb. per square inch
absolute, the work done will be

oe ga | |
ee eres = I—0.083

ye Neem 180 AT
=77,760[1 —0.492] =77,760 X0.508 = 39,502 ft.-lb.

Work of Adiabatic Change.—If a change takes place in a cylinder
surrounded by non-conducting walls preventing energy in the form
of heat from entering or leaving the cylinder, this change is called
adiabatic. ,

During an adiabatic expansion of 1 lb. of air, the work done may
be represented by the equation Ky(71—T+2), as for such a change
AW =—S or W=—778S, and S=C»(T2—T)).

As Pa Rt and Po,=RT >
K (P01 —P 202)

K»(%i—T,)= R

but R=K p—Ky

KyPii—Pod2) _ P1i1— P22
GOK Kp

Kee

This expression must equal the expression for work under an ex-
ponential curve, viz.:

then Ky (T,—T2) =

Pi1—Pove
n—I
hence for an adiabatic change of 1 lb. of air, the exponent in the

equation P11” =P v2” must equal —; that is,

Ky
Kp Kp
Py, Ke =PypKo and for w lb.
Kp Kp

PV Ko PV Kk» =PsV; acae

16 AIR COMPRESSION AND TRANSMISSION

Relations Between P, v and T for Adiabatic and Exponential
Change.—This shows the relations existing between P and v for
an adiabatic change. In order to find the relations between Pand
T or v and T for such a change, it is necessary to turn to the charac-
teristic equation Pu=RT, or

P44 Pods P3033

Ty ah T 2 cs T3
Ey peels
then PRE
But as P4013-49 = Povo! 495
i ‘- 2) 1+405
then es &
: VN om Oal (**) 0-405 Ty
Equating ("*) Sais and : Te
vas © 1.405
As Fat
Pi) van ey ie nao gph
ened Bee teeth Vo . h ) 1.405 2
then & = ence : T,
(a 28825 Tae (*) 0.405
Or, P, te -

In the same way for changes represented by the exponential
equation P41” =P2v”.

("2 pieaae eb S eee
V1 CoN De ae P» RS

Computation of Intrinsic Energy.—These equations enable
calculations to be made of temperature changes during adiabatic
compression and are used in calculating the heat curves of the chart,
Fig. ri.

During an adiabatic expansion of air all the work that is done
must be doneat the expense of the intrinsicenergy. This can be shown
from the equation Q=S+AW, which becomes O=S+AW for an
adiabatic change, for from the definition Q must equal zero. This

being true, Weare or a measure of the work done would give also

CHARACTERISTIC EQUATIONS FOR AIR 17

a measure of the change in intrinsic energy. That is, if the initial
pressure and volume of w Ib. of air is known, the equation
Eee overeat) too VOW tet 1h oP

ToL IeA0 Saat 0.405

will be a measure of the intrinsic energy available. In this equation
P, and P2 represent respectively the initial and final pressures in
pounds per quare foot.

If v; and ve be used representing the specific volumes, that is,
the volumes occupied by 1 lb. weight, then the equation given
represents the work done and the change of internal energy of 1 lb.
weight only; but if V;and V2 be used representing the initial and final

volumes occupied by all the air concerned in the expansion, the
PiVi—P2V2 Fhe ° ° : °

gee _will represent the entire amount of intrinsic
energy available for use in expansion of w lb. from:pressure P; and
volume V; to a pressure P, and volume V2.

If P; and V; are known as well as Po, the final pressure, V2 may be
calculated if the equation of the expansion line is known. ‘This
equation will be

equation

P1V 434% = PoVo! 4

the equation for an adiabatic expansion.

P\Vi-—PoV2
0.405
cannot be used for calculating a change of internal energy. In
order to obtain the amount of internal energy contained in air it is
merely necessary to assume an adiabatic expansion to infinity.

The area under such an expansion curve is finite and amounts
PiV,

If the expansion is not adiabatic, the equation

to ate For example, the internal energy contained in 1 ID. of air
405
at atmospheric pressure and 32° F., with a volume of 12.39 cu. ft. is
P40; RT, ;
eee. OF , which is
0.405 0.405
TBE r 53:2%493 or 64,730 ft.-lb.
0.405 0.405

It is, of course, impossible to obtain this amount of energy from a
pound of air, as it is impossible to secure its complete expansion to
absolute zero of pressure.

CHAPTER IV

GRAPHICAL DIAGRAMS

Construction of Isothermal Curves.—It is frequently necessary
or desirable to construct graphically compression curves represent-
ing isothermal and adiabatic changes of air. A method of con-
structing the isothermal curve is shown in Fig. 7, in which O rep-

ie a 1 4 b

Fic. 7.—Graphical construction of equilateral hyperbola.

resents the intersection of the coordinates of a pressure-volume
plane. If an isothermal line is to be drawn through point a, con-
struct horizontal line a—b and vertical line a—c, as shown, then
draw any diagonal line as o-1 and complete the rectangle a—1—2-3;
3 isa point on the required curve. In thesame way, other diagonals,
as o-4, may be drawn, and the rectangle a—4-5—6 constructed,
giving point 6 as another point on the required curve, etc.

Another method of constructing this curve is shown in Fig. 8.
Assume that it is required to draw the isothermal] line for air through
point a. A diagonal line, as 1-2, may be drawn through this point
and the distance 2-b made equal to 1~-a; 0 is a point on the required
curve. In the same way line 3-4 may be drawn through 0 and the
distance 4—c made equal to the distance 3-0, giving another point c
on the required curve, etc.

In the equation PV”=P,V1"=Po2V.", etc., the value of the
exponent may be found, if the values of P and V for any two points

GRAPHICAL DIAGRAMS 19

are known, by taking logarithms of both sides of the equation,
PiVi"=P2V 2"

log PitnXlog Vi=log Pe+n Xlog Ve, and from this
_ log Pi—log Pe

"log Vo—log V1

Fic. 9.—Graphical construction of exponential curve.

Figure 9 represents graphically a curve whose equation is P}V,"=
P2V,” in which the numerical value of 7 is 1.405. Curves of this
type may be classed as exponential or logarithmic curves. A
simple method (Brauer’s) of constructing such curves graphically
is given below, together with a development of the equations used.

Construction of Exponential Curve.—Brauer’s method of con-
structing an exponential curve may be illustrated by assuming any
two points, as A and B, Fig. 9, of such a curve and drawing lines
through both points perpendicular to both axes. Through C and £

20 AIR COMPRESSION AND TRANSMISSION

draw lines making an angle of 45 degrees the withaxes. D represents
the intersection of CD with BH produced, and F the intersection of
EF with AL produced.

Connecting points D and F with O will give the two angles DOP
or 8 and FOV or a. In order to determine the relations between
these angles, the following demonstration is given:

PPG = Pe GH = eee

DH =P, tan B

P, =P ,(1+tan P) (1)

Vez =s V,ztGB= V,tLE= V,+tiPF

ine = Ve4 tan @

Ve =Va(ittan a)

Ve=V4(rt-+tan a)” (2)

Multiplying equations (1) and (2)

PzV%(1+tan a)" =P,V3(1+tan £)
but, Puvi Pave

(rt+tan a)"=1-+tan 6

tan S=(1-+tan a)"—1

This shows the relation between the angles § and a in terms of the
exponent 7.

For convenience, it is customary to make tangent of angle a=
0.25. Tan ? has been computed for various values of 7 as follows:

n Tang. 2
O.7 0.169 ®
0.8 0.195
0.9 C5223
iO On25
1.0646 0.268
TESS 0. 288
r25 Ou322
I. 333 0.347
T3239 0.358
TAt One7

If, for example, a curve is to be drawn through any point, as A
following the equation PiVi17>=P2V_!>, lay off angle VOF, or a,
with a tangent equal to 0.25 and angle POD, or , with a tangent
of 0.322; then draw AC and AL through A, and construct Jines CD
and FH, making angles CDH and FEL equal to 45 degrees, the in-
tersection of the horizontal line from D and the vertical line from £
will give point B as one of the required points of the curve. In the
same way, other points of the required curve may be obtained.

GRAPHICAL DIAGRAMS 21

Heat added or taken away for Isothermal Change.—The funda-
mental equation showing the effects of applying heat to air, Q0=S-+
AW, enables a determination of Q, the heat to be applied in the case
of expansion, or the heat to be taken away in the case of compression
in order to cause the expansion or compression line to follow a certain
exponential curve. If this required expansion or compression curve
is to be isothermal, S=O and Q=AW, that is, the heat to be applied
to secure isothermal expansion must equal the heat equivalent of the

. P1Vi loge Vi .
work done during the expansion, or — cae * heat units, and in
the same way the heat to be taken away, in order to secure isother-
mal compression, must equal the heat equivalent of the work done
during the compression.

Heat Added or taken away for Exponential Change.— Frequently

it is desired to secure a compression or expansion curve between the

Fic. 10.—Graphical measurement of change of heat energy.

isothermal and adiabatic, following the equation PaVa"=PsVo", as
shown in Fig. 10, in which z is less than 1.405. In this case Q=
S+AW. It has been pointed out that the mechanical equivalent
of the internal energy possessed by air is represented by the area
under an adiabatic curve continued to infinity and in Fig. 10 af
represents such a curve through point a and be such a curve through
point 6. W, or the external work done between a and 8, is repre-
sented by the area abcd, or 1+ 2. SX778, or the internal energy for
point a by the area 1+4, and SX778 for point b by the area 3+-4.
The change of internal energy in going from a to 6 in the case of
expansion will be area (3+4) —area (1+4), or 3—1, and as W=
1+2, the mechanical equivalent of Q must equal SX778+W, or

22 AIR COMPRESSION AND TRANSMISSION

(3-1) +(1+2), or (3-+2). That is, the mechanical equivalent of
the: heat energy to be added during expansion from a to b, or
taken away during compression from 6 to a, in order to follow a
certain curve on the PV diagram, will be the area bounded by the
curve ab and two adiabatics from the ends of the curve to infinity,
PuVa—PoV5 Pov P pV,
n—I By 0.405 - 0.405

Difference between Isothermal and Adiabatic Compression.—

The difference between adiabatic and isothe:mal compression is

or,

367.5

3528
333.1
3234

EEE

CATR

aang
ae NSE SN NIN

ved a

Sl Orme x
af COCCI el Ks
gms a EOLA NRE
BREE SCOUDSENEN:
ee gel LECCE
eo CELLET ALIS NINE
eee TILLELLLELLLL ALAN NINE
gree HCL TAIN De NINE
ep ALESIS CURSES:
breeg CILCCE EEC Net NN e
serge HCL CTT TTS SEINE
deo F ICCC SSE SSE
3 1323 10 Ps S ; °D
em eC LCLLEL LLC NEN
o 102.9 o vie PS PRR NS EPS ‘a
2 ore COSC Ge seel:
pes Cee TTS ie ESSE
58.8 ears Ss 5
441 4 Hee eee 3
29.4 CREE BEE EE i
)4e/ ec Stipes Senses
e Haan 0 76 Seah
02 ean niet fen eae eres

Fic. 11.—Temperature change due to adiabatic compression.

illustrated by the two curves on the PV plane at the left of Fig. 11,
which shows that while an isothermal compression of free air to half
its original volume will raise its pressure to 2 atmospheres, or 14.7

GRAPHICAL DIAGRAMS 23

lb. gauge at sea-level, the same reduction of volume adiabatically will
raise its pressure to 2.82 atmospheres, or 26.75 lb. gage. Jf the re-
duction of volume is to 0.2 of the original volume, the pressure for
isothermal compression will be 5 atmospheres or 58.8 lb. gage at sea-
level, and for adiabatic compression 8.88 atmospheres or 115.83 Ib.
gage.

Temperatures due to Adiabatic Compression.—Adiabatic com-
pression is always accompanied by an increase of temperature fol-

0-405 0-2882
a FZ) . ots Gs). This shows the
ratio of the final absolute temperature to the initia! temperature.

The right-hand part of Fig. 11 shows the resulting temperature
Fahrenheit for adiabatic compression with initial temperatures vary-
ing from 0° to 100° F._ From this it is evident that adiabatic com-
pression to 0.2, the original volume and consequently 8.8 atmos-
pheres, will approximate a final temperature 510° F. if the initial
temperature is 60° F.

Very high temperatures are to be avoided in compressing air as an
explosion may result if the temperature is sufficient to ignite the
volatile matter contained in the lubricating oils used. In addition
to this, the energy represented by the high temperature will soon be
dissipated by radiation. Isothermal compression requires removal
of heat energy during compression, and this cannot be accomplished
satisfactorily with piston or fan compressors operating at modern
speeds. Because of these conditions, high compression is secured in
modern compressors by compressing by stages and cooling the air
between stages. Compression in single-stage compressols Is usually
accompanied by such cooling as can be secured by water-jackets or
other means, but with the speeds required the usual effect is to se-
cure a compression curve with an exponent m varying between 1.25
ance. 32" |

Work done by a Compressor.—The work done in a machine, which
draws in air, compresses it, and then discharges the compressed air,
can be calculated, it the suction and discharge pressures are known
and the character or exponent of the compression curve is given.

Exponential Compression.—Let Fig. 12 represent such a series of
changes for any compressor, in which the effect of clearance is dis-
regarded. In this diagram, which is somewhat similar to the indi-
cator card from a piston compressing cylinder, d—a represents the
intake of air at a pressure of p2 Ib. per square inch, a—d represents
the compression of this air from p2 to pi Ib. per square inch following

lowing the equation

24 AIR COMPRESSION AND TRANSMISSION

the compression curve p1V 1" = p2V 2” anb-c represents the discharged
of the compressed air at a pressure /; lb. per square inch.

The area enclosed by the lines d-a—b-c-d will represent the work
done if V represents the volume in cubic feet and if p, representing
the pressures in pounds per square inch, be multiplied by 144 to give
pressures in pounds per square foot.

—— eo Pimper oe

Fic. 12.—Work diagram of air compressor.

The required area will be area a-b-g-f plus area b-c-e-g minus
area a—d-—e-f,
boVo—paVa
nan alae

or 144(poVo—paVa) (+1) ) or —-144(oV>—PaVa)

Va is usually known in compressed-air calculations, but Vz is not
known directly, but as paVa"=poVo” this equation may be sim-

or +144 poVo—144 PiVa

n—-1 es
plified, for p»Vo=paVa (7) > or poVo= paVa &) m , andethe

expression may be written:

EDN (ere it | |
tees I
Suppose, for example, it is desired to ascertain the work required
to compress 2,500 cu. ft. of air from a suction of atmospheric pressure
or 15 lb. absolute to a gauge pressure of 100 Ib. per square inch, or
115 lb. absolute, following a compression Jine whose exponent is 1.3.
1-3-1

rhe) ‘ I.3 (=) j.2o0e
This will require 7S xX 244X 152,500 | me :

0-23
oy 4.33 X144X15 X 2,500 (7.6) — 4

or, 23,375,000 X (1.59—1), Or 13,780,000 ft.-lb.

GRAPHICAL DIAGRAMS 25

If the compressor is to have this free-air capacity of 2,500 cu. ft.
per minute, the horse-power required will be
13780000
33000

SOL ATO. p),

Isothermal Compression.—If the compression line a—b was an
isothermal line, following the equation paVa=poV», the expression
for the area would be

ie Vig
144 ov loge V+ Ve baVal, or 144 poVo loge +,

or, 144 paVa loge ae Or 144 X15 X2,500 loge a

a
Or, 5,400,000 loge 7.66 = 5,400,000 X 2.036 = 11,000,000 ft.-lb.
If the compressor is to have this free-air capacity of 2,500 cu. ft.
per minute, the horse-power required will be

I I000000
33000

Or, 233 hep:

These calculations show the advantage of having the exponent
of the compression line as low as possible, or in other words keeping
the temperature of the air during compression from rising. The
advantages and methods of attempting this are discussed later.

The effect of valves, clearance and friction on the required horse-
power is considered later in discussing piston compressors.

CHAPTER V

AIR AT PRESSURES BELOW THE ATMOSPHERE

A study of the properties of air, and of its applications would
not be complete without reference to at least a few of the uses of
air at pressures below the atmosphere.

For purposes of experiment and for laboratory uses, these low
pressures are usually obtained by means of the familiar air pump.

Venturi Vacuum Pump.—Another method of securing these low
pressures is by means of a very simple hydraulic air ejector or
‘“‘venturl vacuum pump”’ as it is sometimes called.

This convenient instrument for quickly obtaining an approximate
vacuum depends on the principle that a fluid passing at a high veloc-
ity through a converging and diverging nozzle in which the curves

conform to the shape of the “vena
contracta”’ of a jet from an orifice,
will produce an approximate vacuum
at a point nearest its greatest con-
traction and if an air chamber is
connected through an orifice at
this point the air will be drawn
into the jet and a very good
vacuum formed in the chamber.
In the sketch shown in Fig. 13,
tube A may be connected by a rubber hose to a faucet. The con-
verging-diverging tube through which the water is forced is shown
at D. Tube C, which is connected with the chamber from which
the air is to be exhausted, has a check valve E and is connected
to the smallest diameter of the nozzle. It has been found that
better results are secured when a baffle F is introduced into the dis-
charge pipe B, as shown.

Sprengle Air Pump.—For a more perfect vacuum than the air
pump or the hydraulic air ejector the Sprengle mercurial air pump
is used. This pump depends on the fact that if mercury is forced
through an inverted U tube the mercury going over the bend will

Fic. 13.—Hydraulic air pumps.

AIR AT PRESSURES BELOW THE ATMOSPHERE 27

exhaust the air from any chamber that is connected to the U tube
at the top of the bend.

This pump is used for exhausting the air from incandescent
lamp globes and remarkably low pressures are secured with it.

Measuring Vacuums.—Although the normal pressure of the
atmosphere at sea-level is 14.7 lb. per square inch and pressures
above that are designated in the same units, pressures below the
atmosphere are not usually so designated but instead are expressed
in inches of mercury. If a U tube at the sea-level is filled with
mercury and one end connected with a perfect vacuum while the
other is in contact with the atmosphere, the mercury will rise to a
height of 29.92 in. above the level of the mercury in the leg that is
exposed to the atmosphere (1 in. of mercury=o.49 lb. per square
inch). Consequently a vacuum gage indicating 20 would represent
two-thirds of a perfect vacuum or a pressure of about 10 lb. below
the atmosphere, that is, approximately 5 lb. absolute.

One of the most familiar uses of pressures below the atmosphere
is in a condenser for steam engines in which the back pressure of the
engine is reduced considerably below that of the atmosphere, thereby
increasing the power of the engine.

Condenser Pumps.—Condenser pumps! are of two kinds: cir-
culating pumps and air pumps, the circulating pumps being used
to force the condensing water through the condenser and the air
pumps for removing the condensed steam and air. In some types
of condensers, the condensed steam is removed by gravity, as in the
barometric type, and the air pump removes but air alone, being
in this case called a ‘‘dry-air”’ pump to distinguish from the “‘wet-
air’? pump, which removes condensed steam as well as air.

Wheeler Combined Pump.—As a circulating pump _ usually
lifts the water but a short distance, it is built as a tank pump, but
should it be required to lift water through a long line of pipe, as a
line to the top of a cooling tower, it must be made of heavier con-
struction. Because of the large quantities of water which it handles,
it is very frequently built of the centrifugal type. Fig. 14 shows a
combined air and circulating pump of the Wheeler Condenser and
Engineering Company. The steam cylinder is in the center, the
air-pump cylinder at the left and the circulating-pump cylinder at
the right. |

The circulating pump forms one end support for the condenser.
The water is discharged through A into one set of tubes and then

1 Pumping Machinery, Greene.

28 AIR COMPRESSION AND TRANSMISSION

it returns through B and the upper set of tubes to C, where it dis-
charges. The air pump forms the other support for the shell.
It takes the air and water from the condenser and discharges it

through D. The suction space F is connected to G.

et
SSSSSPSSSSUAY 4
Vee pies

S SASS SSS

Sixes
an

N
G,

rss) SS
jp PIE GELS z

Ve
EA suk NOSE

seal

ISspsmsssssJ

=} NS
GLLLLEL peer
Zz

Air and Condensed Steam Stearn Cylinder Circulating Pump

is)
LELLLN

NN CiLLL/ LLL,
WA
TSS N

Fic. 14.—Wheeler condenser pump.

Size of Water and Air Pumps.—To find the size of the water and
air ends of the pump, suppose that W pounds of steam per hour at
a pressure p are to be condensed. If 7 is heat of vaporization of
the steam, x its quality, ¢.° the temperature of the condensed steam,
and g the heat of the liquid, and if G pounds of water entering at
ti° F. and leaving at ¢.° F. are to be used, G is given by the equation:

5 W (q+%.r— ic)
dto—Qti
in which the subscripts of g indicate the temperature for which the

heat of the liquid is obtained.
If the number of revolutions or double strokes N are assumed,

the displacement of the water end will be

G Ibs. per hr.

D3 = oN cubic feet.

The air end of the pump is made in many cases of empirical
design. Some authors give ratios of volume displaced by the
pump per minute to the volume of the condensed steam or to the
volume of the low-pressure cylinder of the engine which is dis-
charging into the condenser. Several of these are mentioned.

AIR AT PRESSURES BELOW THE ATMOSPHERE 29

RATIO OF AIR CYLINDER DISPLACMENT TO LOW-PRESSURE CYLINDER

Single-acting vertical pump surface condenser........ Ios 13
single-acting ‘vertical pump jet condenser.:..-....-... 1: 9
Double-acting horizontal pump surface condenser..... ae a
Double-acting horizontal pump jet condenser......... be aes
Double-acting horizontal pump-compound engine sur-

PAGEECMTICLCLLSCLa writ Mer ase nak tag ao Tt 26
Single-acting horizontal pump-compound engine sur-

POE SECON COSCL etme Ee kaie (in adn Mee. hehe oe Read STO

RATIO OF AIR CYLINDER De ee TO VOLUME OF CONDENSED

SIIRIACOPCOLICL CHISEL Me Mine ret eee pe, tie ne Sn ee ee Te OG
Net COD CDsct me a Ete Pac ok a heehee Mee oe Ee SAO

This may be a satisfactory way, but it is better to estimate the
volume from the air probably present. Water usually contains
air to about one-fifteenth of its volume. This amount of air is
at atmospheric pressure fa. and it must be cared for by the air
pump at a reduced pressure. In addition to this there are small
leaks in the pipe line which.allow more air to enter. A small hole
will destroy the vacuum of the air pump. To find the volume
of air per minute the following formula will be used, allowing 100
per cent. for leakage.

= Ins I iy thy : :
V=(@X2.Xg) (<-) (bps) T, cubic feet per min.
p = absolute pressure in the condenser, pounds per square inch.

ps = vapor tension or absolute steam pressure corresponding to 7¢.!
T. = absolute temperature in condenser.
Ta = absolute temperature of atmosphere.

This equation shows the importance of making ps as much less
than p as possible. The terms p and ps; do not differ much, and
by taking the mixture of air and vapor on its way to the air pump,
through as cold a passage as possible, the term ps; is made smaller
and the denominator is increased, making V small. This is the
reason for the great advantage in a counter current for condensers,
and even in the condenser, shown in Fig. 14, the coldest water should
enter directly over the air-pump inlet so as to cool the mixture
going to the pump.

From the volume thus computed the displacement of the air
pump is given by:

Dar=sy cubic feet.

1 See discussion on Partial Pressures Appendix C.

30 AIR COMPRESSION AND TRANSMISSION

Knowing the displacements of these pumps a stroke may be
assumed, and from it the area determined.

A es sq. ft. for the water pump.

Aer=—¢ sq. ft. for the air pump.
Steam Cylinder Size.—The cards from the water end are shown in
the lower part of Fig. 15, while those for the air end are shown

above. The combination of these or the addition of them when

¢ Atrnosphere

‘N

n
N
N
N
N
N
N
N
N
N
N
N
N
Ne
N
N
N

7 S fe

Atmosphere ‘ ie
Ure
Fic. 15.—Indicator cards of Fic. 16.—U. S. Navy pump
condenser pump. cylinder.

reduced to the proper scale, on account of the difference in piston
area, will give the total work, and from this the area of the steam
cylinder may be calculated, if the mean effective pressure
(M.E'P.); be found fora given boiler pressure. Allowing 33 per
cent. for friction, which is made large to give certain driving power,
the following results: :

4 LEP. )arAap+(M-EP.)pA p
a iT.0O— 0.23) LUGE ta).

Smit.

U.S. Navy Air Pumps.—Separate air pumps are often used. Fig.
16 shows the air cylinder of a stream-driven pump used in the U. S.
Navy. This air pump is made with two air cylinders driven through
gears from a steam cylinder placed on one side of a pump barrel.

AIR AT PRESSURES BELOW THE ATMOSPHERE 31

The pump is of the bucket type with foot valves AA and head valves
at B. These with the valves in the bucket at C are all spring-con-
trolled metal valves. The foot valves aie placed on an inclined
partition for the purpose of making it easier to discharge the air when
the piston rises and forms a vacuum. The lip around the discharge
valve makes a dam and covers the valve with water. This makes
them air tight. The other valves are also flooded, since all of the
water on the bucket or that over the foot valves cannot be driven
out, as the valves limit the motion of the bucket. On thedown stroke
of the bucket the pressure in the space above it soon falls to a low
vacuum because it had been completely filled with water; this, then,
causes the valves to open and take air from the lower portion of the
cylinder. ‘The air in the water also separates and rises to the top of
the cylinder. Finally the bucket reaches the water below, and this is
driven through the valve openings which are uncovered. It is seen
that the air leaves first in this case. The water is struck by the
bucket surface and will cause considerable shock if the pump is run-
ning too rapidly. |

Edwards Air Pump.—To do away with shock and to decrease
valve resistance, the Edwards air pump, Fig. 17, was introduced.
In this air pump water and air enter
the space A at the bottom of the pump
which is made conical in form. The
piston B, which is driven from the
steam piston by two rods CC extending —
over the shaft and crank, is provided
with a conical bottom. As this piston
descends there is a vacuum produced,
so that when the top of the piston
uncovers the openings #, air enters
{from the space A around the cylinder
barrel, and as the conical bottom enters
the water in the bottom of A, this is forced around the curved passage
and discharged into the openings at HF. This continues even after the
piston starts up, as the momentum causes the water to continue its
motion. This discharge of water into the openings as the piston is
moving upward acts as a valve to keep the air from coming out as the
piston ascends. In a short time, however, the piston covers the
ports or openings E and then the air and water are compressed
until the pressure is sufficient to overcome the atmospheric pressure
on the head valves at H, which are flooded by means of a lip around

NOS

l
NS

<<)
a

Fic. 17.—Edwards air pump
cylinder.

O2 AIR COMPRESSION AND TRANSMISSION

the valve deck. The piston rods CC are carried through long-sleeve
stuffing-boxes so arranged that the point H, at which leakage could
occur, is water sealed, leaving only one stuffing-box at the plate K to
carefor. Thisis a simple matter.

Industrial Uses of Vacuums.—One of the earliest applications of
air pressure below the atmosphere is shown in a patent dated 1833,
for the preparation of leather by the evaporation of certain substances
in a partial vacuum, the object being to avoid intense heat. Water
at atmospheric pressure boils at 212° F. If the pressure is increased
above the atmosphere as in the ordinary boiler, the boiling-point
of the water is raised, and for the same reason when it is desired to
evaporate any substance at a temperature below its boiling-point
for atmospheric pressure it is merely necessary to put the substances
in a partial vacuum and its boiling-point is accordingly lowered.
This results in evaporation at very
low temperatures, a most desir-
able feature especially in the drying
of fruits, etc.

Air at pressures below the at-
mosphere is used for drying all
kinds of food materials such as
meat, fish, fruits, etc. Frequently
a solution of a gelatin sugar or
gum is used as a coating.

Vacuum processes are employed
for pickling and salting meats and
vegetables, evaporating fruits, re-
fining sugar and condensing milk.
Wood may be artificially colored and railway timbers impregnated
with preserving substances by means of a vacuum.

Salt Evaporating Effects.—Probably one of the most interesting
applications of the partial vacuum is found in the manufacture of
salt, the refining of sugar and the concentration of syrups, liquors,
etc., by what is known as the triple effect apparatus.

The apparatus in which this is done usually consists of two or
three ‘‘effects,’’ as they are called, almost identical in construction.
One of these is shown by Fig. 18 and represents one of a “‘train”’
of effects for extracting salt from brine by heating the brine solution,
thereby evaporating the water and leaving the salt as a solid deposit.

B is the heating chamber or section consisting of a series of vertical
flues, conical in section, in which the brine circulates and around

Fic. 18.—Vacuum manufacture of salt.

ALK AIP PRESSURES BELOW THE ATMOSPHERE 33

which the steam flows. This part is very similar in construction to
a vertical flue boiler, with the exception that the flues are conical
instead of cylindrical, to prevent deposits on the tubes.

Steam js furnished to B through pipe £ either from a boiler or
the exhaust of a steam engine, and after giving heat to the brine,
which fills the apparatus as shown, the steam is condensed and
drawn off. ‘The vapor from the brine in the first effect passes through
pipe / into the heating section of the second effect, gives heat to
brine in the second effect and in doing so is condensed. This con-
densation of the vapor produces a partial vacuum in the first effect,
thus lowering the boiling point of the brine in that effect and hence
aiding evaporation.

The vapor of the brine of the second effect is conducted to the
heating chamber of the third effect, imparts heat to the brine in that
effect, producing a partial vacuum as in the first instance. The
vapor from the third effect passes to an air pump. ‘This air pump
maintains a good vacuum in the third effect and in consequence
the boiling-point of the brine in that effect is very low indeed.

The vacuum in the second effect is not good as in the third, and the
boiling-point of the brine in that effect is a little higher. In the
first effect the poorest vacuum exists and in consequence the brine
here has the highest boiling-point of all and therefore requires the
most heat, which is supplied to it from the boiler or the exhaust of
the engine operating the air pump. It is because of the different
boiling-points of the brine in three effects that the heat of the vapor of
brine in the first effect is enabled to evaporate the brine in the
second effect and the vapor of brine in the second effect is able
to evaporate the brine in the third effect. In each one of the three
effects of this apparatus salt is being extracted, the solid matter
settling to the bottom of the chamber C from which it can be with-
drawn by means of the two valves without disturbing the operation.
The salt and some brine with it are deposited on a filter in the cham-
ber D. The salt is here washed and the brine below the filter is
returned to the evaporating chamber through H. The salt is then
removed and a new supply of brine introduced through G. In
this way the operations can be made practically continuous and
in many plants automatic.

In some of the salt machines, instead of having two valves for
withdrawing the salt without disturbing the partial vacuum, the
lower part of the apparatus consists of a pipe running down such a

distance that the partial vacuum will equal the hydrostatic head
3

o4 AIR COMPRESSION AND TRANSMISSION

and consequently the salt may be removed without the use of any
valves whatever.

Concentration of Liquors.—In the concentration of syrups and of
many liquors, it is highly important that the liquors should not be
heated too highly or they will be scorched. To avoid this, the liquor
is moved through pipes at a velocity so great that there is no oppor-
tunity for the syrup to become scorched.

In the concentration of liquors, a partial evaporation is secured in
one effect and the vapor of the liquor separated from the liquor |
itself. Both vapor and liquor are then introduced into a second
effect which has a lower pressure, and here the vapor from the first

o—>
<

aELer
=a)
ze
2)
2

F| |

Fic. 19.—Vacuum concentration of liquors.

effect gives heat to its own liquor and the liquor is still further
concentrated; the resulting liquor and vapor of this second effect
are separated as in the first instance and introduced into a third
effect where the concentration is carried still further. In some
evaporators this is continued in a fourth effect and a still further
concentration of the liquor secured.

Figure 19 shows an evaporator of the type just explained.
The operation is as follows: The steam which may be either the
exhaust from an engine or live steam from a boiler is led into the
cylindrical chamber through A. The liquid to be concentrated is
fed in through the tubes B and enters the evaporator in a small
but continuous stream and immediately begins to boil violently,
becoming a mass of spray containing, as it rushes along the heated
tube, an increasing proportion of steam. The outlet of the tube C
being at a lower pressure than B, the contents are propelled through
the tubes at a high velocity, finally escaping into the separator D.
Here the steam or vapor with its entrained liquid is discharged with
considerable force against the baffle plates H, causing the liquid to
be separated from the vapor, the concentrated liquor being drawn

AIR AT PRESSURES BELOW THE ATMOSPHERE 35

off through a trap F, while the vapor escapes through G to enter the
second effect where its heat still further concentrates the liquid,
which is conducted from F of the first effect to the second effect,
entering through pipes similar to B. The liquid is led from the bot-
tom of the separator of the first effect into the coils of the second
effect and is further concentrated, passing in this way through the
entire system or “train.”

The volume of the liquid is being continually reduced as it passes
through these effects, and as the pressure falls in passing from one
effect to the next the boiling-point is lowered. That in the last
effect being the lowest of all, the required low pressure for this effect
is secured by a vacuum pump. This relative reduction in pressure
and consequently of the boiling temperature automatically adjusts
itself, no matter how many effects are used, thus effecting the boiling
of the liquid by the steam produced from the same liquid when in the
preceding ‘‘effect.”’

One of the advantages claimed for this system of evaporation of a
liquid in the form of a spray subjected 1 to heat under a vacuum is
that it receives the heat
quickly and is concentrated
slightly at one temperature,
then still more at the lower
temperature of ~ the: next
effect, and so on, thus re-
ducing the danger of over-
heating. The rapid move-

ment of the liquid aided by =, ;
the vapor which is moving a
in the same direction keeps )

the liquid in the form of a

spray, thus taking up very

quickly the heat given to it. i OTT,

Evaporation of Cane Juice. pre. 2o,—Vacuum concentration of liquors.

—The cross-section of a still

different type of evaporator is shown in Fig. 20. The liquid or cane
juice is introduced through A and is sprayed from holes in pipe B over
a series of steam-pipes. The partially concentrated liquor falls into
chamber C and is drawn from there by a centrifugal pump D and
forced into the next effect through a pipe similar to A and B of the
first effect. Steam for heating the first effect is introduced through
pipe E and, after imparting heat to the liquor, is condensed. The

36 AIR COMPRESSION AND TRANSMISSION

water falling to the bottom, as shown, is drawn off through F. The
vapor from the liquor that has been partially concentrated escapes
through G and is introduced to the heating chamber of the second
effect through a pipe corresponding to £ of the first effect. As the
pressure of the second effect is lower than the pressure of the first
effect, the condensed steam from this effect is also introduced into
the second effect, and, being at a temperature above that of boiling-
point of water at this lower temperature, it gives heat to this effect
and helps to evaporate the liquor in it.

Vacuum Cleaners.—One of the most recent applications of air
at pressure below the atmosphere consists of vacuum cleaners of
various types for removing dust and dirt fiom floors and walls,
furniture, etc., in buildings. Although many types of machines
for producing the 1equired vacuum are on the market, they may be
grouped under two heads, namely, portable and stationary types.
In the former type of vacuum cleaner, the machine is moved about
the room, drawing dust and air through a pump and discharging

Fic. 21.—Syphon.

into a cloth receptacle, from which the air can escape and in which the
dust is trapped. This type of cleaner has the advantage that it is
comparatively cheap, but has a disadvantage in that germs are
discharged with the air into the room, and from the hygienic point
of view this is considered objectionable.

The objection mentioned is removed in the second type of cleaner,
in which the machine is permanently located in the basement of
the building and vacuum pipes lead to the various rooms and floors,
to which the hose and cleaning tools are attached. This type of
machine is naturally more expensive, but finds considerable favor in
large office and hotel buildings.

The piston type of air pump is used as well as fans, bellows, and

AIR AT PRESSURES BELOW THE ATMOSPHERE 37

rotary blowers. This field of usefulness for air at low pressures has
increased to a remarkable extent, but the industry cannot yet be
said to be on a fixed basis; that is, there is still considerable informa-
tion needed regarding the proper suction pressures to secure the req-
uisite cleanliness without injuring the fabric of rugs: and hangings.

Syphon.—A discussion of the uses of air at pressures below the at-
mosphere would not be complete without some reference to the
syphon (Fig. 21). Bis an air chamber, C a water seal for the valve F,
D a funnel for filling the syphon and also for sealing valve K against
air leakage. After the syphon has been filled, valve K is closed and
G and H opened. This starts the syphon in operation. The air
that comes in with the water and through the joints of the pipe
collects in chamber B and may be discharged by closing valve F,
opening valve K, and filling chamber B with water. Then close
valve K and open valve F and any air below C will rise into chamber
B and the water will take its place without stopping the running
of the syphon.

CHAPTER VI

AIR AT LOW PRESSURES

Uses of Air at Low Pressures.—Probably the principal uses made
of air at low pressure are for cupolas, furnaces, blacksmith fires, for
conveying light materials, as shavings, etc. Avery large field for air
at low pressure is for purposes of ventilation for school buildings
churches, theaters, assembly halls, factories, mines and tunnels, etc.
Its use in this field dates back to the sixteenth century, and while for
many years very little thought was paid to it, to-day considerable
attention is paid to the subject of ventilation. In fact, no heating
system for home, school-house, factory or office building is complete
without some system for removing the foul air and replacing it with
fresh air.

To secure the necessary movement of air in buildings where the
number of cubic feet of room per person is a limited quantity, a
positive circulation is secured by introducing the fresh air at a pres-
sure a few ounces above the atmosphere into the room, or by drawing
the foul air from the room by means of an exhaust fan.

Compressors for Low Pressures.—The principal machines used
for moving air for ventilation and other purposes, either by pressure
or suction, are: the centrifugal fans or blowers, the positive blower
of the piston or rotary type, and the jet pumps from which are dis-
charged jets of steam or compressed air. The requirements for good
ventilation demand that large volumes of air must be moved at com-
paratively low velocity and pressure, which is not a favorable con-
dition for high efficiency and can in general be better satisfied by a
centrifugal fan or blower than by any other machine; it may also
be stated that the fan is comparatively cheap to install, is simple in
construction and possesses a fair efficiency.

Tables giving requisite information regarding fans for various
purposes can be secured from any of the fan manufacturers, and
engineering handbooks usually contain considerable data taken
largely from these catalogs.

AIR AT LOW PRESSURES 39

It is well to remember, however, that these tables are apt to over-
rate the capacity and under-rate the required power for operation.
Centrifugal fans are in use furnishing air at pressures varying from
1/4 0z. to 20 oz., and are constructed in all sizes, the largest, of course,
being used where large volumes of air are to be moved at very low
velocity.

. Air for Forges.—An article by William Sangster in the Transac-
tions of the American Society of Mechanical Engineers, in Volume 22,
page 354, gives the following approximate rules of the air required
for forges and cupolas. The maximum pressure required for forges
is about 4 oz. per square inch, the ordinary pressure about 2 0z.;
140 cu. ft. of free air per minute is ample and it is estimated that it
requires about 1/4 h.p. to furnish air for an ordinary forge. It is
customary to estimate that an exhaust fan for a blacksmith-shop
must remove four times the amount of air delivered at a pressure of
3/4 0z., and that to do this will require about 1/5 h.p. per forge.
Roughly speaking, if the number of forges is divided by 4, the horse-
power required to furnish the blast can be found, and if the number
of forges is divided by 5, the horse-power required to exhaust the
smoke can be found.

These exhaust fans run at a much slower speed than the pressure
fans and as the pressure of the exhaust air is much lower than the
blast the power required for their operation is less, although they
move four times the volume that the pressure fans do. In some
installations one fan does the work of forcing the blast and ex-
hausting the smoke, but as the requirements of a blast fan are so
different from those of an exhaust fan, such a combination is not
economical.

Air for Cupolas.—The air required to melt iron in cupolas may be
taken as 40,000 cu. ft. per ton of iron melted, and the horse-power
required as three-tenths of the number of tons to be melted per hour,
multiplied by pressure of the blast in ounces per square inch. It is
well to remember that these figures do not take account of losses in
the piping system. These results will, no doubt, fall far short if
the pipe system is poorly designed with sharp elbows, small diame-
LErsy tc:

Air for Ventilation.—In estimating air required for ventilation, the
data in Table IV is frequently used:

40 AIR COMPRESSION AND TRANSMISSION

TABLE IV.—AMOUNT OF AIR REQUIRED FOR VENTILATION

Allowable parts of carbonic Cubic feet of air required per person
acid in 10,000 of air
Gr COR Per minute | Per hour
5 100 6,000
6 50° 3,000
7 33 2,000
8 25 1,500
9 20 1,200
IO 16 1,000

TABLE V.—AIR SUPPLY FOR VARIOUS BUILDINGS

Cubic feet | Cubic feet

Air supply per occupant for Pe Aras Heche
Hospitals.2.toy chess Get oo ae le he ae ONTO TLOD. 4,800 to 6,000
High schools. . Le ee ate Oe Brae 50 3,000
Grtcrcaccis® te erty, hc BS en ln ce 40 2,400
Theaters and aesenibly nail eae Wears: De 1,500
Ghrarchee an eee ak a ae 20 1,200

TABLE VI.—AIR SUPPLY FOR VARIOUS ROOMS

Use choos Changes of air

per hour
Public.,waiting-room /.. 1.621, dee ol eee ee 4 to 5
Publicstoilets (ces 85h el eee 5 to 6
Coatvand Hocker-rooms.”.. 2... 4. ee eee ee 4-to 5
MMIUSOUIIS BS ence cite «Sch plc athe EEA oe eee rena 3to4
Offices;*piubltes: his iol a, see 4 to 5
Offices private < asl. cst. c Se ee ee ee ee 3 to 4
Public: dining-rooms, 5395 eee ee 4to5
Living-ToOMS WR eee cn Sao een eee pnd Ne eee 3 to 4
Libraries, public. . 5 safle: Bieta d seiaced, ewe TERee ee tues aa eee mT 4 to 5
Libraries, is ae ch SS ee ince Ee cae 3to4
Fuming cabinets fe eivctaica jaboratorres Lo a AOL eh one we ee ree 30 to 60

The following material on fans and blowers is taken from a lecture
by Mr. H. de B. Parsons, Consulting Engineer, delivered before the
Junior Class of Columbia University:

LECTURE BY MR. H. DE B. PARSONS 4]

FANS OR BLOWERS

‘A fan or blower is a machine for impelling gas, z.e., for producing
a current of gas. In the majority of cases the gas impelled is air.

“There are many purposes for which a fan is used, such as for
heating, cooling and ventilating buildings, either by exhausting air
from, or forcing air into, the apartments; for blowing the fire of a
forge or cupola; for creating artificial draft for fuel combustion; for
work pertaining to drying; for carrying away obnoxious gases and
discharging them at a point where they will not create a nuisance;
for carrying away grindings and waste products so that they may
not affect the workmen; for conveying light materials, such as saw-
dust and small particles, and permitting them to settle in dust
chambers; and for the circulation of air in mines and places where
explosive gases may collect.

“When fans are properly selected for their work they will give
satisfactory and economic results, and will require little attention
for maintenance.

“The conditions of pressure and density of the gas and of speed
and capacity of the fan govern the size, type and proportion of the
fan and its housing. These conditions are closely related, and all
affect the design that should be selected. Even moderate differ-
ences in the conditions of operation will have considerable effect
upon the power necessary to drive the fan. It therefore follows
that a fan should be designed for the conditions under which it is
to operate, and conversely, that a fan should be operated under the
conditions for which it was designed.

‘Fans are not economical machines to operate against high pres-
sures. In such cases a blowing engine or compressor will be the
Detter

Classification

“There are a number of types in use, but nearly all blowers and
fans can be classified under one of the follcwing heads:
(1) Rotary blowing machine.
(2) Disc, axial or propeller wheel fan.
(3) Centrifugal fan, either a fan blast or cone wheel.
(4) Turbine blast or high-speed centrifugal fan.
“Type (1) is a positive or displacement discharge machine, and
is a blower or exhauster.
“Type (2) is an axial discharge fan.
“Type (3) and (4) are peripheral fans.
‘All of the types can be used for exhausting or for blowing,
although some are less suitable for exhausting than others. There
is a material difference in the selection of a type for an exhaust

42 AIR COMPRESSION AND TRANSMISSION

machine, when pressures above the atmosphere on the discharge
side of the fan are considered. The disc fan makes a gcod exhauster
when a pressure above that of the atmosphere does not have to be
maintained on the discharge side, but when such a positive pressure
has to be maintained a centrifugal machine is the more suitable.

Definitions

“There are certain terms used in fan work which are recognized
as having specific meanings.

‘Fan Pressure or Draft——Fan pressure or draft means the
difference between the pressures on the suction side and on the
discharge side of a fan. The difference in pressure is expressed
either in ounces per square inch or inches of water.

“When a fan is used as an exhauster discharging into the atmos-
phere, there will be a partial vacuum on the suction side and slight
pressure on the discharge side. In this case the vacuum is expressed
as the number of ounces or inches of water below the atmosphete,
and the fan pressure or draft is measured by the difference. It is
just the same as if the suction were at atmospheric pressure and the
discharge at the same number of ounces above the atmosphere.

‘Fan Capacity.—Capacity means the maximum discharge of free
air froma fanin cubic feet per minute against a pressure cor-
responding to the speed of the tips of the blades. This condition is
satisfied in the case of a centrifugal machine when the velocity of
the gas entering the inlet is equal to the velocity of the inner edge
of the floats at inlet.

‘““Housing.—The casing in which a fan operates is called the
‘housing.’ It is made of metal, of brick, or of wood. Frequently
the fan is so set as to project into its foundation, and in such cases
the casing only covers the portion which projects above the founda-
tion. ‘The fan is then said to have a three-quarter housing. Of
course the inlet must be above the foundation or a free passage must
be provided to it.

“Free Discharge.—A fan is said to have free discharge when the
blast is free or unrestricted. This condition is maintained when the
total head is practically equivalent to the velocity head. The total
head is equal to the velocity head plus the friction head, and with
a free discharge head the friction head is practically zero.

“Restricted Discharge.—A fan is said to have a restricted dis-
charge when the blast is restricted by ducts or by pressure reservoirs.

“Free and Restricted Suction.—Similarly to free and restricted
discharge, a fan may have either a free or a restricted suction,
whereby the gas has either a free or unrestricted entrance into the fan,
or a restricted entrance caused by ducts or a reduction in the pres-
sure on the suction side.

LECTURE BY MR. H. DE B. PARSONS 43

“Coefficient of Contraction.—The ratio of the area of the vena
contracta to the area of the orifice is called the ‘coefficient of con-
traction.’

‘Coefficient of Velocity.—As the stream of gas passes the vena
contracta its velocity is somewhat increased, and the ratio of the
actual velocity to the theoretical velocity is called the ‘coefficient
of velocity.’ In a well-shaped delivery orifice this coefficient of
-velocity is not far from unity.

“Coefficient of Efflux.—The ‘coefficient of efflux’ is the product
of the coefficient of contraction and the coefficient of velocity.

“Volume of Discharge.—The volume of gas discharged by a fan
is a function of the product of the velocity of the gas times the area
of the outlet, times the coefficient of efflux.

“Blast Area.—The blast area of a fan is the theoretical area of
outlet whose coefficient of efflux is unity. The volume of discharge
is equal to the blast area times the velocity of discharge. Therefore,
the blast area equals the capacity divided by the velocity due to the
velocity head. The stream of gas issuing through an outlet is re-
duced in area, depending on the shape and character of the orifice.
This reduced area is called the ‘vena contracta’ and is usually at a
distance from the opening of about half its diameter. This vena
contracta is caused by the change in direction of the flow of the mole-
cules of the gas as they pass the opening.”’

Measurement of Draft

‘The measurement of draft, either static or both static and veloc-
ity pressures is obtained by noting the difference in level of a liquid
in the arms of a tube bent on the form of a U, of which one end is
connected with a proper tube to the space in which the draft is to
be measured. The liquid is generally water although for heavy
pressures mercury is sometimes used.

“There are different forms of gages which can be bought in the
market but for ordinary work the simplest forms are the best.
Some of these instruments are so made that they will give a con-
tinuous record, and for certain kinds of work these continuous
records are of considerable value.

‘““Anemometers are used for measuring the velocity of the gas.
Readings should be taken at many points in the cross-section of the
current, and even at the same points consecutive readings will not
agree. Multiple readings therefore should be made in order to aver-
age up these irregularities.

“Fan draft is always expressed in ounces per square inch or in
inches of a water column whose weight is equal to the ounces per
square inch. The velocity corresponds to this pressure, when the
friction head is zero.

AIR COMPRESSION AND TRANSMISSION

‘When the pressure exceeds two or three pounds per square inch
as is the case with many positive blowers, the pressure is then
generally expressed in ‘pounds per square inch.’

Corresponding to Various Heads of Water in Inches

TABLE VII.—PRESSURES IN OUNCES PER SQUARE INCH

Decimal parts of an inch

Head
in
inches
0.00) 0.8 ome O28 0.4 On5 0.6 OF 0.8 0.9
Osmaioe ae 0.00510, 12] O78 7 0,23 OF 209) Ones alo. 10 sO ntOm Oma
I 0.58 |)0703° 190.00 | Oa7s 470 Siw Ous7 IFO..03 4207 O0om tie O4mmimoG
2 EALO WAL Rea bay a Fee eT ee) MAT aTR COM Tec Op) at Ot ee
3 Te73UESFO |) Te BS |ATOL | 1.06) 2.02 1e2. OS 2b TA hee Om meee
4 2SAT VW 2e27 N24 2 I 2048 PO RAT eaAOO NO Oe OO aD. 72 mire ae james
5 2280 1-2.04 11-3500) | 3200) - 297 2elee rio 47 24 mg 20m) ec ns 5 amr ey
6 P3247 13-521 3258 |, 3404 53 FOS <7 6 Woh Olas ee 7a) ns cOe mt ReaD
7 AVOAY 4.10 144.10 4322 WAI 28 CA ea AO ed Calafate ee
89) 4024 4007 A273 104.90 W485 des Ot 4.07 15.03 195 Om es ae
9 5.20) 6:26 | 8.35 | a7 te Ae US lace sd ee se O081 SOO mms age
TABLE VIII.—HEIGHT OF WATER COLUMN IN INCHES
Corresponding to Various Pressures in Ounces perc Square Inch
Eres: Decimal parts of an ounce
sure
in oz
per
: ‘ : : ; é .6 : 8 O:
rae 0.0 Cut One 0.3 0.4 O25 fe) On ° 9
fe) Sf) OnE 1) O35.) -Ol'5 2 a2 0G LOT IAA OAM Walwo2 Taler sot ate et
I £272) (1.00) -2.08:) 92, 00 12. Al) 22800 aan 7 7 Od abet ke
2 3. AON 935631053. 81 }) 3.08 ATS (ASS etn5O eed. OF sd O4 sees mon
3 ST LOW eS Se Gal ee a7 EMS oe 106.1. 6503 1 Gi40 8102571 Oars
4 6.02) IvOON 727") 72 AA FROEN Aeon E00 sO aes terol sO MO wAS
5 8.65]. 8782 |G. 00," 0.17) 02344 0.5218 On00 sO. 004510503) Ione.
6 10.38 | 10. $5 (10.73 [TO-DO MIT O7 UT En20 hin A set COG tis 7 Fer reo
7 12.1} 12,28) @o. 40142 363 12, SON Os Lowel os 2 el oi Ou Toe
8 13.844 TA, 00 | T4479. 014. 260 14a Sa 14s 71 ae Sow 5 Obe LS. Sea
Oh 18.871 15 574 hFS) 02) 16-00 i1G220 10 48 nOuOe W 10. 70) 20.00) fy

DCECTORE BY MR. H. DE BY PARSONS 45

Fan Efficiency

“In the operation of a fan or blower there are certain losses
utes must exist, the principal losses being:

‘rt. Fluid friction and eddies caused by the movement of the
gases.

‘““9, Leakage of the gases backward through the fan or blower.
This is sometimes called the ‘slip.’

“2. Mechanical friction of the moving parts of the apparatus.

‘“‘ Generally speaking, these losses increase with the speed of revolu-
tion of the fan and also as the difference in pressure between the
suction and discharge sides increases.

haiine efficiency of a fan or blower is the ratio of the useful work
done on the air divided by the work required to drive the fan. Fans
are generally driven by steam engines or by motors, and frequently
the denominator of the efficiency ratio includes the work of driving
the engine or motor. Such an efficiency is really the combined effi-
ciency of the prime mover and fan.

“The efficiency of a fan wheel with a housing varies with the ratio
of diameter of inlet to diameter of wheel. The smaller this ratio
the greater will be the theoretical efficiency so long as the area of
outlet times the coefficient of efflux is not less than the blast area.

‘““When a centrifugal fan has to work against high pressures, it is
desirable, therefore, that the ratio of inlet to wheel diameter be small
in order to get the benefit of this increase in efficiency.”’

Flow of Gas Through an Orifice

‘““Gas flowing through an orifice does not obey the same law as
the flow of fluids. The reason of this is that gas expands from
the higher pressure to the lower pressure as it issues through the
orifice.

‘“‘Tmagine the gas in a reservoir R (Fig. 22) flowing from the short
cylindrical orifice of section a. Imagine that the reservoir is kept
supplied with gas so that its pressure remains constant. Suppose
that the division S represents a pound of gas.

‘““As the gas escapes through the orifice a, the pressure is kept
constant, and the work OA EQ has been done upon the gas. The
gas in expanding develops the expansive work EIMQ, EI being an
adiabatic curve.

“The outer pressure P, absorbs the work JVOM, and the balance,
AEIN, is devoted to accelerating the particles of the pound of gas
to a velocity v.

46 AIR COMPRESSION AND TRANSMISSION

“Hence the work AEIJN equals the actual energy of 1 lb. of gas
moving with a velocity of v ft. per second or

a
AEIN =(t lb.) X —
2g

Therefore v=/2g (AEIN) ft. per second
Taking the law of the curve FI as p101)"=fov2"=constant we
have:

fi
nN pe a
AEIN=""_ xX 144 piS ae (2) |

Fic. 22.—Flow of gas through an orifice.

“Letting m =1.405 the ratio of the specific heats and the exponent
in adiabatic changes of air

‘ihe a [28x25 X 144 p:S|1— @) be

v= 1794 | P15 E —~ C) oe) ft. per second

“Tt is found by taking the pressure at the orifice by a gauge G that
if the gas flows into another reservoir kept at the back pressure py»
the orifice pressure is identical with 2 if the latter is more than
about 0.581. That is, if p1 is r00 lb. per square inch absolute, any
value of the back pressure greater than about 58 lb. gives this
pressure to the orifice, but if the back pressure is the atmosphere
the orifice pressure remains 58 lb.

“Tt therefore follows that for ao 58 the velocity given by the
1
above formula exists at some point of the jet beyond the orifice at

LECTURE BY MR. H. DE B. PARSONS 47

a section a,> a due to the natural spread of the jet, while the velocity
at the orifice or throat of the jet is that given by the formula for

Loss of Head Due to Friction in Ducts

“The frictional resistance to the movement of a gas in a duct is
proportional to the surface of the duct. It is, therefore, directly
proportional to length and inversely proportional to diameter. It
also varies as the square of the velocity.

“Therefore, the ducts should be of ample area, or the power lost
in friction will be very great. Small pipes and high velocities
should be avoided.

“Tt is evident that after a certain size of duct is reached, any
further change in size or velocity of movement will only have a rel-
atively small effect upon friction loss. The limit, therefore, is
reached when the increase in space required and the cost will turn the
saving in friction into a loss from a commercial standpoint.

“Usual Velocity in Ducts.—In heating and ventilating work for
theaters, hospitals, churches and large buildings, the limiting
velocities usually selected are:

(a) In ducts leading from force fans—

In borizontalimaineducts 27 427. 2-1 ,000 It, per minute
imshorizontalimatnibranches #42. 72.428... a: 1,300 ft, per minute
In horizontal branches to risers............. 650 ft. per minute
tie WerticalsTisers eau woe es Sin ison Sa bes Ae 800 ft. per minute

(b) In ducts leading to exhaust fans—
EWevertical riscrs ea erty or aan Goo 1b per minute
Tnehorizonraduets tolans 2 sane tears £000 ttn per minute

‘“The frictional loss in ducts can be calculated from the formule
for the movement of fluids. In addition to the friction loss of head
caused by the passage of a gas through a straight duct, there is a
loss at each bend or change of section.

“In order to overcome these friction losses, it is necessary that the
pressure at the fan end of the duct should equal the sum of the
pressure desired at the open end of the duct, and the pressure
necessary to overcome the losses in frictional head.

“Tn all the following formule the following notation has been used.

Notation of Symbols

A denotes the area of the duct in square inches.
a denotes the blast area, or the ‘effective area of discharge’ in
square inches.

48 AIR COMPRESSION AND TRANSMISSION

denotes the diameter of duct in inches.

denotes the perimeter of duct in inches.

denotes a constant.

denotes the diameter of the fan wheel in inches.

denotes the density of the gas, 7.e., its weight in pounds

DerecuLaLe.

E denotes the combined efficiency of fan and its prime mover.

e denotes fan efficiency, or ratio of useful work to work of driving
fan.

g denotes the acceleration due to gravity in feet at the end of one
second, 32.16 ft.

h_ denotes the equivalent head, 7.e., the height of a column of gas
in feet having a density d, Sihaee mene will produce the velocity
pressure p ounces per square inch.

K denotes the capacity of a fan in cubic feet per minute.

1 denotes the length of duct in feet.

nm denotes the number of revolutions per minute of fan wheel.

Stef ope.

denotes the total pressure against which a fan is working.!
denotes the velocity pressure in ounces per square inch (or inch
of water) against which a fan is working.
t denotes the absolute temperature, or 460°+ F.
Q denotes the volume of gas discharged by a fan in cubic feet per
second.
V_ denotes the peripheral velocity of fan wheel in feet per second.
vy denotes the velocity of gases in feet per second due to pressure p.
W denotes the width of fan wheel in inches.
w denotes the width of blades of fan wheel at periphery in inches.
‘Pipe Losses.—Frictional losses are very hard to calculate, as
so much depends on the smoothness of the surface and the material
of which the ducts are made.
“The loss due to surface friction can be estimated from the
formule:
“For circular ducts of galvanized iron, carefully made,

“For rectangular ducts of galvanized iron, carefully made,
ly?C

100,000 A

in which p denotes the loss of pressure in ounces per square inch.
This is an empirical formula based on Weisbach’s general formula
for the flow of fluids.

1 The total pressure against which the fan is working is +s, in which ps is
the static pressure.

LECTURE BY MR. H. DE B. PARSONS 49

‘“‘Bends create an additional loss which are hard to estimate. For
all practical purposes the frictional loss due to bends can be estimated
sufficiently accurately as follows, when the ducts are of galvanized
iron, carefully made and of fairly smooth surface:

(a) For right-angle bends with the radius at the root of the bend
equal to one duct diameter, allow an equivalent length of
straight pipe equal to
11.1 times the diameter
Of the. duct.. [hus in
Pig2923, 11 B =20-1n.,
allow for the bend 11.1
X20 0r 222 in. or roft. of N
pipe: :

(b) For right-angle bends
with a radius at the root
of the bend equal to one-
half the duct diameter,
allow an equivalent
length of straight duct equal to 29.5 times the diameter of
the duct.

Fic. 23.—Right angle bend resistance.

Rotary Blowing Machines

‘“‘A rotary blower is a positive pressure blower or exhauster, and
is not a fan, although it is used for similar purposes. It is positive
in its action and it operates by displacement.

‘A rotary blower costs more than a fan of equal capacity, but it is
more economical than a fan when operating against high pressures,
that is 8 oz. per square inch or more. Turbine blowers, however,
are now being built giving efficiencies fully equal to that of rotary
blowers. (See Chapter XI).

‘A rotary blower is more economical than a compressor when
operating against pressures less than 7 lb. per square inch. Gener
ally speaking, the compressor is more economical at pressures In ex-
cess of 7 lb. bn Lie

‘‘Rotary blowers can be arranged to give constant pressures or
constant volumes. They can also handle liquids as well as gases.

‘““A rotary blower (Fig. 24) consists of a casing in which two
impellers revolve in. opposite directions. Each impeller is of a
double-lobe section symmetrical with its shaft. The impellers are
set so that the lobe of one impeller fits into the recess of the other.

The impellers do not touch each other, nor do they touch'the casing,
4

50 AIR COMPRESSION AND TRANSMISSION

although they should work as close as is possible without touching
so as to prevent loss through leakage.

‘The air is drawn in through the inlet, is caught between the lobe
of an impeller and the casing and forced around as the impeller
revolves, and discharged through an opening situated in the casing
diametrically opposite to the inlet. In order to keep the impellers
at their proper relative speeds, one shaft is driven by the other shaft
through a pair of gears.

+ aa e+ CEPUCH DG a rn Tere
I I< - Pitch Diameter. >|

¢--~-- pitch Dia, -—---—-9|

Fic. 24.—Cross-section through standard blower.

“The pitch diameter of these gears controls the size and capacity
of the machine. The radius of an impeller, or its half length, is
made three-quarters of the pitch diameter of the gears. The casing
consists of two semi-cylinders separated by a parallel section. The
radius of the cylinders is equal to that of the impellers plus clearance.
The width of the parallel section is equal to the pitch diameter of the
gears plus the clearance. The speed of revolution is regulated by
the safe speed at which the gears can be operated.

‘“‘Blower Pressures and Capacities.—The limit of the gas pressure,
in commercial sizes, is about 12 lb. per square inch. The standard
commercial sizes have capacities varying from one-quarter of a cubic
foot to 400 cu. ft. per revolution. |

“Two of the types of rotary blowers in use, are described by the
shapes of the ends of the impellers, as cycloidal or involute.
When the impeller ends are cycloidal they fit close to each other and
leave no waste spaces or pockets. Such machines are adapted to
handle wet gases and liquids as well as dry gases. When the
impellers are cycloidal, the capacity per each revolution is equal to

LBC LU REeDYe Vik He DEO APARKSON S ol

the area of the pitch circle of the gears times the length of the
cylinder.

‘When the impellers are involute, the capacity is somewhat greater
than the cycloidal and depends on the diameter of the generating
circle for the involute. This diameter is variable to suit the duty of
the blower.

“The slip is largest in small machines, and least in large ones.
Thus for machines displacing three-quarters of a cubic foot per revo-
lution at 1-lb. pressure the slip is about 60 to 70 revolutions, 7.e.,
the machine has to make that number of revolutions to hold the
pressure against leakage. For machines displacing 300 cu. ft. per
revolution at 1-lb. pressure, the slip is from 3 to 5 revolutions. The
slip for intermediate sizes is about proportional and for pressures
other than 1 lb. the slip will vary closely as the square root of the
pressures.

“For cycloidal types, the casing is 1 1/2-pitch diameters high by
2 1/2-pitch diameters wide. For involute types, the casing section
is nearly the same, but depends on the circle on which the involute
is rolled, and this depends on the duty for which the machine is
designed.

“The efficiency is variable, and for the larger sizes is between 80
and 86 per cent. falling off gradually as the pressures exceed 3 lb.
per square inch. For smaller sizes the efficiency is less.

‘‘Power for Rotary Blowers.—The horse-power required at the
shaft or pulley to drive a rotary positive blower is proportional to
the volume and pressure of the air discharged. It is safe to assume
that for each 1,000 cu. ft. of free air discharged per minute at 1-lb.
pressure, 5 h.p. is required. The following formule are sometimes
used in calculating the horse-power. The first two formule give
the theoretical horse-power required; and in order to determine the
horse-power necessary to drive the rotary positive blower it is
necessary to divide the results obtained by the efficiency of the ma-
chine. The usual efficiency is between 80 and go per cent.

er ee |

11,000

(Cr) a,

“This formula is used when it may be assumed that the air is
compressed so quickly that it does not have time to cool to atmos-
pheric temperature, as in nearly all blower work.

Oa = bs)

33,000

(2) hp. =

52 AIR COMPRESSION AND TRANSMISSION

“This formula is the ordinary “hydraulic formula” and is ordi-
narily used for pressures up to 5 oz.

lb. per Sa. ins x
GQ) hpi v

200

“This formula is frequently used by makers of positive or rotary
blowers for determining the horse-power required to operate the
machine. In this formula Q represents the volume of air in cubic
feet per minute displaced by the impellers, no allowance being made
for slippage. In the above formule P, represents the pressure of

‘5 0.06
= Horse Power Curves.

3 Machines of 1500 cv. ff. per
ra min. capacity operating the
20: samme tubes in the New York
SS Postal Service, March 1909.

c0)

Ee 0.

Le}

AZ,

ce)

qui

Horse Power Re

afr Sees in ee Saves per square [heh

Fic. 25.—Power consumed by rotary and piston compressor.

the atmosphere or the suction pressure absolute in pounds per square
foot and P the compression or discharge pressure in the same units.
(See Fig..25.)”

Mechanics of the Fan

“The laws that govern the flow of gases are the same as those for
the movement of liquids. If p, the pressure in ounces per square
inch, is divided by 16 and this result multiplied by 144, the pres-
sure will be expressed in pounds per square foot. This may also
be done by multiplying d or the weight in pounds of one cubic
foot of the gas by its height or head /# expressed in feet.

That is:

Ate = _ 9?

and as the fundamental formula for velocity is v?= gh.

v=V2 gh= 4/18 9

PEGCOURPepyeV RSA DEI BYPARSONS 53

“When # is given in inches of water:

62.4 p=hd; na SP

V=4/10.4 gt

“The theoretical velocity obtained by using this last formula is
greater than the actual velocity produced by the fan, because friction
and eddies will restrict the freedom of flow. The formula, however,
shows that the flow of gases through an orifice increases as the square
root of the pressure and inversely as the square root of the density.

“The head is made up of two parts—that necessary to overcome
the friction and eddy losses and that necessary to produce the veloc-
ity obtained.

‘““The pressure produced by afan may be considered as equal to
the weight of a column of gas one square foot in area which the fan is
supporting. This weight is equal to the height of the column times
the density of the gas. The “equivalent head’ is the height of
this column of gas. Therefore, for any given pressure, the greater
the head the less will be the density, and vice versa. Also, the
greater the head required to produce a given pressure the greater
will be the velocity.

‘““As liquids have greater densities than gases, their equivalent
heads for equal pressures will be less than the equivalent heads for
gases. As velocities vary as the square roots of the head, the veloc-
ity of gases will be greater than those of liquids under the same
conditions of pressure. That is the reason why gases issue through
orifices at greater velocity than liquids under the same pressure
conditions.

‘“‘As gases are compressible, their density will vary with the pres-
sure. Their density also varies with the temperature and with the
humidity contained. Since the velocity varies as the square root
of the head, and as the head varies inversely as the density, any in-
crease in density due to increase in pressure will reduce the head and
consequently the velocity.

“Conversely any increase in temperature reduces the density
and consequently increases the head and also the velocity. The
velocity is entirely dependent upon the head. Therefore, in mak-
ing calculations for fan operations the effect of both temperature
and density must be considered. For fan operation the standards
generally adopted are: for temperature 60° F. and for density the
weight of a cubic foot in pounds at atmospheric pressure or 14.7 lb.

Therefore

54 AIR COMPRESSION AND TRANSMISSION

per square inch absolute. When the density of the gas is given
for any pressure and temperature its density at any other pressure
or temperature can be found with sufficient accuracy for all ordinary
fan operations, by assuming that the density will vary inversely as
the absolute temperatures, and directly as the absolute pressures.

‘Thus d) = Z
by

“Tf a cubic foot of dry air weighs 0.077884 Ib. at 50° F. its weight
uy (50-++460) X0.077884

at 600° would be dy Ie es = 0.03751 lb.
‘Also, as the density varies directly as the absolute pressure
jab
Pp

“The pressure per square inch at atmospheric pressure is 14.7 lb.
absolute or 235 0z. Therefore, the density at 3 oz. gauge pressure
would be:

_ (235+3) X0.077884

dy
235

=0.0788

“The head can be expressed for 50° F. when # is given in ounces
per square inch thus:

jp TAA ANP KP ROSS ee MPs
6XdK 23ST IX(235+P) (235d-+dp)
235 |
‘“When # is expressed in inches of water as
62.4

vie 5-2 PK 400; a ae Ss eee

ene PSEA) SE

400.7

“Since 62.4 lb. the weight of a cubic foot of water at 50° F. divided
by 12 is the weight of a column of water 12 in. square and 1 in. high,
and since the pressure at one atmosphere (14.7) would sustain a
column of water 33.9 ft. or 406.7 in. high;

“Substituting these values in the formula for velocity and ex-
pressing p as the velocity pressure in ounces per square inch

04 |26X ALLS Pi. HOA 32K 21 US pee [136036.8p
235d+dp (235+p)d (235+p)d

EECTURE BY MR. H. DE B. PARSONS 55

When # is expressed in inches of water

0=4)28X 2115p = (RE CEO | ee
| 406.7d+dp N (406.7+p)d (406.7-+p)d

For dry air at 50° F. d=0.077884.

“These two formule are those used to calculate tables giving the
theoretical velocities expected at different pressures. If the gas
is not dry, but contains some vapor or moisture, its density will vary

es Fahrenheit

ie
a
Pe
ee
Fala
a a

Temperature of Air, Degre

0 10 foe 30 40 50 60 70 80
Vapor Content, Grains per Cu. Ft. (7000 Grains = | Pound)

Fic. 26.—Humidity of air.

by the quantity of mcisture which it contains, Fig. 26. (See Appen-
dix C and chart.)

“Tf the gas is at any other temperature than so” F. its density
will decrease as the temperature rises, and conversely will increase
as the temperature falls. As the temperature varies so will the veloc-
ity. As the gas becomes lighter from increases of temperature,
the velocity will increase as the square root of the ratio of the abso-
lute temperature considered to the absolute temperature of 50° F.
The converse is also true, as the gas becomes heavier from decreases
of temperature, the velocity will decrease in the same ratio.”

Effect of Outlet on Capacity

‘The shape of the opening through which the gas is discharged
from a fan affects the volume discharged in a given time. The

56 AIR COMPRESSION AND TRANSMISSION

shape of the orifice and the form of the duct affect the size of the vena
contracta and therefore the blast area of the fan and the volume of
the gas discharged.

‘““As stated, the volume of discharge is a function of the product of
the blast area times the velocity and the blast area is determined
by multiplying the area of the orifice by the coefficient of efflux.
The coefficients of efflux commonly used in practice for different
types of orifices are:

Orifice-in-a ‘thin plates... sn nraein ee a ee SSO
Short cylindrcalspipessies cane eee eee ee eee ee O75
Rounded ot-conicalmouth pieccss seat t eee 0.98
Conical pipe, angle of convergence about 6 degrees..... 0.92

‘With peripheral discharge fans, when the area of the outlet of a
fan multiplied by the proper coefficient of discharge is less than the
blast area of the fan, the pressure in the housing will equal that
corresponding to the velocity of the tips of the blades, and the vol-
ume of discharge will be less than the capacity of the fan.

‘When the area of the outlet multiplied by its coefficient of dis-
charge is greater than the blast area, the volume of discharge will
be greater than the capacity of the fan, and the velocity of the gas
as it enters the inlet must be greater than the speed of the inner edges
of the blades. Consequently, the pressure in the housing will be
less than that corresponding to the speed of the tips of the blades.”

Work Required to Move a Volume of Gas

‘A fan operating against 1 oz. of pressure per square inch and dis-
charging the gas through an orifice having too sq. in. performs work
which may be calculated in the following manner:

“The total pressure against the fan is 100 sq. in. times the coeff-
cient of efflux (say 0.75) times 1 0z., or 75 oz. or 4.7 lb.

‘“‘ Assuming that the gas is air, then from the formula for velocity
of dry air at 50° F. the air will have a theoretical discharge velocity
through the blast area of 5,162 ft. per minute.

‘The effective work is, therefore, 5,162 4.7 or 24,250 ft.-lb.
per minute or 0.735 h.p. The actual work of driving the fan is
greater than this result by the amount of power required to over-
come the mechanical resistance and losses in the fan. This resist-
ance is made up of the losses due to friction, windage and leakage.
If these losses aggregate as much as the network then the power to
drive the fan would be twice the network, 7.e., the efficiency of the
fan as a machine would be 50 per cent. The actual power required

to drive the fan would be Be 1.47 hip,

EEUGLUKE BY VMK.Hs DEB, PARSONS 57

“Placing the above in form of formule and taking f in ounces per
square inch,

Useful work =Py=—2 Hits 0S Fel See,
Since jee
16

p _ dv
De 18g

From previous formula v?= 18 g

adv
Therefore useful work =—.— ft. lbs. per sec.
2882

‘““When # is given in inches of water

Useful work =Py= 52” {t. lbs. per sec.
144
Since pete cee ued
144°* 12 144

p dv?

From orevious formula v2 =10.4 ¢*, or p=
: 4§ @ p 10.42

5.2adv° ~—_—adv®
I44X10.4g 288¢

Therefore useful work = ft. lbs. per sec.

“In these formule, av is proportional to the volume of gas dis-
charged by the fan. Since a is in square inches, the volume of cubic
feet per second will be |

av
aaerr

“Representing the efficiency by E the work to drive the fan may be
fag eee es ah
expressed as 238g Pols. Del sec,

‘From these formule it will be seen that the power varies as the
cube of the gas velocity, while the pressure varies as the square of
the velocity and the volume directly as the velocity.

‘From a consideration of these factors it is evident that fans are
more economical when used to move large volumes of gas at low
pressure than small volumes at high pressure. For this reason fans
are not economical machines for compressing gases. In addition
to the above, fans always have a clearance space between the
revolving wheel and its housing, through which space the gases have
a tendency when under pressure to leak backward, which tendency
we have seen increases as the square root of the pressure. Fans are
seldom used for pressures exceeding about ro oz., when higher pres-
sures are desired the positive blowers are more efficient and are used

58 AIR COMPRESSION AND TRANSMISSION

for pressures as high as 8 lb. When still higher pressures are desired,
compressors or blowing engines, such as described later should fa
used.

“Tn the formula above the value of v is that due to velocity head.
When the dynamic head is known, that is the velocity head plus the
friction or static head, a simple formula for brake horse-power to
drive the fan is:

‘““When ? is given in ounces per square inch

ee
brake h.p.= Bese
“When # is given in inches of water
OX5.2 p
brake h.p.= ssoKE

Design of Fans

“Tt must be evident that unless a fan is properly designed for
the work which it has to perform, there wil! be considerable loss in
power required to drive it.

“The peripheral speed of a fan must be such as to create the -
desired pressure. The pressure against which the fan has to oper-
ate is first determined, and having settled on the pressure the
peripheral speed is made to conform with it.

‘Furthermore, if the fan be direct connected either to an engine
or to a motor, dhe speed of the fan will have to conform to that of
the prime mover.

“The work formule given above are all based on the blast area
of the fan. The way in which these formule will apply to the differ-
ent types of fans will be made clearer under ‘ Description of Fans.’

‘‘For any size of centrifugal fan there exists a certain maximum
area over which a given pressure may be maintained, depending
upon and proportional to the speed at which the fan is operated.
If this area, sometimes called the ‘capacity area,’ ‘blast area’ or
‘effective area of discharge’ be increased, the pressure is lower while’
_ the volume is increased. Contrariwise, if this area be decreased,
the pressure remains constant while the volume is increased. In
practice the outlet of a fan rarely exceeds the ‘blast area.’’’

- Description of Fans

‘A disc, radial or propeller wheel fan consists of a machine having
blades so mounted on an axle or shaft that when the shaft revolves
these blades operate like a screw, and the gas is impelled ee
in the direction of the axis.

‘The blades may be straight and flat or curved. The blades may
be curved in different ways so as to increase the screw effect and

LECTURE BY MK. Ae DE B..PAKSONS 59

diminish the centrifugal effect. Disc fans with curved blades will
operate against slightly higher pressures and deliver more gas than
those with straight blades.

‘“‘As the gas enters the fan it will be forced forward with some
centrifugal effect; and this centrifugal effect can be somewhat
reduced by having the blades revolve inside of a tube so as to pre-
vent the gas from escaping over the outer edges of the blades.

‘“‘Disc fans will not operate economically against a pressure, as
the pressure will increase the slip and the leakage of the air from the
blades at the tip. If the pressure is at all high the gas will be drawn
backward near the axis and will be blown forward near the outer tips
of the blades, or in other words, the fan disc will simply circulate
the air without making any delivery.

‘Disc fans operate best when drawing gas from a practically free
suction and discharging it at no pressure. When these fans are
set up care must be taken that they do not operate against the wind,
as the wind pressure will vitiate the operation of delivery.

“The number of blades appear to have a small effect upon the
discharge, provided, of course, that the number is neither too large
nor too small. Too many blades will simply churn the air and
produce the effect of cavitation. Too few blades will not give a
sufficient grip on the air to force it forward with the proper
delivery.

‘The gas delivered by a disc fan is very irregular in velocity. If
anemometer readings are taken at different points in front of the
disc, the recorded velocities will be found to vary without apparent
reason, and the variation will not remain constant. It is, therefore,
very hard to determine the mean velocity of discharge of the gas.

“The number of revolutions is limited by the strength of the fan
and by the fact that a high velocity will cause the fan to hum and
be noisy. The revolutions are usually limited so that the velocity
at the tips of the blades shall not exceed 8,500 ft. per minute. (For
noiseless operation, 4,000. Usual maximum 7,000). On this as-
sumption, if D denotes the diameter of a fan in inches and

nm denotes the number of revolutions per minute.

zDn
es see sures and Dn=32,000 (nearly).
“This last equation may be used to determine the limiting revo-
lutions or diameter by assuming one or the other.
“The volume of gases discharged by a disc fan with a free suction
and discharge can be estimated from the formula:

D? :
=1/4 2—v in which v=0.20V
Q=1/4 a 39

60 AIR COMPRESSION AND TRANSMISSION
“The brake horse-power for the fan with the above value of Q
can be estimated from the formula:

_OXxdXv3/? ( 13.5 is a constant
RLS OED 550X 28 ‘S \found from experience

‘When a disc fan operates against a pressure, 7.e., not with a free
suction and discharge the above formule must be changed as Q
becomes less because the slip becomes greater.

‘‘ Approximately 0 = 1/47 (=) vy in which v! equals 1.25v less 45
per cent. of the theoretical velocity due to the pressure against
which the fan is working.

“The brake horse-power will be the same as if the fan were work-
ing without restriction, although the volume of discharge Q will be
less.

“Example: Free suction and discharge. Fan wheel 48 in. in
diameter running at 450 revolutions per minute. Find Q and
brake horse-power. Dry air at 50° F.

a Se ee ee
V=0.39X3.14X— Xoo 36.8

8 2,
Q=1/4X3.14Xx (E) X36.8=462

462 X0.077884 X 36.83/2 X 13-5
550 X 64.32

Brake horse-power = 3:07

‘‘Also restricted discharge. Find Q and brake horse-power for
the same fan and conditions when operating at 5/8 in. of water
pressure.

“The velocity due to 5/8-in. water pressure is 51.8 ft. per second.

vi=1,.25 X36.8—0.45 X51.8= 22.8
Q\ 2
Q=1/4X3.14X (~) X22 9 = 250

“The brake horse-power would be 3.07 because it is approximately
the same as if the fan were working unrestricted. It is found by
substituting the unrestricted value of Q instead of the actual re-
stricted value.

“Centrifugal Fans.—Centrifugal fans operate on the principle of
the vortex. They suck the gas in and discharge it off the periphery
of the wheel by centrifugal action.

‘Fan Blast or Steel Plate Machine.—The fan wheel consists of
an axle or shaft on which are mounted radial arms carrying floats or
blades. Each blade is narrower across the tip than it is across the

LECTURE BY MR. H. DE B. PARSONS 61

body. The blades are mounted inside of side plates, so that the
gas is confined in the spaces between the blades, which thus form
passages from the suction to the discharge side of the fan. Thése
side plates also prevent the loss of friction between the revolving
air and the sides of the housing.

‘“‘Sometimes the blades are curved backward at the tips so as to
make the fan run more quietly, and sometimes the blades are curved
backward for their whole depth so
that the gas may enter the wheel
and pass through it without shock.

‘““When fan wheels have flat
blades, they can be run equally
well in either direction, but when
the blades are curved, the wheels
should revolve with the convex
sides of the blades in advance.

‘““When these fans are used for
blowers, there is usually an inlet
on both sides of the housing; and
when used as exhausters, usually
an inlet on one side only, as it
facilitates the connection with the
suction duct. w+ 04D

‘

“The diameter of inlet is gener- wid ae 9

_ 1
ally 0.6 or 0.7 of the diameter of 4-4 /east ia, but beection
Ahad heal actually from ' of Fan
e Ian wheel. 2a to $a. : Wheel

“For high efficiency the area of Lv

inlet should not exceed 4o per cent.
Ol athe dise, areayo! they wheel.
The full width of the blades is
generally made either one-half or three-eights of the diameter of
the “wheel. The blades are generally cut off at the upper outer
corners so as to taper at the tips at an angle of about 20 degrees
with the side edges, but their width at the periphery should be not
less than 0.6 to 0.7 of the width at the root, 7.e., not less than their
maximum width times the same ratio as was chosen with the ratio of
inlet to wheel diameter. Usually w=o0.4 D and w=o.5 D (see
Fig. 27). |

‘The width of the fan is made such as to provide the proper area
for the flow of the gases through it so as to discharge the required
volume. If the diameter of the fan wheel is made too small, it may
not be possible to give the wheel sufficient width to permit the
necessary discharge of volume, unless the fan is run at a very high
rate of speed. This increased speed will result in raising the pres-
sure above that required, and will, therefore, increase the power

Fic. 27.—Steel plate fans.

62 AIR COMPRESSION AND TRANSMISSION

necessary to drive the fan. Contrariwise, if the wheel be given
a large diameter it may have to be made so narrow, in order to dis-
charge the required volume, as to become impracticable. It will,
therefore, be seen that under any given conditions there will prob-
ably be one diameter and width which will be best suited for the
work.

“The blades of this type of fan are given sufficient depth so as to
project inside of the circle of the inlet in order that they may better
grip the incoming gas and force it through the wheel.

‘With a peripheral discharge fan enclosed in a housing, the limit
of its capacity to maintain a given pressure is measured by its blast
area. In other words the velocity of discharge will be approxi-
mately equal to the peripheral speed of the fan, and the volume will
be measured by this velocity times the blast area.

‘“‘Tf the blast area be increased the pressure will be less, and if the
blast area be decreased the pressure will remain the same. For a
peripheral discharge fan with a housing the blast area can be cal-
culated as follows:

Let D denote diameter of fan wheel in inches;

w denote width of fan wheel at periphery in inches;
c denote a constant, depending upon the design of the fan
and its housing, but which has a value not far from 2 1/2 to

?
a denote the blast area in square inches.

D D
Blastrarea—o — es = = g(nearly)

“Tf the shape of the discharge orifice and duct be known, and the
coefficient of contraction determined, the area of the discharge orifice
would be the blast area, as determined from the above formula
multiplied by the reciprocal of the coefficient of contraction.

“The usual maximum peripheral velocity for standard fans is
6,600 (for noiseless operation about 4,200) ft. per minute, but should
not exceed 8,000 ft. per minute. This latter figure limits the pres-
sure to 1 3/4 oz. per square inch but special fans may be designed to
maintain a pressure as high as about 12 oz.

“The volume past the blast area is about 86 per cent. of the
peripheral speed. In other words, the peripheral speed must be
1.16 times the velocity due to the pressure of V =1.16 2.

V
wheel circumference

Therefore n

“The efficiencies without prime movers vary from 45 to 50 per
cent. for commercial sizes when using dynamic head, or from 30 to 35

EPEGCLOREOBY MRA Ho DE B- PARSONS 63

per cent. when using velocity lead. The outlet is generally made
square, and its area is usually about two and one-half times the
blast area, or A =2 1/2 a, but never less than 1 1/2 a. -

“This proportioning will make the bottom of the outlet below the -
periphery of the wheel. The efficiency of commercial sizes is about
45 to 50 per cent. without a prime mover. If the prime mover
efficiency is taken at 85 to 90 per cent. then the total efficiency of
the fan and prime mover would be between 38 and 45 per cent.
Example:

‘“‘Given the quantity of air per minute, 65,000 cu. ft., the temper-
ature of dry air 70° F. and the pressure 1 3/4 in. of water. Deter-
mine the diameter of fan, revolutions and brake horse-power.

Under these conditions the density of the gas or its weight per
cubic foot may be taken as 0.0754 lbs.

a] 136036.8 X1.75 e
v= = 88.0
(406.7-+1.75) X0.0754
65000 «X88.0 _ 65000X144 _
eee: Therefore a= 60% 88.0 =1,770

Making w=0.4D

dak
1 7O= 91 D*=11,500,, D=107

ez", 10 6 00.0 == 102

Wheel circumference x14 = 27.0 it.
ESE
27-9
With E as 45 per cent. the power to drive the fan is

Therefore nN 220

1770 X0.0754 X 88 By
550X288 X 32.16 X0.45

“‘Ffousing.—The housing is placed around the wheel in an eccen-
tric position and has a form approaching the spiral. This arrange-
ment facilitates the gas delivery from the wheel. The openings
for discharge of the gas are tangential to the wheel. There may be
one or more openings as circumstances demand but their combined
area of discharge should not exceed the fan capacity. It makes no
difference whether these discharge outlets are placed horizontally
or vertically.

“The arrangement of discharge outlet, however, gives a name to
the fan—as a horizontal top discharge, a vertical discharge, a hori-

Brake horse-power

64 AIR COMPRESSION AND TRANSMISSION

SS
SS
S wf

———
SS Ss
SSS Ss
="
=

SSeS
See
Se
SSeS sss

Sse
SS

== 35553 [22SS
= =

SSS
ee

Fic. 28.—Full housed steel plate Fic. 29.—Full housed steel plate fan.
fan. Left-hand bottom horizontal Right-hand top horizontal discharge.
discharge.

Fic. 30.—Three-quarter housed steel Fic. 31.—Three-quarter housed
plate fan. Right-hand bottom horizon- steel plate fan. Left-hand top
tal discharge. horizontal discharge.

LECTURE BY MR. H. DE B. PARSONS 65

zontal bottom discharge, a double discharge, etc. (Figs. 28, 29, 30,
2randes 2.)

“The spiral or scroll form of the casing should be such as to let
the gas escape with freedom from all parts of the periphery. The

smaller diameter of the scroll should not be less than Di and the

re ence

D denotes the diameter of the wheel in inches;
a denotes the blast area in square inches;
W denotes the maximum width of blades in inches.

larger diameter not less than D+

de
fae

O oy”
A!

qu
Fic. 32.—Allis Chalmers steel ventilating fan.

‘Cone Wheel Fans.—The cone wheel fan is a single inlet pe-
ripheral discharge fan. It is used both with and without a housing.
Cone wheel fans are not efficient for use against pressures in excess
of 1 oz. per square inch and are seldom used against pressures as high
as this limit. Generally speaking, they are not as economical in
the handling of gases as centrifugal fan-blast machinery properly
encased in a well-designed close-fitting housing.

‘““Cone wheels should have a perfectly free inlet and be arranged
to have a free discharge of air from all points of the periphery.
When cone wheels are encased in a housing the housing is usually
much larger than the fan wheel to permit a perfectly free and un-
restricted discharge. As ordinarily arranged, the inlet to a cone
wheel is a hole in a wall of the apartment from which the gas is to be
sucked-(Figs. 33 and 34).

“On the axle or shaft of the fan there is mounted a cone with its
apex turned toward the inlet. Between the cone and the periphery
of the wheel there are blades or floats, and these blades are encased

66 AIR COMPRESSION AND TRANSMISSION

inside of side plates. As the air enters the inlet it is deflected by
the cone to the floats, which together with the side plates, continue
to change the direction of the air so that it is discharged off the pe-
riphery in a plane at right angles to the shaft or line of entrance.
The width of cone wheels is generally one-quarter the diameter of
the wheel, and the inlet opening is generally three-quarters of the
diameter of the wheel. The floats are curved backward and tapered

Yt]

Fic. 33.—Cone fan inlet side. Fic. 34.—Cone fan discharge side.

toward the periphery so that they have a width at the tips of about
three-quarters the width of the wheel.

“Assuming that the air is discharged at a velocity equal to the
speed of the tips of the floats, the capacity of a properly designed
cone wheel in cubic feet per second is approximately:

QO=6.4D%*/p
for exhausting and O=s.0D?\/p for blowing in which

D denotes the diameter of the wheel in inches and
p denotes the pressure in ounces per square inch corre-

sponding to the velocity of the tips of the blades.
‘The horse-power required to operate a cone wheel including

efficiency is approximately brake horse-power = for exhausting
43

and

brake horse-power=? for blowing.

PECTOREPUBY MK ae DE BeePARSONS 67

“The limits of peripheral speed are about the same as for a disc
fan so that assuming the speed of the tips of the blades, the revo-
lutions and diameter can be calculated by assuming one or the other
as with disc fans.

“Turbine Blast or ‘Sirocco’ Fan.—The name Sirocco is a trade
name. ‘These fans are centriufgal in their action and have a pe-
ripheral discharge. ‘The runner or blast wheel is built up of steel,
and consists essentially of three parts—the interior cone to deflect
and turn the air as it enters the inlet toward the blades, the blades,
and the side plates. The runner is shaped like a drum. The
blades are Jong and narrow radially, being generally between six
and nine times as long as they are wide. The runners are usually
equipped with about 64 blades. The blades are curved so that the
concave side revolves forward. The blades are not quite as deep
as the side plates, and the side plates are made one-sixteenth the
diameter of the wheel.

‘““These fans are usualy made with the inlet on one side only, the
other side being closed by the back of the internal cone. When
double inlets are desired, two internal cones are placed back to
back, and this practically means that the wheel is made double.

‘‘The peripheral speed at which these fans can be run before they
begin to hum or become noisy is higher than that of the steel plate
fan. With ordinary sizes they remain quiet until peripheral speeds
of 10,000 ft. per minute are reached, but a peripheral speed of over
15,000 ft. per minute is often used.

“The design of these fans is quite different from that of a steel
plate fan. Having the blades curved forward, which is quite correct
in principle, results in an increase in the velocity of discharge, which
with properly shaped blades, is twice that of the “‘steel plate’’ fan.
In other words, the velocity of discharge from a properly designed
turbine fan, is theoretically twice that of the periperal speed, and
actually about 117 per cent. of the peripheral speed. As these fans
are made about twice as wide as the steel plate fan, and as the
velocity of discharge is twice as great, a turbine fan wheel will dis-
charge about four times the volume of gas as a steel plate fan of
equal diameter and run at equal speed.

‘These fans are mounted in a housing, and the ratio of outlet to
inlet is one totwo. There are three types of these fan wheels which
are distinguished by the manner in which they are bladed (Fig. 35).

(a) In the forward inclined wheel the blades are curved
forward of the radii.

(b) In the radially inclined wheel, the blades are curved on
the radii.

(c)’ In the backward inclined wheel, the blades are curved
behind the radii.

68 AIR COMPRESSION AND TRANSMISSION

“These different forms rank in the order of mechanical efficiency
as given above, and in the matter of speed of revolutions in the
reverse order. Owing to the high speed at which these turbine fans
can be run they are well adapted for direct connection to high-speed
prime movers (Fig. 36).

“The commercial efficiencies when used for blowing are about 59
per cent. for forward blades, 53 per cent. for radial blades and 50
per cent. for backward blades.

SSS
WNZE NZ

C
Fic. 35.—Three types of blading. Fic. 36.—Sirocco double inlet runner.

‘The width of the runner is made two-thirds the diameter for
standard sizes.
‘The volume of discharge can be calculated from the formula

av
=———. in which y=V X11;
Car v, 44
2) Dae met
a SS Se eee en
277 aes] SO 225
‘‘When the dynamic head is known the same formula for power
required can be applied to this fan as to others.”

CHAPTER VII

PISTON COMPRESSORS

In the practical applications of air for power and other purposes,
the simplest method of compressing air above those pressures
for which the fan and positive pressure blower are particularly
adapted is by means of a piston compressor.

This method of compression is used more than any other method
and for that reason considerable attention will be given to it. Fig.
37 shows in cross-section such a compressor, its piston, piston-rod

\ IZA i
A
ER SC
Kea LS

Of LLIN

Soom

LYtEY

le
aes)

RNewa
[DOISSSSSSS

| Ze

iY,
Sinn

Fic. 37.—Cross section of piston compressor.

and valves. D represents the inlet valve through which the “free
air’? is drawn into the compressor, and E the outlet or discharge
valves through which the compressed air is discharged into a reservoir
or receiver as it is quite commonly called.

Naturally with this type of compressor the piston movement
is limited, so that at its extreme positions the piston will not strike
the cylinder ends or heads. That volume of the cylinder through
which the piston does not move, together with the volume occupied

70 AIR COMPRESSION AND TRANSMISSION

by the passages leading from the cylinder to the valves, is called the
clearance.

Action of Piston Compressor.—Suppose the piston is to start
at the right end of its stroke. As it moves to the left, a slight
vacuum will be created in the cylinder on the right side of the piston
and valve D will be opened. Free air will rush in, following the
movement of tke piston and filling the cylinder. As the piston
starts to the right, this same operation will take place on the left
side of the piston, while on the right compression will commence as
the volume occupied by the air in that part of the cylinder is reduced.

If the valve E is in communication with a reservoir of com-
pressed air at 30-lb. gage pressure, no air can escape from the cylinder
until the pressure has risen to a little above 30 Ib. When this is
done the valve E will be lifted from its seat and the compressed air
will be pushed bodily out of the cylinder into the reservoir.

At the end of the stroke, the clearance volume will be filled with
this compressed air and as the piston starts back valve EF will close
and the compressed air in the clearance space will expand to fill
the gradually increasing volume in the cylinder and will continue
to do so until the pressure in the cylinder is lower than that of the
atmosphere, when valve D will open and a new supply of free air
be drawn in. The clearance space reduces the volume of free air
‘ drawn into the compressor.

Indicator Card of Piston Compressor.—These changes are shown -
by the indicator card given in Fig. 38, which shows the changes in

|
|
|
|
|
|
!
!
I
j
r
|
|
|

[aie

cap)

V 0.

Fic. 38.—Indicator card of piston compressor.

pressure and volume taking place on one side of the piston. The
distance LZ represents the volume displaced by the piston during
one stroke. Starting with the piston at G and the cylinder full of
air, the pressure of the confined air will gradually rise as the volume
is being reduced by the moving piston until a pressure H, a little
above the pressure in the reservoir is reached. From here until

PISTON COMPRESSORS 71

the end of the stroke the piston will force the compressed air out
of the cylinder into the reservoir. The wavy line H—J represents
this expulsion, the inertia of the moving parts of the indicator and
the fluttering of the discharge valve on its seat causing the irregu-
larities of this discharge line. When the stroke is completed the
pressure of the air in the reservoir closes the valve F, leaving the
clearance space full of compressed air at the high pressure J. As
the piston moves to the left the compressed air in this clearance
space will expand to fill the constantly increasing volume until a
pressure a little below that of the atmosphere is reached (K), when
valve D will open and free air rush in to fill the cylinder as the piston
continues its movement.

The more air in the clearance space the further wil] the piston B
have to move before the compressed air in the clearance space can
have the opportunity to expand low enough to open the inlet valve.
For this reason the clearance space for a piston alr compressor
is made as small as possible.

Effect of Clearance.—The loss due to clearance is not a loss of
power, for most of the energy used in compressing the air into the
clearance space is given back in expanding and helping to move ‘
the piston. The only loss in power is the heat loss through radia-
tion. The loss due to clearance is mainly a loss of capacity which,
in many cases, is a rather serious matter. Engineers have sought
to reduce this to a minimum with the result
that the clearance volume of a modern piston
. air compressor varies from 0.02 to 0.0094 of
the volume of the piston displacement.

Methods of Reducing Clearance.—Various
methods have been devised to reduce this
‘ clearance, some even going to the extent of SEES WY
putting in spring heads on the cylinder which MMM LL
the piston could strike at each end of the F'6. 39.—Uncovering
stroke without serious injury. This method, ? PEA EARL gee

: . Bing pressure.
however, introduces complications that are
not always desirable.

Another method of reducing the clearance loss that has been
suggested is to let the piston uncover a passage Jeading from the
clearance space just at the end of the stroke, and thus allow the
compressed air in the clearance space to expand into the cylinder
on the other sid: of the piston, as shown by Fig. 39, without reducing
the capacity of the compressor. This, however, is open to the

72 AIR COMPRESSION AND TRANSMISSION

objection that the compressed air in the clearance space, which
normally acts as a spring, is released and the piston will pound
at the end of each stroke unless some other means is used to prevent
it.

Some piston air compressors are so designed that when the
machine is cold the piston will almost touch the cylinder-head
when at the crank end of its stroke. As the compressor is operated
the heat of the air being compressed will be partly transmitted to
the piston-rod and cause it to expand slightly, thus increasing

Cc B B
LZ, ae. IZ, eae p, K

== |")
Z %

y y

4__y

ZZZZLLLETLE y

sxx

Fic. 40.—Hydraulic piston compressor.

the crank-end clearance and decreasing the head-end clearance,
and by designing the compressor so that when hot, the piston will just
miss touching the cylinder-head when at head-end dead-center,
and when cold just miss the other cylinder-head when at dead-
center, the clearance volume of the cylinder is reduced to a minimum.

Some of the early air compressors were built, as indicated by
Fig. 40, so as to reduce the clearance volume by using a water column
as a piston. As the piston A is moved back and forth the water
column in each upright cylinder is caused to alternately rise and
fall. As it falls, valve C is opened and free air rushed in. As the
water column is raised, valve C is closed and the confined air is
compressed to a pressure sufficient to open valve B and permit
the compressed air to escape to a reservoir. The water at its upper
height can fill the entire space, including the passage to the valves,
which a metal piston could not do, so the clearance is in this way
reduced to a minimum. This type of compressor was practically

PISTON COMPRESSORS 73

abandoned, but its principle has been recently revived in air com-
pressors working on the principle of the Humphrey pump.

Suction Line.—If in the design of a piston compressor the inlet
valve or the passages for the same should be too small, then the air
cannot rush in as fast as the piston
moves and the suction line, instead
of being straight, will fall below the

horizontal, to rise again near the StS.
end of the stroke as the piston
- ee
velocity decreases. Ne

A suction line that is not hori-
zontal indicates restricted inlet for
admission of air.

It may sometimes happen that as the piston, which has its max-
imum velocity near the middle of the stroke, nears the end with a
decreasing speed, the inlet valve will close before the end of the
stroke is reached, and the admission line will fall slightly. As
compression starts, this line will be retraced until a pressure greater
than the admission pressure is reached, as shown by Fig. 41. A
good indicator card for a piston air compressor will have the com-
pression line start very close to the end, as shown by Fig. 38.

Compression Line.—One reason for the ideal method of air
compression being isothermal, as already explained, is because

Fic. 41.—Effect of early closing of
inlet valve.

0 V

Fic. 42.—Card showing isothermal and adiabatic compression.

any energy stored as heat in the compressed air above the temperature
of the surrounding atmosphere will soon be radiated and hence lost.
Another reason, as shown by the chart, Fig. 11, is that if the com-
pression is not isothermal, the pressure due to the increased tempera-

74 AIR COMPRESSION AND TRANSMISSION

ture will rise above that which would result from isothermal com-
pression and hence cause an increased expenditure of energy to
operate the compressor, as.shown in Fig. 42. This increased
expenditure of power may be avoided by isothermal compression

The approximate horse-power required to compress air under
isothermal or adiabatic conditions may be determined as indicated
by the formule in Chapter IV. This has been done for various
pressures indicated in Table IX, which will give an idea of the
saving to be secured from isothermal compression.

TABLE IX
Single-stage compression, from atmospheric pressure at sea-
level. Inital temperature, 60° F. Horse-power required
to compress 1 cu. ft. of free air
Atmos-
Gage | pheres Calculated horse-power Actual horse-power (approx.)
pres- | absolute .
qhien aCe Sac Allowance for | Allowance for
pounds} of com-
: : : losses above losses above
pression | Isothermal | Adiabatic ; : j ;
; ‘ adiabatic com- | adiabatic com-
compression | compression 3 :
pression, 15 per | pression, 20
cent. percent:
20 2.30 0.0551 0.0626 0.0720 0.0751
25 AR ie 0.0637 0.0741 0.0852 0.0890
30 3.04 ORO7 Ts 0.0843 0.0970 ©. LODI
35 3.38 0.0782 0.0941 0.1082 ©.II29
40 aa72 0.0842 0.1029 oO, T1S2 0.1234
45 4.06 0.0895, Ovrrir 0.1282 0.1338
50 4.40 0.0950 OnlLor 0.1470 0.1430
55 Ant A 0.0994 0.1269 0.1460 Oris22
60 5.08 O.IO4I Or1337 0.1537 oO. 1604
65 5.42 0.1081 ©. 1401 0.1610 0.1681
70 5.70 Omer res 0.1468 ©. 1690 0.1761
75 6.10 O.1162 O7E535 0.1765 0.1842
80 6.44 O.IIQ5 On1s0n 0.1830 ©.I9I0
85 6.78 Ov1224 OBTOSE 0.1900 oO. 1961
go ipl Ont250 CO: 1702 0.1955 ©. 2040
95 7.46 0.1287 0.1760 0.2024 O;2112
100 7.80 eyes 8 0.1807 0.2080 0°. 2168
IIo 8.48 0.1366 0.1894 0.2180 O, 2272
125 9.50 ©.1442 0.2025 0.2328 ©. 2430

Wet and Dry Compression.—The two principal systems which
have been used in attempting to secure isothermal compression in a

PISTON COMPRESSORS | 75

piston air compressor are the ‘‘wet system” and the “dry system.”
A. wet compressor is one which introduces water directly into the
cylinder during compression.

A dry compressor is one which admits no water to the air during
compression but surrounds the cylinder with a jacket of circulating
water in order to reduce the heat of compression.

There are two kinds of wet compressors: First, those which in-
ject water in the form of a spray into the cylinder during compres-
sion; second, those which use a water piston in compressing the
air, such as shown in Fig. 4o.

Numerous tests have been made of these different methcds of air
compression showing that the compression cen be brought closest
to the ideal isothermal by means 0i injecting a spray of water directly
into the cylinder during the compression.

Although the best results have been secured by the wet system of
compression, still it has been practically abandoned in favor of a
dry system of compression using a water-jacket, for the following
reasons:

First. The mechanical difficulty of introducing the water in a
fine enough spray to reduce the temperature of compression as It is
being produced.

Second. Impurities in the water through mechanical and chem-
ical action destroy the metallic surfaces of the cylinder and piston.

Third. Wear due to insufficient lubiication.

Fourth. Difficulty of regulating the amount of water to be
introduced.

Fifth. Limitations of speed due to the presence of water.

Actual Compression.—If an isothermal and an adiabatic line be
drawn on an indicator card taken from an actual modern air com-
pressor starting at the point where compression begins, the actual
compression line wiJl come very close to the adiabatic.

If the compressor operates at a very slow speed, there is an
opportunity for the heat that is generated by compression to be
radiated and the compression line will come closer to the ideal
isothermal.

That is, with a high-speed air compressor, the compression is
approximately adiabatic, while with a slow-speed compressor with
efficient water-jacket the compression line may approach the
isothermal, speed being an important element in determining the
slope of the compression line.

Cards from Air Compressors.—In taking indicator cards from an

76 AIR COMPRESSION AND TRANSMISSION

air compressor all the precautions that are necessary for taking
steam cards apply with equal force, as erroneous conclusions may
be very easily drawn from the cards if care is not taken. For
instance, if the piston of the air compressor leaks slightly, then the
compression line will approach the ideal isothermal line, indicating
' a very desirable compression line, while in reality the compressor is
defective. For this reason, extra precautions must be taken in
order that the card may indicate the true state of affairs in the
cylinder.

When air is compressed in a cylinder to a pressure of 100 lb. per
square inch without cooling, temperatures ranging from 475° to
550° are reached, and these high temperatures are not only produc-
tive of poor economy and low efficiency but are dangerous because
of the explosive nature of the compressed air containing the vapors
of the cylinder oil used for lubricating the piston.

CHAPTER VIII

EFFICIENCIES AND ENERGY COMPENSATION

In the discussion of Chapter IV, clearance was disregarded, but in
calculating the required dimensions of an air compressor the effect
of clearance must be considered.

It is usual to express clearance as a certain percentage of the piston
displacement. If this percentage expressed as a decimal fraction
is represented by C, the volume occupied by the air to be com-
pressed at the end of a suction stroke will be (1+C) times the
piston displacement. —

Volumetric Efficiency.—The effect of clearance upon the capacity
of a compressor js usually expressed in terms of the “volumetric

C B

'
I
'
{
|!
1
i
!
!
1
i
1
|
1
'
G

Fic. 43.—Ideal card for piston compressor.

efficiency,” but as this term is not always interpreted in the same
way it is advisable to use two terms “apparent volumetric efficiency”
and “‘real volumetric efficiency.”’

Apparent Volumetric Efficiency.—The apparent volumetric eff-
ciency is the apparent volume of free air drawn in as shown by
’ the indicator card divided by the volume of the piston displacement,

Oralieice. a AG of ideal card, Fig. 43. In an actual card, Fig. 44, this

AD
ratio is also shown by 1G"
Hi

78 AIR COMPRESSION AND TRANSMISSION

In Fig.-43, the clearance line C—D will follow the equation
beVe" = paVa”
but, as V-=C, this may be written

pPeC” =paVa"™

be\n
va=(F) C

from which

Fic. 44.—Actual card of piston compressor.

The apparent volumetric efficiency, or en may be written

AG—DG

AGES

or id
DG
DEINE

pee

Calling the piston displacement AG unity, the apparent volumet-
ric efficiency may be written

Be ; Pe \n
_-ofts)ixccex-ef (22)

The effect upon the capacity may be illustrated by assuming a
compressor in which #¢ is 80 lb. per square inch gage, or 94.7 lb. ab-
solute, and C is 2 per cent and 7 is 1.4.

EFFICIENCIES AND ENERGY COMPENSATION ico

Substituting in the above,

| for Three Stage

pe\eu
Po

| for Two Stage (

14
“| for Single Stage

0.5

Fic. 45.—Loss of capacity due to clearance.

That is, such a compressor would only take in 94 per cent. of the
piston displacement in free air, and poor valve action would reduce
this capacity still further. The loss of capacity due to clearance for
various pressure ratios is shown in Fig. 45.

80 AIR COMPRESSION AND TRANSMISSION

True Volumetric Efficiency.—The above illustration would be
true if the temperature of the air after being drawn into the cylinder
were the same as the atmosphere, and the pressure at the instant of
compression equal to the atmosphere. As this is seldom true, it is
necessary to make correction for this by multiplying the above

expression by
ae A
iP Pam
in which the subscript am stands for atmospheric conditions and
subscript 1 for conditions at the beginning of compression.
The true volumetric efficiency is the ratio of the free air actu-
ally drawn in to the piston displacement and is represented by the

formula
Alas Pi Pe ae
Tie lad(enie |!

Cylinder Efficiency.—The cylinder efficiency of an air compressor
may be defined as the ratio of the work done in a complete cycle to

Fic. 46.—Cylinder efficiency.

compress isothermally a volume of air at atmospheric pressure equal
to the intake piston displacement divided by the actual work done
in the air cylinder.

This would be Fig. 46, the area AKCG divided by the shaded area,
or the actual work done in the air cylinder.

Efficiency of Compression.—The efficiency of compression may be
defined as the product of the cylinder efficiency and the true volumet-
ric efficiency, or it is the work done in a complete cycle to compress
isothermally, without clearance, a given volume of free air divided
by the work actually expended in compressing the same volume of
free air.

EFFICIENCIES AND ENERGY COMPENSATION 81

Mechanical Efficiency.—The mechanical efficiency of an air com-
pressor is the work done in the air cylinders divided by the work done
in the steam cylinders, if driven direct by steam, or in the gas-engine
cylinders, if gas engines are used, or the work delivered at the belt
if the compressor is belt driven.

Net Efficiency.—The net efficiency of a compressor unit driven by
a steam engine or turbine direct is the ratio of the internal energy
available in the compressed air at room temperature to the heat
energy available in the steam supplied; or it is the energy available by
adiabatic expansion of the compressed air at room temperature to
atmospheric pressure divided by the energy available in the steam
supplied, if expanded adiabatically in a Rankine cycle.

In considering efficiencies of air compressors, it is important to
distinguish between a machine used for compressing air as a means
of storing and transmitting mechanical energy, in which the ideal
compression is isothermal, and a machine used for supplying air
under pressure for purposes of combustion, as in forges, cupolas and
blast furnaces. In these last cases the pressures are comparatively
low and the resulting increase of temperature due to adiabatic com-
pression is not objectionable. In fact there is ample justification
for taking, in these cases, adiabatic compression as the standard.

Blower Efficiency.—Henry F. Schmidt in an article in the Journal
A. S. M. E. of Nov., 1912, on “‘Centrifugal Blowers” indicates a
“blower efficiency’? for any blower not water-jacketed, by dividing
the rise of temperature, as calculated from adiabatic compression
from the suction to the discharge pressure, by the actual rise of
temperature taking place during the compression in the blower.

The losses in a blower are principally friction, eddies and leakage.
All energy losses reappear as heat and bring the temperature after
compression higher than that due to adiabatic compression, and in

the article the author proves that this ratio will reduce to the form
11—T2. ; ae ;
ToT, in which Tis the initial temperature of the air, 71’ its actual
final temperature, and 7, the final temperature if the compression
had been adiabatic. This formula is open to the criticism that the
radiation is disregarded, but as its value is comparatively small the
‘blower efficiency” expression has the decided advantage of sim-
plicity and ease of determination.

Economic Efficiency.—Franz zur Nedden in his articles on Turbo-
blowers and Compressors in the Engineering Magazine for Nov.,

1912, states that the thermic losses of a compressed gas may be
6

82 AIR COMPRESSION AND TRANSMISSION

expressed by the contraction which it undergoes in cooling. In
place of the larger volume of power medium which leaves the com-
pressor, a diminished volume only at the same pressure reaches the
destination. As in perfect gases contraction due to cooling is in
direct proportion to the absolute temperature, the fraction formed
by taking the absolute temperature of the atmosphere as the numera-
tor and the absolute temperature of the air or gas leaving the
compressor as the denominator, might be taken as a fair expression
of the losses caused by the unutilized heating of the gas or air in the
compressor.

As this loss would not occur if the gases were compressed isother-
mally it is debited entirely to the compressor. He cites in illustra-
tion a compressor of the piston type of 140,000 cu. ft. per hour
capacity working against 115 lb. per square inch at go r.p.m., in
which the temperature leaving the compressor was 197° F. and the
temperature of the atmosphere 41° F., this gives an ‘‘economic
f 460+41 __ 501

400197 657

Energy Compensation.—If an air compressor is driven direct by a
steam engine with the steam and air cylinders tandem and one

efficiency”’ o = 76.1 per cent.

V (LLL LLL]

VE LLIILZLIL LLL)

y | y Z == Z

BL Ae y

ey (=
Steam ‘Air

Fic. 47.—Direct acting steam compressor.

common piston-rod as shown in Fig. 47, with the valves arranged to
give a steam and air card as shown, the greatest force is exerted on
the piston-rod at the time when the least is required in the air cylin-
der and when the air cylinder needs the greatest force applied to
expel the compressed air, the least is being applied in the steam
cylinder.

Many ingenious contrivances have been devised for storing the
excess energy developed in the steam cylinder during the beginning
of the stroke and drawing on this excess during the last part of the
stroke.

EFFICIENCIES AND ENERGY COMPENSATION _ .88

When a fly-wheel is used it must of necessity be very large in
order to do this, as the amount of energy that can be stored in the fly-
wheel will depend upon its weight and speed.

Hydraulic Compensator.—One form of energy compensator is
shown in Fig. 48, which represents a sketch of a D’Auria non-
rotative air compressor.

The desired result is obtained by using a ‘“‘ hydraulic compensator,”
which consists of a cylinder A fitted with a plunger B carried by
the same piston-rod that connects the steam and air piston. The
ends of the compensator cylinder communicate with each other by

Seomcesas A Comm nant

Y au, é
H gf D

SESE

Fic. 48.—D’Auria System of energy compensation.

means of a loop of pipe c-c—c so constructed as to form a very rigid
bed-plate for the machine, a very desirable feature, as it helps to
keep the machine in alignment. The cylinder and pipe are filled
with water, or any other liquid, leakage being made up through a
pipe.

When the compensator is in action, the liquid column contained
in the compensator is moved reciprocally and as it requires energy
to start a mass moving and also to stop it after it gets in motion, the
excess energy of the steam cylinder is used up or rather stored in
starting the liquid in motion during the first part of the stroke, and
this excess energy is given back during the last part of the stroke as
the pistons near the end of their stroke.

Lever Compensation.—Sometimes two steam air compressors are
placed side by side and the piston-rods connected by a system of
levers as shown in Fig. 49, so that the excess energy that is not needed
in one air cylinder is conveyed by the system of levers to the other
air compressor and aids that near the end of its stroke. By this ~
arrangement one compressor supplements the other.

Weight Compensation.—A method adopted by the Norwalk Iron
Works is best shown in their two-stage compressor, driven by a
tandem compound steam engine, as shown in Fig. 50.

84 AIR COMPRESSION AND TRANSMISSION

By arranging air and steam cylinders, as shown, with a common
piston-rod, an excessively heavy moving piece is secured, which
requires considerable energy to start in motion and also to bring to
rest near the end of the stroke. That is, a large share of the energy

Gx ZZ

i
ae
SAAS

Fic. 49.—Lever system of energy compensation.

developed in the steam cylinder during the beginning of the
stroke is used in starting this heavy piece in motion and the extra
energy required in the air cylinders during the last part of the
stroke is taken from this moving mass in bringing it to rest.

gure “ur by

SSS OSS has yp Ee 2 el
ae} Ne

ve
S =a fe Ml
S| ae = as E er
“
Pipl ie :
SS

Life i= = =

== 9_ ira ey

| a
aes

Fic. 50.—Norwalk compressor.

Straight-line Compressor.—The balance is aided still further
by a fly-wheel which with shaft and eccentrics is used to operate
the valves. It is evident that an air compressor which has the steam
cylinder and the air cylinder on the same piston-rod will apply the
power in the most direct manner and will involve the simplest
mechanism in construction.

This type of compressor (Fig. 51) is usually referred to as a
straight-line air compressor and is usually equipped with one or

EFFICIENCIES AND ENERGY COMPENSATION 85

two fly-wheels to act as energy compensators. Even then it is
difficult to secure a very good economy, especially with light fly-
wheels. In order to secure maximum economy of steam an early
cut-off is desirable, but if no fly-wheels are used this cannot be
obtained, and it is necessary to admit boiler steam for almost the
entire stroke.

The air compressor used by the Westinghouse Air Brake Company
in their familiar system of train brakes is of this type. It is admira-
bly suited for this purpose because of its simplicity and the fact

Fic. 51.—Straight-line air compressor.

that it does the most work when the engine is at rest or using only
a portion of its steam, and for this reason it utilizes steam that
might otherwise escape out of the safety valve.

Many efforts have been made to equalize the steam power and
air resistance by using a crank shaft and placing the crank pins
of the steam and air-connecting rods at an angle with each other
so that the greatest force would be exerted in the steam cylinder
at the time the greatest resistance was being encountered in the
air cylinder. The same thing may also be accomplished by placing
the cylinders at an angle with each other. Various compressors
have been built on this principle, the angle between the cylinders
varying in different designs, being in some 45 degrees, in others
go degrees, and in still others 135 degrees. The best results, how-
ever, have been secured with an angle of go degrees.

This arrangement has been adopted by some manufacturers
of compressors for refrigerating plants, but has not been used by
manufacturers of air compressors to any extent. Fig. 52 may make

86 AIR COMPRESSION AND TRANSMISSION

this clearer, with the horizontal cylinder for steam and the vertical
one for the ammonia compressor. When the steam piston is at
dead center the air piston has completed about half its stroke,
and the high steam pressure admitted to the steam cylinder during
the first part of the stroke will be available for moving the compressor
piston through the last half of its stroke when the greatest resistance
is encountered. As the steam piston is completing the last half
of its stroke, the compressor piston starts down compressing a new
supply of free air on its lower side if of the double acting type, and
as the work of the first half of the stroke of the air piston is com-

CHEAT TEC EOL CUCE CAE COLEEOEEGy

Fic. 52.—Horizontal-vertical arrangement of cylinders.

paratively slight, the pressure in the steam cylinder can be reduced
for the last half of its stroke, giving both economy of steam and
uniformity of speed.

Duplex Compressor.— More frequently this result is accomplished
by placing the two cylinders in a horizontal plane with the crank
pins at an angle of 90 degrees as shown by Fig. 53. This arrange-
ment is frequently adopted when air compressors are driven by
gas engines and if an air compressor is driven by a belt the compressor
will operate much more evenly and hence with a more uniform pull
on the belt if two or more cylinders are used with the crank pins
of each placed at an angle with each other.

The “duplex air compressor” is designed on this plan with two
cylinders side by side, the crank pins for the two compressors
being at an angle of 90 degrees with each other. The motive

EFFICIENCIES AND ENERGY COMPENSATION 87

power may be either belt, electric motor or steam engine. If the
latter, it is not uncommon to place the steam cylinders tandem
with the air cylinders, using a common piston rod, as shown in
Fig, 54.

NST SY NL__ift
eS TM Vy

—Z
J

a} N = sl | Dy

Fic. 54.—Duplex steam-driven compressor.

The steam cylinders may be either cylinders of a cross-compound
engine, or two separate simple steam engines. Similarly, the two
air cylinders may be cylinders of two separate air compressors or

88 AIR COMPRESSION AND TRANSMISSION

cylinders of a two-stage compressor. The name ‘“Duplex”’ is
applied to any of these designs.

Figure 55 shows a sketch of the arrangement of cylinders for
a two-stage duplex compressor driven by a cross-compound steam
engine.

A little study of these sketches will make it clear that with such
a duplex arrangement when the greatest power is developed in
one steam cylinder, this excess power can be utilized by means
of the common crank shaft in overcoming the maximum resistance
that is being encountered in the other air cylinder. With the cranks

Fic. 55.—Duplex cross compound steam, two-stage air compressor.

go degrees apart there is little difficulty in starting, even if compound
steam cylinders are used, for if the compressor would stop with the
high-pressure cylinder at dead center, live steam may be admitted
to the low-pressure cylinder by means of a by-pass.

Commercially, the duplex compressor appeals to the trade in
that one side or half of the machine may be furnished with fly-
wheel and out-board bearing designed for a complete machine,
and as the demand for compressed air increases, the output may
be increased by installing the remaining side of the machine.

The belt compressor is probably the best type for small capacities
when it can be used conveniently, as is the case in a great many
factories, for the losses in a steam cylinder, especially of small
power, are excessive as compared with the loss of power due to
belt transmission.

Cit’ ili Re EX

MULTI-STAGE COMPRESSION

It was pointed out in Chapter VII that it was not advisable to
attempt compression above 80 lb. per square inch in a single cylinder
because of the loss of energy and danger of explosion due to the re-
sulting high temperatures. 7

It frequently happens, however, that pressures much higher than
this are demanded for commercial purposes, and in order to satisfy
this demand, avoid the danger just referred to, and reduce the losses

MEREBEU AS ses

SSS SSS

b
f
q
i

f es
je OPO La
=e SESE EELS ERASE LUE UNEENANUNENENENESSND e

Fic. 56.—Saving due to multi-stage compression.

due to adiabatic compression, engineers have adopted a multi-stage
system of compression; compressing the air partly in one cylinder,
passing it through an intercooler where its temperature and volume
are reduced, then compressing it still further in a second cylinder,
and, if the pressures required are high, this compressed air is passed to
a second intercooler, thence to a third cylinder and in some cases a

90 AIR COMPRESSION AND TRANSMISSION

third intercooler and a fourth cylinder are required to secure the
desired compression pressure economically.

Advantage of Multi-stage Compression.—The advantages of
this system of compression more than offset the extra expense in
constructing the compressor. The saving in power required may be
illustrated by Fig. 56, where a—d represents the adiabatic line from
atmospheric pressure to the required receiver pressure, a—c an iso-
thermal line between the same pressures. The shaded area repre-
sents the total work of compression in the four cylinders, the differ-
ence between this area and the area abde representing the saving in
power due to the multi-stage system of compression. afne repre-
sents the work done in the first cylinder, fn the volume occupied by
the air as it leaves this cylinder. In the intercooler the temperature
of the air, if this part of the apparatus is properly designed, will be
reduced to the inlet temperature, and in consequence the volume will
be reduced from fx to on. Compression in the second cylinder will
raise the pressure to g and reduce the volume of the compressed air to
gm. In the second intercooler the volume will be reduced as the
temperature is reduced to the inlet temperature from gm to pm, and
so.on. ‘This secures a compression that requires a smaller expendi-
ture of energy than adiabatic compression, giving results that com-
pare very favorably with the ideal isothermal compression without
serious difficulty.

Pressures Used for Various Stages.—Of course this arrangement
increases the first cost of the compressor and for that reason the ad-
visability of installing multi-stage compression will depend upon the
pressure required. Some authorities recommend two-stage com-
pression for pressures as low as 50 lb., but this practice is unusual.
It is certain, however, that for pressures from 80 to 500 lb. the two-
stage compressor should be used; for pressures from 500 to 1,000 lb.
the three-stage, and for pressure between 1,000 and 3,000 lb. the four-
stage compressor.

Intercoolers.—To secure best results care should be taken to see
that the intercooler between the different cylinders reduces the
temperature of the air as nearly as possible to that of the air at the
compressor inlet. As it is important that the flow of air through
the intercooler should be as low as possible, it is desirable to reduce
the pulsating effect of the discharge of partially compressed air to
the intercooler. This is usually accomplished by using large ports
and passages.

The larger the volume of the intercooler, the more time for the

MULTI-STAGE COMPRESSION

Fic. 57.—Intercooler for duplex two-stage compressor.

Wa ter Outlet

>
KS

as See SS | See
Uy N=
UJ
4
U

.
eet oe ak
id Site esses SY

eee sf
(e ~]

LO: meas ee ‘Of
$22 Yor
Ogee Ss (|
wz J

cc OL Ne

(a hig a\'=|
YTZZALLIP OTP LIZZ ne

ZZA Rae.
ML
T
| y
i

Cima waa waar ar ae

ll Robeech

ZA U
— if N MY]
eS
<a Ss
a

ee ey

howe
‘AMIN KADY
SS Pe rae \ Z) '
AN 7
is
H Ly
4
i wy wanes
1 GN NOE
f Zia tH Tomi ah =
‘a
HAT SS
Dp in
az eee:
Ac LX ‘

Sy H
we ees Jacket Pipe-
? Air Discharge///////7

Fic. 58.—Intercooler for tandem two-stage compressor,

92 AIR COMPRESSION AND TRANSMISSION

compressed air to cool; for this reason ‘‘receiver intercoolers,” as
they are called, are more efficient than those of small volumetric
capacity.

Fic. 59.—Nordberg intercooler.

Types of Intercoolers.—Figs. 57, 58 and 59 show various types of
intercoolers. The horizontal type is more frequently used because
of its greater adaptability to
compressor construction.

In accordance with the
fundamental principles of
economical transference of heat,
it is customary in better types
of intercoolers to have the cir-
culation of the water opposite
in direction to that of the air
being cooled, and also to have
ond i the air broken up into as fine
Ce ir L—- streams as possible.

fi ¥ The tubes of intercoolers are
usually of iron unless the

character of the water used is
bad. In this case the tubes
may be galvanized or, if the
water is salt or contains mate-
rials having a corrosive effect on iron, brass or copper tubes are used.

In cooling the air, moisture is frequently deposited, and provision

Fic. 60.—Intercooler with separator.

MULTI-STAGE COMPRESSION 93

is made to remove this by traps or separators, as shown in Fig. 60.

Perfect intercooling implies that the temperature of the partially
compressed air leaving the intercooler shall be as low as the atmos-
phere. This naturally requires different ratios of cooling surface to
cubic feet capacity for different water temperatures.

Cooling Surface and Capacity.— Mr. F. V. D. Longacre gives two
charts covering this matter, shown in Figs. 61 and 62. The first
shows the intercooler surfacer equired for various water tempera-
tures to secure perfect intercooling for two-stage compression to 100
Ib. discharge pressure at sea-level; and the second shows the amount
of water required to secure perfect intercooling for this pressure if
the cylinder jackets and intercooler are in series, also the amount of
water required with a separate jacket, and when the low-pressure
and high-pressure jackets are connected in series.

Intercooler Pressure.—In considering multi-stage compression,
it is necessary to determine the proper intercooler pressure to secure
the most economical results.

It was shown in Chapter IV that the area of an ideal indicator
diagram disregarding clearance could be expressed as

nN Po na
—"— 144 paVe| (22) -| ft-lb.

and if Va represents the capacity of the machine in free air per
minute, the horse-power required, disregarding friction and other

losses, will be:
144 Po
33000 rane aVa (ere -:|

If pp represents the intercooler pressure, this would represent the
horse-power required to operate the low-pressure cylinder, and if
the discharge pressure from the high-pressure cylinder be represented
by pa the horse-power required to operate this cylinder would be:

144 n pa —
33000 pave (22 |

but if perfect intercooling were secured, poVc would equal paVa
and the total horse-power required to operate both cylinders of
the two-stage compressor would be cee

Petr nt paVe| (22) "= + (2) "* =.

94 AIR COMPRESSION AND TRANSMISSION

S 20 Curve showing intercooler surtace|

4 10 100 |b. discharge pressure at sea

c- ae level to obtain pertect intercooling. ee eae
22)

a= |b

z SCS et eee
£ ad
= [so one ase ae
= Tee eee a
Z, Bae eee Bere
‘ a [ES ete hit] | a
t ea [9 5] a a a
8°9 5 45 50

35
Dereace ieniperarre pee Air Peand and Water enone

Fic. 61.—Intercooler surface required.

IS
cS

BS
i Se |

D
ro

Gallons of Water per Hour

Volume of water Sanilgd
for perfect intercoo

100 Ib. disch. pressure

sea level operation Wath
and without cylinder

Jacket water.

ay, on

Fala

ELE
280

Siew
=o an ea
v 260
77 a
(allel a NR ESE
sca)
tp
Coy ye
lay Ae
sae

BOP ROES
eee ine |
SSE

120

VALS
PAV
BBVA

90
ewetee rCiaiestern beneee Fahrenert

Fic. 62.—Water required for intercooling.

MULTI-STAGE COMPRESSION 95

This expression will be a minimum when the part within the
brackets is a minimum. As fa, or the inlet pressure, and pa, the
discharge from the high-pressure cylinder, are fixed, the only
variable is the intercooler pressure, or po.

(jak n—1
Differentiating eo) Oe ian cp ies | with respect to pp and

equating to zero will give the requisite condition of proper inter-
cooler pressure for minimum expenditure of energy.

n—1 in n—1

1h el fh —1 I Aba AP === == 1
Pay 3 ee + Se, Ci =o
Raa
Laas’ a =— in I TNS =
: Did ae Doak ee Ts - py
HES
ny pe n b—n 1—n
Pe = a5 a aSpE De wD ee
po=NV Paba

That is, for two-stage compression the most economical expendi-
ture of energy is secured when the intercooler pressure is the square
root of the product of the given suction and discharge pressure of
the machine. As perfect intercooling is assumed, paVa=poVo
and the areas of the two cards must be equal, that is, the most
economical results are secured when the work of compression is
divided equally between the two cylinders.

Let Fig. 63 represent an ideal card of a three-stage air compressor
without clearance, in which p3 represents the pressure in the first
intercooler, and 4 the pressure in the second intercooler. These
first two stages may be considered as two-stage compressors between
bp; and p,4 in which, for the most economical results,

b3s=WV pips

and in the same way and for the same reason,

pi=V Papo
from which

p3= V obo
and

pa=V pips?

The effect of clearance on the above discussion can be shown
by referring to Fig. 64, showing cards for a two-stage compressor

96 AIR COMPRESSION AND TRANSMISSION

with clearance. The area showing the work done is AJKLSTZ,
which may be considered as AJNF+KLEN—ZSEF, This will
evidently be a minimum when the expressions for these areas are
a minimum, but as the expression for ZSEF does not contain the

Pon a gs eae

me--—---——-—— — --— -— - -— -

| Lo
o
NY, Se ee ee eee

Fic. 63.—Proper receiver pressure for multi-stage compression.

variable pz, this term will drop out in differentiating and, as a
result, it will follow that the intercooler pressure giving the most
economical result will be with clearance as without clearance.

Px = iy

Fic. 64.—Effect of clearance on receiver pressure.

The same method will show that most economical receiver pres-
sures for a three-stage compressor are:

p3= V pbs and p4= V pipe?

MULTI-STAGE COMPRESSION 97

when clearance is considered as when clearance is omitted in the
discussion.

Effect of Clearance on Volumetric Efficiency.—It was pointed
out in Chapter VIII that the real volumetric efficiency of an air
compressor could be expressed as

re helmet]

Figure 64 has assumed the clearance in the various cylinders
to be proportional, that is, the ratio of clearance volume to piston
displacement in each cylinder was such that clearance lines of each
cylinder unite to form a continuous expansion line.

If C represent the clearance of the low pressure cylinder and po
represent the intercooler pressure, pa the suction pressure and pe the
discharge pressure from the high-pressure cylinder, then the real
volumetric efficiency of a two-stage air compressor may be ex-

pressed
I Bpe 4 pi oe Po a |
re Batee| eI]

but as Po=N Pabey this may be written

Tam pi'| | (Patbe?\ > |)
Tr fe: el ( Pa. es
Tom Pr}, _¢| (Pe) on— ||
arena aa}

In the same way, the true volumetric efficiency of a three-stage
air compressor may be expressed as

Tam Ps is| (2 nor]
16s Pam | - Pa J
in which fe is the discharge pressure from the last cylinder and pa
the suction pressure of the low-pressure cylinder.

Figure 45 shows graphically the effect on volumetric efficiency
of compressing by stages and the resulting advantage in capacity.

CHAPTER X

DETAILS OF PISTON AIR COMPRESSORS

Classification of Valves.—Most of the various types of inlet
valves for piston air compressors may be divided into two general
classes: first, those which are automatically opened by atmospheric
pressure and closed by means of their own inertia or weight, by
springs, or by air pressure; and second, those which are opened
and closed by direct and positive mechanical connection with the
crank-shaft or some other moving part of the machine. Each of
these classes include many forms of valve design.

Valves of the first class are entirely automatic in their action,
their opening and closing points depending entirely upon the con-
ditions of pressure within the cylinder. However, they have
certain advantages which will be considered later. Valves of the
second class, with one or two exceptions, have their points of opening
and closing fixed without regard to changes in operating conditions,
and the present tendency among designers and manufacturers
seems to be toward valves of this class.

Mechanical Valves.—Nothing can be superior to mechanically
operated valves when properly adjusted to operating conditions,
as by their aid several of the losses of air compression have been
reduced to a minimum.

On the other hand, faulty adjustment of valves, sometimes
combined with improper design, renders them extremely low in
both efficiency and capacity.

Inlet Valve Setting.—If inlet valves are so set that they open

almost exactly when the piston is at the end of its stroke, the card
will indicate absolutely no clearance at either end of the cylinder,
the clearance air being exhausted into the intake. If the inlet
valve closes slightly before the piston reaches the end of its suction
stroke, the volumetric efficiency is also reduced.

In case the inlet valves are so constructed that they cannot open
until the clearance air has been expanded to atmospheric pressure,
the only loss due to this clearance is one of capacity, which may be
Overcome by an increase in size or speed of piston. If, however,

Pt yer

DETAILS OF PISTON AIR COMPRESSORS wh

the inlet valve opens when the piston is in its extreme position,
the clearance air is exhausted through the intake, making a direct
loss of power as well as of capacity.

Figure 65 is reproduced from a card taken from a machine in
which the inlet valves were set to open when the piston was exactly
at the end of its stroke. In most of these cases the exhaust through
the intake is sufficient to cause considerable noise. Figure 66 shows
a card from the same machine with the valves set properly for their
particular pressure. This change in the time of opening the inlet

pm.

Fic. 65.—Mechanically Fic. 66.—Mechanically oper-
operated inlet valve opened at ated inlet valve properly set.
end of stroke.

valve has not effected the volume of air discharged, but the power
required to operate the compressor has been considerably reduced
and the machine will run more smoothly with less shock to the
moving parts at the end of the stroke.

Effect of Changing Discharge Pressure.—lIf the pressure of dis-
charge is now increased, the former troubles appear again, resulting
in a card shown by Fig. 67. If this pressure is to be maintained con-

Z)_4

Fic. 67.—Effect of increasing Fic. 68.—Effect of decreas-
discharge pressure. ing discharge pressure.

tinuously, the inlet valve will have to be adjusted to open a little later
in order to give the best results.

In the same way, if for any reason the discharge pressure should
be reduced after the valves have been set correctly, the indicator
card will resemble Fig. 68, and if the compressor is to operate con-

”

100 AIR COMPRESSION AND TRANSMISSION

tinuously at this lower pressure the inlet valve will have to be ad-
justed to open a little earlier.

Figures 69 and 7o show indicator cards from improperly set
mechanically operated discharge valves with the defect indicated
under each.

Every machine with mechanically operated valves should be
carefully examined to determine whether they operate at the correct

Fic. 69.—Mechanically Fic. 70.—Mechanically
operated discharge valve operated discharge valve
opening too early. opening too late.

time, andif not, should be so adjusted in order to raise the efficiency
of operation.

The principal disadvantages of the mechancally operated valves
are the increased cost and the extra attention required to keep the
valves set properly.

Automatic Valves.—In compressors using automatic valves, how-
ever, there is no necessity of timing the valves to suit changes of
pressure, as the operation of the valves is controlled entirely by the
conditions or pressure within and without the cylinder.

It must be remembered that the majority of air compressors
operate at a fixed pressure of discharge and after the valve is once set
for this discharge, there is no further need of changing it.

Both types of valves have advantages and disadvantages peculiar
to each and a choice of valves should not be made in any important
instance without a thorough investigation of all the variable factors.
involved.

Valve Area.—A very important matter to be considered is the
inlet valve area or port opening required for the proper action of a
machine. Asin other points of design, it is necessary to compromise
the desired ends, for the larger the inlet valve the less will be the
water-jacketed cylinder surface, and as both are desirable it is im-
possible to give absolute ratios of inlet areas to cylinder sizes. Some
designers make inlet areas 5 per cent. of the piston area, and other

DETAILS OF PISTON AIR COMPRESSORS 101

designers use as high as 14 per cent. asthe ratio. The design of the
valve, the cylinder proportions, and the speed of the machine, all
have an influence in determining this point.

The following data is given by the chief draftsman of a large com-
pressor company as the practice of that company resulting from an
experience of many years:

“Roughly speaking,” 5,000 ft. per minute for the velocity of the
air through the valve gives good results. This being the case, a
slow-running machine would require a smaller valve than a high-
speed compressor with a ‘piston-inlet’ valve, having a piston speed
of from 300 to 350 ft. per minute, the inlet area is from 5 to 6 per
cent. of the piston area. On large compressors with a piston speed
of from 500 to 600 ft. per minute, the valve area ranges from 6 1/2
to 7 per cent: of the piston area.

The discharge valves which are of the poppet type are from ro
to 12 per cent. of the piston area.

On machines having both inlet and discharge valves of the poppet
type, the ratio should be about 12 per cent. for machines of that
speed. For piston speeds not exceeding 4oo ft. per minute it is
probable that ro per cent. is enough.

The area of the discharge valve should not be less than that of the
inlet, for although the volume of discharge is less than the volume of
admission, this discharge must take place in a considerably shorter
space of time.

Forms of Poppet Valves.—Probably the automatic poppet valve
is the most common form of valve in use. A few designs are shown
in Figs. 71 and 72.

Fic. 71.—Air inlet valve.

The principal difficulty to guard against in the design of an auto-
matic valve is to avoid the possibility of the valve itself being
drawn into the cylinder with the in-rushing air. This may happen
through the breaking of the spring and disastrous results frequently
happen, for on the return stroke of the piston, the cylinder head,
or the piston, or some other part of the apparatus is sure to suffer.

Figure 73 illustrates a peculiar valve designed for a single-acting
compressor, 7.¢., a type of piston compressor in which the air com-

102 AIR COMPRESSION AND TRANSMISSION

pression takes place on only one side of the piston instead of both as
is usually the case. JB is the inlet valve which is located in the center

Fic. 72.—Air discharge valve.

of the piston and is held on its seat by the spring D. The discharge
valve A is a radical departure from the older designs of compressor

Fic. 73.—Valve in cylinder fee

valves, being a flat disc covering the entire area of the cylinder and
held in its seat by a guide and spring.

AS
AA rz raedsaiteg ae
Ee
a = ——ews

AL OH a A fa

om |
zee ‘* 4
is Ae
Se iA eek ay
S

Le ed

N
Vy

GPR

iN
yee

Fic. 74.—Piston inlet valve.

Its face and the face of the piston are perfectly flat, so that the
piston may strike the valve and deliver all the air with no clearance

DETAILS OF PISTON AIR COMPRESSORS 103

space to reduce its capacity. A large area of discharge is obtained
by a very small movement of the valve and no pounding is made
by its action, for the compressor is of the straight-line type and the
compression in the steam cylinder acts as a cushion to relieve the
pounding that might otherwise occur with no clearance.

Piston-inlet Valves.—One of the most interesting forms of inlet
valve is the piston-inlet valve, as manufactured by the Ingersoll-
Rand Company, a sketch of which is shown in Fig. 74. By
this arrangement the entering air comes in through the tube 4,
which projects through the head end of the cylinder. The air
passes through this to the center of the piston which is hollow.
Communication is obtained from this hollow piston to the cylinder
through the ring-shaped valves B, which are made of open-hearth
steel in one piece without a weld. These valves have a movement
of about 1/4 in., and are held in place by pins which are set in slots
in the valve.

The two inlet valves B and the tube A are carried back and forth
with the piston. The valve on that face of the piston which is
towaid the right is closed as the piston moves to the right, while
that on the left side is open to admit a fresh supply of air to the
left side of the piston while air is being compressed on the right
side. -

When the piston reaches the end of its stroke, the inlet valve
closes because of its own inertia and as the piston starts on the return
stroke the valve that was formerly closed is now left behind for about
1/4 in. of the piston travel and remains open during the entire
stroke. When the valves are closed, their face is almost flush with
the piston face, thus reducing the clearance space to minimum.

There are no springs in the construction of this valve, and it has
been found to work equally well with slow or high speed. These
valves are guaranteed by the company for five years.

Discharge valves are shown at H. These conduct the com-
pressed air to the discharge pipe F.

All these types of automatically operated valves have the ad-
vantage that they adjust themselves to meet varying changes in air
pressure automatically.

Semi-mechanical Valves.—There are several types of semi-
mechanically operated valves on the market. Some of these con-
sist essentially of an arrangement of levers to remove the action of
the spring on the inlet valves during admission, permitting the valve
to open instantly and freely and remain open without any clattering

104 AIR COMPRESSION AND TRANSMISSION

until the end of the stroke, when the spring tension is permitted to
act on the valve and close it.

Attention has already been called to the fact that many auto-
matic valves close before the end of the stroke, due to the fact that the
piston is rapidly reducing its speed at that time. If the inlet valve

Fic. 75.—Mechanical valve of Corliss type.

closes before the end of the stroke, the volumetric efficiency is natur-
ally reduced, and on this account the mechanically operated inlet
valve is preferred by many engineers.

In addition to the mechanical operation of poppet spring valves
just mentioned, some air compressors are equipped with valves

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Fic. 76.—Southwork blowing engine valve.

which resemble in action and appearance Corliss steam valves.
Fig. 75 gives an illustration of this form of valve, which is opened
and closed by a rotating motion, given to them by levers from a
wrist plate or eccentric.

This type of valve is sometimes used for the discharge valve on
compressors, but cannot be operated successfully for very high

DETAILS OF PISTON AIR COMPRESSORS 105

pressures because the clearance is made excessive. Mechanically
operated valves are usually used on large blowing engines for blast
furnaces. One type is shown in Figs. 76 and 77. In this type of
compressor, it is desirable to secure large, free opening for suction,
and one of the latest designs consists of a large cylinder on the out-
side of the compressing cylinder which reciprocates back and forth,
and in so doing opens large slots at the end of the cylinder, giving
very free opening for inlet.

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Fic. 77.—Kennedy blowing engine valve.

Regulators, Unloading Devices, Etc.

It is often essential that the pressure of air in an air receiver be
kept constantly at a fixed point and as the number of tools using
air at the same time will vary in any installation, some automatic
device must be used so that the_compressor will be furnishing air
when needed and when no air is needed this device must prevent any
unnecessary work being done at the compressor. There will be
times when every tool that is taking air from the receiver will be in
operation and the compressor must have a capacity sufficient for
such occasions; and again there will be times when none of the tools
are in operation and the work at the time being done at the com-
pressor would be in excess of the needs if some automatic system
of regulation is not used.

Belt Regulator.—Probably the simplest form of regulator is the
one that is often used on belt-driven air compressors. ‘This consists
of a belt-shifting device so arranged that when the pressure gets
above the desired point the belt is shifted off the compressor wheel
and onto a loose pulley. When the pressure falls below this fixed
point the belt is shifted back again, and the compressor is thus

106 AIR COMPRESSION AND TRANSMISSION

automatically started and stopped to suit the changing amount
of compressed air that is needed.

Westinghouse Governor.—Fig. 78 shows a sketch of the governor
used on the Westinghouse air brake. It consists of a piston A
moving in a cylinder and directly connected to the steam valve C
which supplies steam to the air compressor, or air pump as it is
more commonly called. A spring, D, helps to hold this up and
hence keep the steam valve open. Pipe £ leads from the air reser-

%, D>

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Fic. 78.—Westinghouse governor.

voir to the governor. Communication between E and the cylinder
above A is closed by a needle valve Ff, which is held on its seat by
the governor spring K. When the air pressure in the main reservoir
gets up to its maximum, the pressure in £ is sufficient to raise the
small piston H against the governor spring K, lift Ff from its seat
and allow the air to press on A and thus close the steam valve and
stop the air pump. A small opening, L, allows the air above A to
escape gradually into the atmosphere. As air is used in releasing

DETAILS OF PISTON AIR COMPRESSORS 107

the brakes, the pressure in the reservoir will be reduced, and when
this happens, spring K can overcome the air pressure and seat F and
spring D will then raise piston A, open the steam valve C and start
the pump.

The governor used for controlling the compressor for electric-
driven air-brake systems consists of an ordinary Bourbon pressure
gage with a special needle or hand, which upon coming in contact
with a stud at the position of minimum pressure causes an electric
current to flow through a magnet coil. This coil operates a plunger
to which the contact pieces for the motor circuit are attached and

MeL

Fic. 79.—Belt regulator.

in this way the circuit is closed and the motor started. As soon
as the air pressure reaches the desired maximum the gage hand
strikes another stud, causing current to pass through a second
solenoid magnet which pulls the plunger referred to in the opposite
direction and stops the compressor motor.

By this mechanism it is possible to get a close margin between
the maximum and minimum pressure. This margin can be changed
by moving the studs.

Figure 79 shows a form of regulator for belt-driven compressors
~which stops the compressor when the desired maximum pressure

108 AIR COMPRESSION AND TRANSMISSION

is reached. When the desired upper limit is reached the belt is
shifted from the tight to the loose pulley.

Norwalk Regulator.—One of the simplest forms of regulators
for steam-driven air compressors is the one made by the Norwalk
Iron Works shown in Fig. 80. It consists of a balanced steam
valve A placed in the steam-pipe near the steam cylinder and con-
trolled by the air pressure in the receiver. A small cylinder B

contains a piston connected with the
opp balanced valve A by the stem C. Above
s this small piston is a stop screw D pro-
jecting above the cylinder head for regulat-
ing the lift of the piston. The air from the
receiver is led through a small safety valve
E which regulates the pressure at which air
can escape into the cylinder B to move the
piston. Above the disc of the small safety
y valve is a spring whose tension is regulated
i by ascrew F allowing the pressure at which
air is permitted to enter cylinder B to be
changed at will. The air passes into cylin-
der B below the piston and if no escape
were provided would drive the piston to
the top of B. To regulate this a very fine
slot is cut in the side of the small cylinder. When the piston
rises it uncovers this slot and thus furnishes an escape for the
air which is passing the safety valve. If only a little air enters
then a small part of the slot will accommodate it and the
piston will take a low position. With more air escaping the piston
will rise higher and uncover more of the slot, thus providing a
larger opening for its exit. As the slot is very fine, a very little
difference in the quantity of air will cause the piston to assume a
high or low position. After the small safety valve begins to blow
an almost insensible increase of pressure in the reservoir will furnish
enough more air to carry the piston to the top of the cylinder.
Thus any degree of regulation is obtained by a very little difference
of pressure, as the air which works on the piston in the small cylinder
has only to perform the work of lifting the piston and valve suff-
ciently to uncover enough of the slot so that it can escape; its pressure
is very slight.

The piston is fitted loosely and the whole apparatus moves as

nearly without friction as can be imagined.

aaear
Hy
sere
Hj

TI tis

Fic. 80.—Norwalk
governor.

DETAILS OF PISTON AIR COMPRESSORS 109

When this regulator is applied to compressors having a single
steam cylinder, it is possible for the valve to be carried so high as to
cut off all steam and to stop the engine on the center. This would
be objectionable. To obviate this, there is placed on the top of
the small cylinder a screw stop which can be set to prevent the
closing of the steam valve more than is sufficient to run the engine
at the slowest possible speed.

Combined Governor and Regulator.—Another combined speed
and air-pressure governor is shown in Fig. 81. This not only

Breas

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ISSISSEB SS

Fic. 81.—Clayton governor.

performs the functions of an air governor, but also prevents the
compressor from operating at an injurious speed should a sudden
drop in the air pressure produce a greater demand upon the com-
pressor than its highest reasonable speed can supply. It consists
simply of an adjustable stop attached to an ordinary centrifugal
ball governor. This stop is adjusted to suit varying pressures of
air in the receiver caused by the varying demands that are made on it.

Nordberg Governor.—A combined air and speed governor manu-
factured by the Nordberg Manufacturing Company of Milwaukee,
Wisconsin, is shown by Fig. 82. In this type of governor the speed

110 AIR COMPRESSION AND TRANSMISSION

of the engine is controlled not only by the centrifugal action of the
governor but also by any variation of the air pressure. The arm A
controls the point of ‘‘cut-off”’ for the steam cylinder and is operated
by the movement of the bell-crank C about the fixed point D. The
rod E controls the bell-crank and is connected to what is called a
floating lever B. This lever B is con-
nected with the centrifugal governor FP
at J, and with a piston which is in com-
munication with the air pressure at G and
is held up by a weight ZH.

It is evident from this arrangement
that if the point G should remain station-
ary and the point J should lower, rod E
will be forced downward and Ato the
right; also, if point J should remain sta-
tionary and point G should rise, the same
movement will occur, and vice versa.

That is, if the air pressure should rise
above normal, the engine will have its
supply of steam per stroke reduced, and
if the air pressure should fall, the supply
of steam will be increased; or, if the pres-
sure of air remains constant, the governor
will have the same control over the speed
of the engine that the ordinary centrifugal
governor has.

Fic. 82.—Nordberg Unloading Devices.—Sometimes an air

governor. compressor must be kept running at con-
/ stant speed, and in order to prevent it
from doing unnecessary work when the consumption is not equal to the
capacity of the compressor, a device is used to remove the work or
load of the air piston and allow it to move back and forth in its cylinder
without doing any work. ‘These are called unloading devices. Fig.
83 shows the principle upon which many of them operate. 61 repre-
sents a valve on the inlet pipe which is closed when the load is to be
removed, preventing air from entering the cylinder. These unload-
ing devices are frequently made use of in starting air compressors
without any load until full speed is reached, when the load is put on
as desired.

Clearance Unloader.—One of the most recent unloading devices is

arranged to vary the clearance on the compressor as the load changes.

DETAILS OF PISTON ALR COMPRESSORS 111

This device is illustrated in Fig. 84 and its effect or operation is
shown by the indicator cards of Fig. 85.

This automatic clearance controller, as it is called, consists of a
number of clearance pockets which are thrown automatically into
communication with the ends of each air cylinder in proper succession,
this process being controlled by a predetermined variation in re-
celver pressure. _

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Fic. 83.—Rand imperial unloader. 41, unloader body; 49, unloader cup
leather; 50, unloader follower; 51, unloader follower screw; 52, unloader cylinder;
53, unloader cylinder cap screw; 54, unloader regulator cylinder body; 55, unloader
adjusting plug; 56, unloader valve spring; 57, unloader piston; 58, unloader inlet
plug; 59, unloader nipple; 60, unloader dirt collector; 61, unloader plunger.

Regulation is obtained in five stages, viz., full-load; three-quarter-
load; half-load; quarter-load, and no-load.

When the compressor is operating at partial capacity, the clear-
ance space of the compressor is increased and, as a result, its volumet-
ric efficiency is reduced without changing the suction or discharge
pressures, or the speed of the compressor. —

On two-stage compressors these controllers are placed on both
cylinders, thus maintaining a constant ratio of compression. The
device is automatic and very satisfactory for use on compressors
which are motor driven, or must, because of their method of opera-
tion, be driven at constant speed.

112 AIR COMPRESSION AND TRANSMISSION

sey
= Y
oe

Fic. 84.—Clearance unloader.

Low Pressure Side

High Pressure Side
Scale 24 Scale 60
LHP 3323, AP Input38E
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Fic. 85.—Cards showing clearance unloader.

CHAPTER XI

TURBO-COMPRESSORS

The Engineering Magazine in recent issues has given a series of
five articles on turbo-blowers and compressors by Franz zur Nedden,
Superintending Engineer of Weise and Monski, Halle-Saale, Ger-
many, from which the following material and illustrations have been
gathered.

The introduction of the steam turbine as a competitor of the
reciprocating engine has necessitated a similar change in the design
and construction of pumps, blowers, and compressors and has natu-
rally led to the production of turbine machines built for compress-
ing air or gases.

The advantages of turbo-compressors, however, are not so apparent
in small as in large size units, and the high cost of such units in the
experimental development of this machine has made its introduction
rather slow.

The recent development, however, of exhaust steam turbines has
stimulated the use of turbo-blowers.

‘“‘Tt is well known that the economy of the steam turbine increases
directly with its rotative speed, and even electric generators of the
highest speeds are slow-running machines when compared with the
steam turbine operating at the number of revolutions required to
secure the best economy. The turbo-compressor, like the steam
turbine, becomes more and more economical the faster it runs, and
is therefore a proper companion of the steam turbine. Speeds of
4,000 revolutions per minute and even more are not unusual to the
design of turbo-compressors.

“Tf large volumes of compressed air are wanted in plants where
considerable quantities of exhaust steam are available at the same
time, the coupling of an exhaust-steam turbine with a centrifugal
compressor becomes an ideal arrangement, and the combination is
far superior in economy to the piston compressor driven by an
electric motor or a high-pressure steam engine.”’

Design of Turbo-compressors

“In studying the development of turbo-compressors it is most
interesting to observe that Rateau and Parsons dealt with the prob-
lems of design quite differently.

8 113

114 AIR COMPRESSION AND TRANSMISSION

‘“‘Rateau Blower.—Prof. Rateau did not take the structural
features of his new machine from his steam turbine, but from his
high-lift turbine pump. His turbo-compressor and turbo-pump are
so similar that a superficial inspection of the drawing of the two
~ machines might not reveal the difference (Fig 86).

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Fic. 86.—Original Rateau turbine blower.

“The air upon entering the impeller near its nave is seized by
the impeller blades and thrown outward radially. Its kinetic energy
due to velocity upon leaving the periphery is changed into pressure
in the fixed diffusor channel, and being led back toward the center
the air enters the second impeller to undergo the same process in a
second stage and so on. Some essential differences in the design
of details will be taken up later on.

“The Parsons Blower.—Mr. Parsons, on the other hand, made
the turbo-blower merely an inversion of his steam turbine. Fig. 87
shows a unit consisting of a standard Parsons steam turbine (on

a eae:

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wv ie z My \ Rspstissssscy SE)

AAO oti | \ “ft vi i
eth = fal ONIN Ne ANE

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laisipeseas oa d

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a £6 A 2 Ea TS | SS ay |
KOS ASES SE SSS iS
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ANGE

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Fic. 87.—Parsons turbine blower with steam turbine.

the left-hand side) coupled direct with a turbo-blower of the usual
Parsons type, and delivering 1,600 cu. ft. per minute against a pres-
sure of 6 to 20 in. of mercury at speeds varying from 2,400 to 3,400
revolutions.

“The air is drawn into the chamber B, and conducted into the
periphery blades of the runner A by fixed guide blades D. The
following blades are not shown in the section merely for simplicity
of outline.

TURBO-COMPRESSORS 115

“The principal divergence from the Rateau design is that the
impellers of the Parsons turbo-blower throw the air in an axial direc-
tion to the next guide apparatus. Parsons undertook to transform
the kinetic energy of the air as it leaves the impeller blades into
pressure by simply opposing plain straight blades against its flow.
The second guide-wheel transmits the air axially to the second
impeller, which again throws it axially into the third guide-wheel,
and soon. Fig. 88 shows a developed section through several rows
of blades. |

“The excellent reputation of the Parsons machines helped the
introduction of his turbo-blower, which was rapidly accepted and

PVE be

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| | | Guide Me

POON oe

Fic.” 88.—Developed section of Parson’s blades.

put into practical operation. At a time when more than a dozen
Parsons blowers were in operation or under construction, Rateau
was still experimenting with his first turbo-blower. Nevertheless,
Prof. Rateau succeeded in making up this delay and soon advanced
to the point of combining several of his blowers in series, thus pro-
ceeding to obtain final pressures of 100 to 150 lb. per square inch.
The excellent results which he and his assistant, Piof. Armengaud,
obtained from their high-pressure machines induced even the
licensees of Parsons steam-turbine patents to secure rights for the
Rateau turbo-compressors. A careful comparison of both systems
will disclose seme reasons for the rapid adoption of the Rateau
system.

“Cooling Turbo-compressors.—Increase of temperature makes
special cooling arrangements indispensable, especially with turbo-
compressors, 7.¢., with machines compressing air to more than 20
lb. per square inch absolute pressure. Economical cooling becomes

116 AIR COMPRESSION AND TRANSMISSION

a vital question in the thermal efficiency of the compressor. On
this point it decidedly excels the piston compressor, as it is impos-
sible to cool the air continuously when it is compressed in cylinders.
(See Figs. 89 and 90.) Here jacket cooling is the most important
part of the whole cooling system, and the special intermediate
coolers used between the separate cylinders of compound piston

Quant.
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Yolume, Cubic Feet per Haur

Fic. 89.—Diagram three-stage piston compression.

compressors are generally considered wholly unnecessary in the turbo-
compressor. In the turbo-machine compression of the air proceeds
much more gradually, the distance traveled by every particle of
air is consequently much greater than with piston compressors, and
the entire area available for the cooling influence of the water is

BUS
Va KK Vili
eee SSS Coble Feet wee

0 500 ; 1000 1500
Volume, Cubic Feet per Hour”

Fic. 90.—Diagram of tubo compression.

many times as large as that in the piston compressor of equal capac-
ity; therefore, the air pumped by the tucbo-compressor can be kept
at nearly constant temperature throughout the operation. Cooling
arrangements of the counter-current type can easily be used to give
maximum effectiveness, a condition not readily attained in com-
pressors of piston type.

TURBO-COMPRESSORS isles

COOLING DEVICES

“The principal differences noticeable between various turbo-
compressor systems are in their cooling arrangements. The various
licensees of Prof. Rateau do not use a uniform cooling device. Fig.
g1 shows one of the first water-cooled Rateau compressors which
has been successful in practical operation. Fig. 92 shows its internal
features. Each of the three groups coupled in series contains seven

oY Ni Nees iN MMos A)

Yi
YA ae :
Al aime 4

—— ns |
if = zm 7 ain iff i
Ue il, — LH Wii i

Fic. 91.—Water-cooled turbo compressor.

or eight stages. Each casing is separable horizontally into two parts,
a form which seems to have become standard for turbo-compressors
as it has for steam turbines. The cooling water enters the casing
from below at the highest pressure stage of the group. It passes
thence to the upper part through copper tubes, shown in Fig. 91,
goes through a core-hole at the top of the highest stage into the
upper half of the highest ons but one, and again through copper

“ HT bn '
ae _—————Sa
i oy

WT, ———— LLLLLLLLLLLLL LLL Ae

ee ecg er cnemene es teen ce: PAs0-a=>
:

Fic. 92.—Turbo compressor built by Brown, Boveri and Co.

tubes into the lower half, whence it goes through a core-hole into
the lower half of the next lower stage, and soon. Later, the copper
pipes were replaced by small bored holes passing through the
horizontal joints. (See Fig. 92.) This system has the disadvantage
that there is no assurance that the water shall completely fill up the
cooling chamber, as the core-holes are not always at the very highest
points of the chambers.

‘““Messrs. Brown, Boveri & Co. avoided this difficulty, and, more-

118 AIR COMPRESSION AND TRANSMISSION

over, greatly enlarged the cooling area by casting the vanes hollow,
thus leading the air back to the center and creating a separate inte-
rior cooling chamber B, Fig. 93. The water, after filling chamber A,
flows through the hollow guide vane into chamber B, and thence by

Pas ek pipe D (which is screwed into

highest point of chamber A of the
next stage. Though the excellence of
this system cannot be denied, it is, on
the other hand, very expensive, for the
castings become highly complicated,
and the foundry work, moreover, must
be such as to guarantee that all sur-
faces are absolutely smooth, as the
elle frictional resistance of air is dependent
on the roughness of the surface over
which it passes. The cost of casting
these casings was about 5 1/4 cents a
pound in place of 3 1/2 for average casings, and even then this
famous foundry could not avoid 15 to 20 per cent. waste.

E “‘C. H. Jaeger of Leipzig wisely separates the casing of his turbo-
compressors into as many chambers as there are stages, and screws
these together as shown in Fig. 94. Furthermore, he separates each

SSSSOQq

Y
A
Y
WV,
y
y
V)
mY,
j
bod)

Fic. 93.—Ipmroved cooling sys-
tem for turbo compression.

Sen wal NS avtall uu
Winch aaa wee Ne i ee a a= i ed,
META iy LLL LL LULL i! UML AA As ze
‘ oa = ee
IN = W\

ill
NS WA WA

Fic. 94.—Water-cooled Jaeger turbo compressor.

stage into an upper and lower half. Though this system increases
considerably the number of machined surfaces, it nevertheless
insures small casings and enables the maker to manufacture the
single stages on a large scale and to combine out of stock as many as
are needed for any special requirement.

“Expansion of Casing.—Another effect of temperature rise, and
one which acts on the machine, is the expansion of the casing by
heat. Special preventive measures must therefore be taken to avoid

TURBO-COMPRESSORS 119

any alteration in the relative position of the casing and the runners.
As the casing rests by lateral supports.on the bed-plate, the absolute
height of its center above that bed-plate will change as soon as
the casing becomes heated and expands. If the shaft, revolving
within this casing with a clearance as small as 1/1000 of an inch, is
supported by bearings which remain practically cool, the distance
between the center of the shaft
and the bed-plate will remain
unchanged. Therefore, the
centers of the casing and of
the shaft, which may coincide
when both are cold, must
differ as soon as the blower
comes into operation and the
temperature of the casing
rises. The clearances will
then not only become eccen-
tric, but very probably the
shaft and the vanes of the im-
pellers will come into close
contact with the fixed parts,
causing heavy friction, and
because of the absence of any
lubricating medium they would very likely seize.

“Designers of turbo-compressors have overcome the difficulty of
axial expansion by means already well known in the steam engine
and the piston compressor. These difficulties are, of course, larger
with long-extended turbo-compressors than with single-casing turbo-
blowers. The blower illustrated in Fig. 95 rests with one end only
on the bed-plate, the other end being free to move or expand. When
additional stages are required, the design must be altered tc a form
now quite generally adopted by the licensees both of Prof. Rateau
and Messrs. C. H. Jaeger & Co. The body is supported only by the
two terminal covers which carry the bearings. One of the bearings
is fixed rigidly to the bed-plate, while the other is allowed to slide
to some extent on the machine rest. (See Fig. 94.) With very long
blowers, it might perhaps be advisable to support the body by
lateral feet, the machined surfaces of which might move freely on
the bed-plate on exactly the same horizontal plane as the axis of the
blower. In this way radial expansion would not alter the height of
the geometric center of the casing above the bed-plate.

‘“fRunners.— Many of the problems that had been solved in design-
ing steam turbines assisted in the solution of the design of turbo-
blowers, but although Parsons was able to adapt his steam-turbine

S Wsr¥%
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/ Wi
Fic. 95.—Jaeger’s turbo-blower.

120 AIR COMPRESSION AND TRANSMISSION

runners to his blower, the original designs of Rateau and Jaeger had
unsymmetrical runners, which had a tendency to become deformed
under high speed. This difficulty was finally overcome by the use
- of impellers made with a solid hub with blades and lateral flanks of
pressed sheet-nickel steel, as shown in Fig. 96.

“It is not very difficult to insure tightness between the single
stages of turbo-blowers and turbo-compressors. The pressure

Fic. 96.—Jaeger’s patent impeller.

differences against which the clearances have to be kept tight are
comparatively small, as the delivery pressure is distributed over
many stages; and, on the other hand, long clearances are secured
almost automatically as the stages are placed one after another in
the casing. With air, as with water, the leakage resistance of a
clearance is greater the longer and narrower it is, but the problem

Z )
iMlalal:
Fic. 97.—Labyrinth bushing.

of maintaining tightness against leakage with gaseous media is
facilitated by the labyrinth effect which is utilized on a large scale
in the manufacture of steam turbines.

‘“‘Figure 97 shows such a labyrinth as used by Brown, Boveri &
Co., the action being briefly described thus:

“The air passing through the small clearance A expands as it
enters the following chamber B. By this expansion its pressure is
greatly decreased, and it traverses the second clearance C at a con-

TURBO-COMPRESSORS a

siderably lower pressure than that at which it passed through 4;
owing to the expansion in the second chamber D, still less pressure
is left for forcing the air through the third clearance F, and so on,
and it is therefore impossible for any considerable quantity of air to
pass the labyrinth.

“Balancing Axial Thrust.—The only point at which any great loss
of air occurs is in connection with the usual methods for balancing
the axial thrust. Special arrangements for this balancing become
necessary, for, as with centrifugal pumps, the annular area opposite
the entrance to each impeller is subject to a heavier pressure than
the entrance itself. And, as with centrifugal pumps, there are sev-
eral ways of obtaining perfect balance, perhaps the best being that
illustrated in Fig. 98.

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Fic. 98.—Turbo blower of 25,000 cu. ft. capacity.

“Balance by Counter-position.—Here the blower draws the air
from both sides, and delivers it after both halves of the entire quan-
tity have been compressed separately in wheels of the same dimen-
sions through which they pass in opposite directions. Thus, any
axial thrust arising in a wheel on one side is balanced by an equal
thrust exerted by a corresponding wheel on the other side. This
advantage of balancing axial thrust most perfectly, and practically
without any leakage losses, is paid for, however, in this instance by
a cumbersome arrangement and poor efficiency. It is obvious that
the use of double the number of rotating sheaves for compressing
the same quantity of air doubles also the amount of energy lost by
frictional resistance. In other words, a blower like that shown in
Fig. 98 is simply two blowers, each for half the delivery, coupled
in parallel; and, as the efficiency of all rotating machines drops with

122 AIR COMPRESSION AND TRANSMISSION

decreasing delivery, it is clear that the efficiency of these two blowers
must be lower than if the whole quantity were dealt with in one single
machine. .

‘Balancing by Diminishing Back Area.—Another design which is
standard in the blowers of some of Prof. Rateau’s licensees, for

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example the Skoda-Werke at Pilsen, and Messrs. Kuehnle, Kopp
& Kausch, at Frankenthal, is adopted from the well-known Rateau
turbine pumps. ‘The excess pressure acting on the back of each

TURBO-COMPRESSORS 123

impeller can be cancelled by simply diminishing this back area pro
rata with the increase in pressure. This is done by leaving free an
annular margin at the periphery of the back of each impeller as
shown by Fig. 99. The disadvantage of heavy wear and tear,
which appears when this system is used in turbine pumps, is certainly
very much less serious when atmospheric air is dealt with. Never-
theless, there is a certain loss due to the formation of vortices in the
cells of the impeller where the air is confined by a rotating disc on
one side and an immovable casing on the other side, and this defect
is unavoidable in this otherwise excellent balancing system.

‘“‘Balancing by Balancing Piston.—The design which is now prac-
tically standard, Fig. 100, is characterized by a continuous flow of
air through the impellers on one direction, while the consequent
axial thrust is reduced by other means.

‘“‘Beyond the last stage there is fixed on the shaft a piston which
is of the same diameter as the entrance of the impellers and extends
as far as possible axially. It revolves with a very small clearance in
a box containing a labyrinth, such as that described above.

“‘One side of this piston is under the full pressure of the last stage,
the other side being connected by a large pipe to the entrance cham-
ber. The effort which the pressures of all the stages exert on the
piston area is thus equal to the effort which the pressures of the
single stages altogether exert on the single impellers, and as the two
forces act in» opposite directions the balance is perfect. Here
again is another reason for limiting the number of stages that may
be coupled together in a single casing, for although it is possible
normally to keep the losses caused by the balancing piston of a
blower within limits of 1 per cent., this would not be possible if the
pressure differences were as great as 50 lb. or perhaps 100 lb. per
square inch.

“‘Stuffing-boxes.—The problem of a reliable, tight stuffing-box,
which is so difficult in the design of turbo-pumps, can be most per-
fectly and easily solved in the turbo-compressor. The first group
of a compressor, or the whole of any blower which is of the piston-
balance type just described, needs, we might say, no stuffing-box
at all, as both free ends of the machine are under atmospheric pres-
sure, and if air should be sucked in around the shafts, no harm would
be done, as air is just what the blower requires; only in cases where
the blower has to deal with very poisonous or valuable gases would
it be necessary to provide a special packing. In all the following
casings (that is, in casings containing the high-pressure groups),
stuffing-boxes may be wholly dispensed with by using a fully capped
bearing, as in Fig. 92. The probability of water mixing with the
oil would, of course, forbid this design with any good centrifugal
pump or water turbine.”

124 AIR COMPRESSION AND TRANSMISSION

“Figure ror shows test curves taken during the practical operation
of one of the latest turbine blowers supplied by C. H. Jaeger & Co.
to the Grillo Zinc Works, Ltd., Hamborn on the Rhine. The curves
show the relation between pressure, power absorbed by the blower
spindle, efficiency, and duty at constant running speed. The pressure,
horse-power, and efficiency, are ordinates over the respective duties
as abscisse, in like manner to that followed in making these curves
for turbine pumps. It is easily seen that over a fairly wide range
the pressure and efficiency remain nearly constant, and the regulation
of the capacity can therefore be made by simply throttling down
the surplus quantity. The speed of revolution need not be altered,
and alternating-current motors oran exhaust-steam turbine may
readily be used for driving these blowers in their simplest form.

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Fic. to1.—Test curves, Jaeger’s turbo-blower.

‘Coupling Compressors.—In some cases, as, for instance, in
blast furnaces, it may be necessary to generate an extra high pres-
sure for short periods. Such necessities may arise, for example, if
the resistance of the air nozzles or of the column of melting ore is
increased. The centrifugal blower running at constant speed would
not be able to drive the air through the mains at a pressure much
higher than normal. If, therefore, no means for regulating the run-
ning speed are available, some airangement must be supplied such as
that furnished by Sautter, Harle and Cie, to the iron works at
Chasse. Fig. 102 shows how coupling either in parallel or in series
is combined with perfect balancing of the impellers.

‘Although coupling the stages alternately either in parallel or in
series permits us either to deliver a la1ge quantity at normal pressure
or a reduced quantity at double that pressure, this mode of regula-

TURBO-COMPRESSORS 125

tion is unsuited to a great many cases. For instance, in the majority
of chemical processes, the delivery of constant volumes of gas is
necessary for the economical working of the process and the uniform
quality of the product. One of the best known of these processes
is the melting of pig iron in the cupola. The problem of delivering

Fic. 102.—Arrangements for coupling turbo blowers.

the constant volume against varying head can be solved only by
using varying speeds, but under this condition it can be solved by
the turbo-blower with almost unexcelled exactness. The means by
which this is effected is an apparatus which has been called by its
inventor, Prof. Rateau, the ‘Multiplicator.’

“Rateau Multiplicator—tFig. 103 shows a cross-section of the
Multiplicator as applied in the well-known turbo-blower plant at

‘Fic. 103.—Rateau multiplicator.

Rothe Erde. The principle is the same as that of the Venturi water
meter; that is, by tapering a pipe the velocity with which the gas
passes through it is increased as the diameter is narrowed. As
no addition or subtraction of energy is made during the passage of
gas through the pipe, any increase of velocity must be accompanied
by a decrease of static pressure. Ifthe air be tapped from the narrow-
est and widest points of the tapered pipe, and the tapping pipes be
led to opposite sides of a movable piston, it is clear that the difference
in static pressure on the two sides of the piston will either move that
piston or exert a certain effort on the piston-rod.

126 AIR COMPRESSION AND TRANSMISSION

“The greater the reduction of cross-section in the tapered pipe
is made, the greater becomes the effect exerted through the piston-
rod, and the greater also becomes the variation in that effort
caused by any increase or reduction of velocity in the main; that is,
of the quantity passing through this main per second. ‘Therefore,
to obtain a very sensitive piece of apparatus powerful enough to
move a governor gearing when the variation in the delivery is ©
but 1 or 2 per cent., it would be necessary to make such a great
difference of cross-sectional area that the resistance of the mains
would be considerably increased by the throttling effect of the taper.

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Fic. 104.—Piston controlled by Fic. tos.—Connection between piston
multiplicator. and regulators.

“Here the ingenuity of Prof. Rateau’s method appears. He
tapers the main but very slightly and inserts at the narrowest end
a system of pipes as shown in Fig. 103. At the point a the static
pressure is already reduced somewhat below that in the normal
pressure main. At the point 6 the pressuie of the small quantity
tapped off at a is again decreased. Finally, the static pressure at
c is in turn much lower than at 6. The effect of this arrangement
is so great that with a velocity of 60 ft. per second in the mains,
the difference of static pressure between a and c was about 6 in. of
mercury, while the loss occasioned by the whole installation was at
the same moment not mote than 3/8 in. of water column. By put-
ting the two ends of the cylinder A, Fig. 104, in connection with the
narrowest and widest points of the tapered piping system, a consid-

TURBO-COMPRESSORS | 127

erable force can be exerted on the spring B. It can easily be cal-
culated that a difference of 2 1/2 lb. in total pressure is generated
by a variation of about 1 per cent. in the velocity of the air current
passing through the mains; that is, when the quantity delivered
by the blower varies by about 1 per cent., a force of about 2 1/2
Ib. becomes available for moving the regulator.

“Figure 105 gives an idea of the manner in which the gearing was
arranged in a special case. The speed regulator and regulating
throttle valve of a Parsons steam turbine were influenced simulta-
neously. In like manner the regulating lever of any driving electric
motor can be moved in exact proportion to the momentum of the
air piston, as shown in Fig. 103.

‘It is very interesting to see how this achievement enabled turbo-
blowers of the Rateau system to create an entirely new field for them-
selves. One of the licensees of Prof. Rateau, the machine-manu-
facturing establishment of Kuehnle, Kopp & Kauschat Frankenthal,
delivered some turbo-blowers for the Anilin and Soda Factory at
Baden for the purpose of blowing air through the electric arc in the
newly invented process of obtaining nitric acid directly from the at-
mosphere. These turbo-blowels were to replace reciprocating blow-
ers, which had caused great trouble and expense. It was necessary
to connect them with a very large air-tank in order to produce reason-
able steadiness of the air current, and when inspection of the recipro-
cating blowers was necessary, it was impossible except by the skill
of very experienced mechanics, and even then only with the greatest
risk, to take one of the blowers out of service and at the same time
start ancther without interfering with the continuous current of
air. The turbine-blower not only gave an absolutely continuous
air current, but proved so safe in operation that no change of blowers
was needed during the entire process, which generally lasts uninter-
ruptedly for several months.

“Mixing Blower.—This extraordinary success of the turbo-
blower impelled the Badische Anilin and Soda Fabiik to order
fiom Kuehnle, Kopp & Kausch another kind of turbo-machine—
that is, a mixing blower, which is shown in Fig. 106. Two different
gases ale drawn by different sets of impellers, keyed on the same shaft,
and are delivered to two different delivery pipes. This alrangement
has the advantage that owing to the compulsory equality of speed
of both impeller groups, the relation between the quantities of the
gases continuously delivered is absolutely the same (that is, it is
proportionate to the cross-sections of the impellers) providing the
resistance remains equal in both delivery pipes. As this last con-
dition cannot be kept uniform during the chemical process, auxili-
ary throttle valves are inserted in the delivery pipes and worked
by two Multiplicators. These latter, after careful adjustment,

128 AIR COMPRESSION AND TRANSMISSION

insure the maintenance of absolutely constant mixing rates between
the two deliveries under all conditions. The makers were required
to guarantee that the mixing ratio should be kept constant within
a margin of 1 per cent., and their machines wete so perfectly designed

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Fic. 107.—Rateau turbo-compressor, 140,000 cu. ft. per hour.

that the Badische Anilin and Soda Fabrik at once began to develop
new processes for the electric synthesis of gases, which were made

possible only by the new mixing turbo-blower.”
The cross-section of a Rateau Turbo-compressor of 140,000 cu.

ft. per hour running at 4,600 r.p.m. is shown as Fig. 107.

CHAPTER XII
HYDRAULIC COMPRESSION OF AIR

The method of compressing air by means of falling water, without
the use of any other moving part whatever, forms one of the most
interesting topics in the subject of air compression.

The large installations in northern Michigan, together with the
large compressors of the same type in British Columbia, Quebec
and Connecticut, give some idea of the extent to which this very
simple method of utilizing the energy of falling water is being ap-
plied. All of these installations have been completed within very

Fic. 108.—The trompe. Fic. tog.—Frizell’s hydraulic compressor.

SNe

recent years and their success gives promise of many more such
plants being planned.

Trompe.—One of the oldest forms of compressing air is by means
of a trompe or water bellows, a device of historic interest, in which
water was lead from a higher to a lower level through a pipe or bam-
boo pole with openings in the side through which air entered and
mingled with the descending water and was later trapped from it,
as shown in Fig. 108, for use in forges.

A great many impovements have been made on this aan ap-
paratus and quite distinct types developed from it.

Frizell’s Compressor.—One of these is shown in Fig. 109, the
invention of J. P. Frizell of Boston, Massachusetts. This device
utilizes an inverted syphon having a horizontal passage C between

9 129

130 AIR COMPRESSION AND TRANSMISSION

the two legs, Band. A stream of water is led into the upper end
of the longer leg B and at the top of the horizontal passage C’ be-
tween the two legs of the syphon, an enlarged chamber, D, is con-
structed in which the air separates from the water. The water
freed from the air passes up the shorter leg, F, of the syphon to the
tail race. The pressure of air accumulating in the chamber is
determined by the height of water in the shorter leg.

This application of carrying upward the water after the air is
separated from it seems to have been original with Mr. Frizell, and
in this respect his device differs from the old trompe.

Mr. Frizell made two working models of this type of apparatus.
In the first one, the legs of the syphon were 3 in. in diameter,
the head of water being 25 in. and an efficiency of 26 1/2 per cent.
was obtained. A larger apparatus was then constructed at the
Falls of St. Anthony on the Mississippi River a few miles above
St. Paul; the longer leg of the syphon in this plant was 15 X30 in.
and the shorter leg of the syphon 24X48 in. in section; the height
of water above the air chamber was 29 ft. The head in feet varied
from 0.98 to 5.02; the first head being just sufficient to cause a flow
through the pipes. With the working head changed from 2.54 ft.
to 5.02 ft., the efficiency varied from 40.4 per-cent. to 50.7 per cent.,
the quantity of water in these cases varying from 5.92 to 11.89 cu.
ft. per second.

Mr. Frizell estimates from the experiments he has made that with
a shaft ro ft. in diameter, a depth of 120 ft. and a fall of 15 ft., the
efficiency would be 76 per cent.; and with a head of 30 ft. and a fall
of 230 ft. the efficiency would be 81 per cent.

Mr. Frizell’s first experiments involved a large outlay in cost of
plant and were not entirely satisfactory; but where there is a mod-
erate water fall and plenty of water, this is no doubt a very simple
method of compressing air.

This system is applicable to either high or low falls and although
no installations of this type of air compressors were made until a
number of years after Mr. Frizell’s patents were obtained, the fact
that he is the pioneer in this line entitles him to a great deal of
credit. |

The following explanation of this system is taken from The
Railway and Engineering Review, Sept. 17, 1898.

“The general principles underlying this method of compression is
familiar to most in one form or another. For instance, it is well
known how readily water breaks into foam, which is due to its being

HYDRAULIC COMPRESSION OF AIR 131

impregnated with airin minute bubbles. Since bubbles rise in water
at a velocity depending on the size of the bubble, it is obvious that
air drawn into a current of water moving downward with a velocity
in excess of that at which the bubbles rise will be carried down and
subjected to a pressure corresponding to the depth attained, and
moreover the compression will take place isothermally, a process
which is not accomplished by any method of piston compression.
If the direction of the water be then altered to a horizontal one, the
air will rise in a few seconds to the top of the passage and accumulate
in a suitable chamber under the desired pressure. The length of
the horizontal tunnel will be controlled by the necessity of placing
the entrance to the air chamber far enough from the descending
branch to admit of the complete escape of the air bubbles.

‘‘A method of introducing air into the descending column of water
is to surround the shaft with a bulkhead of masonry, over which

Fic. t10.—Syphon bulkhead.

the water is led in a covered channel, the bottom of which rises a
little above the highest level of the water. This forms a syphon
as shown in Fig. r1o.

‘“‘At the point A the pressure within the syphon is less than that
of the external air, and the latter will flow in through any opening.
This is evident, because the flow of water depends upon the syphon
principle. This space A extends around the masonry bulkhead and
is in communication with the atmosphere. It is also connected with
a pump for the purpose of removing any water that may collect in it,
the amount of air being regulated by opening or closing holes in
chamber A.”

Baloche and Krahnass Compressor.—Another device, shown in
Fig. rrr, differs somewhat from that of Mr. Frizell. It was invented
by A. Baloche and A. Krahnass in 1885 and consisted of a syphon,
B, carrying water from an upper to a lower reservoir, the lower end
of the syphon being projected through an inverted vessel, R, placed
nearly at the bottom of the second reservoir. Just beyond the bend

132 AIR COMPRESSION AND TRANSMISSION

in the syphon and in line with the axis of its longer leg, an air pipe,
T, projected into the descending leg of the syphon. This entrained
the air with the descending column and carried it down into the in-
verted chamber, R, from which the air escaped at the top, while the
water passed out from the bottom into the lower reservoir. This
apparatus produced pressure on the air in the top of the inverted
chamber due to the height of the water column upon it.

Arthur Compressor.—Another device, shown in Fig. 112, patented
by Thomas Arthur in 1888, differs from the last in having a stream

Fic. 111.—Baloche and Krah- Fic. 112.—Arthur’s hydraulic com-
nass’s hydraulic compressor. pressor.

of water led directly into the top of the vertical pipe A. Inserted
into the mouth of this pipe is a double cylindrical cone, C, forming
an annular air passage between it and the walls of the pipe A. Owing
to the increase in the velocity of the water passing through the nar-
row throat of the double cone, air is inhaled through the pipe D,
through the annular space mentioned and through perforations in
the lower cone and is entrained with the falling water.

Through the down-flow pipe A rises a vertical delivery pipe, Z,
for the compressed air, having its lower end, H, enlarged and open at
the bottom. Projecting upward into this enlarged air-delivery pipe
is a water escape pipe, /, through which the water passes after
parting with the air. The escape pipe is in the form of an inverted
syphon and maintains on the air in the delivery pipe Z a pressure due
to the elevation of the water at the discharge point above the air line
in the large end of the delivery pipe.

HYDRAULIC COMPRESSION OF AIR 133

Taylor Compressor.—The hydraulic compressor system of Mr.
Taylor is shown by Fig. 113. The large recent installations referred
to are principally based upon his patents.

With Taylor’s system a series of small air pipes placed vertically
in the upper end of the falling column of water introduce the air into
the water. The compressed air and water are separated at the
bottom of the shaft.

Mr. Taylor seems to have been the first to introduce the plan
of dividing the air inlets into a great number of small openings
evenly distributed over the area of the water inlet.

Fic. 113.—Taylor’s hydraulic compressor.

In the figure shown, these air tubes are represented at C, all
terminating at the conical entrance B to the down-flow pipe H. The
water supply is furnished to this down-flow pipe through a flume D.
As the water falls it draws air through the small tubes, carrying it
down to the separating tank G, where it is liberated at a pressure
depending on the weight of the water in the vertical pipe Z.

The compressed air is then conducted through the pipe K to the
place to be used. The distance from M to the tail race L represents
the height or fall of water that is available.

Taylor at first seems to have attempted to utilize centrifugal
action in causing the separation of the air and water in the large
chamber at the bottom, but afterward abandoned this scheme and
‘used instead forms of deflector plates in combination with a gradu-
ally enlarging section of the lower end of the down-flow column in

134 AIR COMPRESSION AND TRANSMISSION

order to decrease the velocity of the air and water and cause the
water to part more readily from the air.

The position of the hopper or frame carrying the air inlet tubes
regulates the amount of water that is admitted to the vertical pipe.
The quantity of air regulates itself and is neither more nor less than
the given quantity of water can carry. If the descending column is
so loaded with air that it does not preponderate sufficiently over the
ascending column, the water in the former will rise, the commotion
will diminish and less water will enter. In the contrary case the
water falls, commotion increases and more air is taken in. ;

Taylor Compressor at Magog, Quebec.—The first one of these com-
pressors on the Taylor principle was installed at Magog, Quebec, to
furnish power for the print works of the Dominion Cotton Mills
Company. The head of water is 22 ft., the down-flow pipe is 44 in.
in diameter and extends downward through a vertical shaft ro ft.
in section and 128 ft. deep. At the bottom of the shaft the com-
pressor pipe enters a large tank 17 ft. in diameter and tro ft. high,
which is known as the air chamber and separator.

A series of very careful experiments have been conducted at the
Magog plant by Professor Kennedy and others; and it has been
demonstrated that with a head of 19 1/2 ft. of water using 4,292 cu.
ft. of water per minute, the equivalent of 1,148 cu. ft. of free air per
minute was recovered at a pressure of 53.3 lb. showing that of a
gross horse-power of 158.1, 117.7 h.p. of effective work was used in
compressing air, giving an efficiency of 71 per cent. which is very
satisfactory.

_This compressed air was then used in an old Corliss engine, with-
out changing the valve gear in any way from what it was when
adjusted for steam, and 81 h.p. was recovered, showing a total of
work recovered from the falling water of 51.2 per cent. When the
compressed air was heated to 276° before being used in the engine,
11m h.p. was recovered. The heating required 115 lb. of coke
per hour, equal to about 23 h.p. The efficiency, therefore, including
the falling water and the fuel consumed, was 61 1/2 per cent. It
has been calculated from other experiments that if the compressed
air had been heated to 300° the total efficiency secured would have
been 87 1/2 per cent.

When it is considered that a good water turbine will give an eff-
ciency of 85 per cent. and that part of the power developed in ,
the turbine will be lost through transmission before the power is
available, it is evident that this system is a very efficient method

HVDRAULIC COMPRESSION OF AIR 135

of generating and transmitting power. For if the efficiency of the
turbine is 85 per cent. and that of the system that is used for con-
verting the power in the turbine into a more.available form 80 per
cent., the total efficiency of the system will be 0.80X0.85 or 68
per cent.
This shows the immense importance of this device. Its field of
usefulness is certainly a large one.

Taylor Compressor at Ainsworth, B. C.—Figure 114 illustrates
a sketch of the upper part of the Taylor Hydraulic compressing
plant at Ainsworth, B. C., which is

< A é 6 q ——oF
quite unique in that it did not re- ) AT
quire the sinking of a very deep B i B
shaft. Theapparatusis constructed aAadianng

against the vertical wall of the
canyon in the rugged mountain dis-
trict in which it was built. The
plant is located on Coffee Creek to
the south of Ainsworth and about
2 1/2 miles from the principal mines
to which it supplies compressed air.
‘The creek has a flow varying from
2,500 cu. ft. per minute to several
thousand, and the flume used is
stave barrel construction, round
steel bands being bolted around it
every at... Phe wtilume? is 17250711,
in length, 5 ft. in diameter in the
clear, the available head at the
compressor being 107 1/2 ft. The
water at the compressor tower is received in a wooden tank 12 ft.
in diameter and 20 ft. in height. A down-flow pipe passes from
the water level through the bottom of this tank down perpendic-
ularly and at the creek level a shaft extends to a depth of 210 ft.,
making a total vertical height to the shaft of over 300 ft.

This down-flow pipe, which is 2 ft. 9 in. in diameter outside, is
also of stave construction throughout, the bands being placed
from 6 in. to 3 ft. apart, depending on the pressure to which a
particular section is subjected.

This terminates in a great bell-shaped chamber at the bottom
of the shaft 17 ft. in diameter and 17 ft. high, the bottom of this
chamber being about 2 ft. above the bottom of the shaft, thus allow-

Fic. 114.—Taylor’s compressor at
Ainsworth, B. C.

136 AIR COMPRESSION AND TRANSMISSION

ing the water to pass out and up the shaft to the tail race. A deep
circular groove was dug in the bottom of the shaft to aid in separating
the air from the water.

As the distance from the water level of the air chamber to the
tail race is about 200 ft., the pressure on the air is about 87 lb.
per square inch.

The air is conducted from-the compressor through a g-in. pipe
which supplies compressed air through several branches to over
I5 mining properties. The total length of pipe is over 2 miles.
A pipe reaches from the surface of the creek level—that is, the tail
race—to the dividing line between the air and water of the large
chamber at the bottom; so that if more air is being compressed
than is being used, the water line in this chamber will be lowered
and the surplus air escape, while if the pressure of air falls, the escape
pipe will be closed.

The actual effective head of water in the apparatus is 107 1/2 ft.
and if a turbine had been used, with an efficiency of 75 per cent.
the available horse-power generated would amount to 620.

This installation has cost in the neighborhood of $60,000, including
incorporation, water-power, development and pipe line. Of this
investment, $20,000 will cover the pipe-line cost, $10,000 the water-
power improvements, and $30,000 the compressor cost. This
last item was unusually high because. of the extremely hard founda-
tion through which the shaft was sunk.

Taylor Compressor at Victoria Mine, Mich.—In 1906 a large
plant of this type was installed at Victoria Copper Mine near Rock-
land, Ontonagon County, Michigan, which consisted of three com-
pressing units with a total capacity of from 34,000 to 36,000 cu. ft.
of free air per minute. A series of tests made on a single intake
head by Prof. F. W. Sperr, gave the following results:

TABLE X.—AIR MEASUREMENTS

: , Absolute pressures
Velocity | Cubic feet |_
Square feet

feet per per

area 3 Free air, | Compressed
second minute :
pounds | air, pounds

Horse-power

4 44.09 10,580 14 128 1,430
49.74 11,930 14 128 T,627
4 38.50 9,238 14 128 1,248

HYDRAULIC COMPRESSION OF AIR 137
TABLE XI.—WATER MEASUREMENTS
Mune Velocity | Cubic feet Bibcen cr:
feet per per Head, feet | Horse-power
area : per cent.
second minute
71.75 3-933 13,057 79.5 1,741 82.17
67.03 3.684 14,820 70.0 1,961 82.27
7210 2.936 12,710 70.6 1,700 7350

Phenomena of Hydraulic Air Compression.—There are several
phenomena in connection with this method of compressing air that
at first thought seem paradoxial.

In compressing air by hydraulic means, the air becomes drier
during the compression, but no matter what may be its initial
condition as to humidity at the end of compression it will, in all
probability, be saturated with moisture.

Air almost always contains moisture. Its capacity for moisture
is determined by the combined conditions of pressure and tempera-
ture to which it is at the time subjected.

Changes, either of pressure or of temperature, immediately change
the capacity of air for water, and if the free air is saturated with
moisture its capacity for water will be reduced whenever the pres-
sure is increased or the temperature decreased, and in consequence
water will be precipitated.

When air is compressed by hydraulic means, isothermal com-
pression is secured and, generally speaking, at uniform temperature
a given volume of air implies a capacity for a certain weight of
water whether the air is at a pressure of one or one hundred at-
mospheres, but if the air is compressed through a range from 1 to
too atmospheres, its volume will be reduced, if the compression is
isothermal, to 1/100 the original volume, and in consequence
99/toc of the weight of moisture it originally held will be precipi-
tated. The air is still saturated, but the total weight of water held
in suspension has been reduced. That is, this method of com-
pression has reduced the weight of moisture present in the air and
hence dried it, but at the end of compression the air is saturated
with moisture.

Another interesting phenomenon in connection with this type of
compressor has recently been brought to light. It has been found
that air compressed by this method contains less oxygen than free

138 AIR COMPRESSION AND TRANSMISSION

air of the atmosphere and in consequence its use in mines is not as
beneficial as air from other types of compressors.

It will be observed that, with this construction, the material used
for the down-flow pipe need only be of sufficient strength to carry
the weight of water and pressure generated in the working head of
the water-power, as once it reaches the tail race level the internal
pressure is gradually neutralized from that point down by the pres-
sure in the return water surrounding the down-flow pipe; so that any
pressure almost may be reached without increasing the strength of
the down-flow pipe. The material for the down-flow pipe may be
iron or wood hooped with iron, and the shaft may be constructed
of inexpensive timber as it is preserved by being constantly in the
water.

By this method, low falls, otherwise useless, are made available
and the same pressures can be obtained as from high falls, the horse-
power being determined by the diameter of the down-flow pipe, and
the height and volume of water in fall, while the pressure depends
solely upon the depth of the well or shaft; therefore, any desired
pressure can be obtained.

Briefly stated, the air is compressed by the direct pressure
of falling water without the aid of any moving mechanism and prac-
tically without expense for maintenance or attendance after instal-
lation.

By this system any fall of water varying in working head may be
utilized, any pressure required can be produced and uniformly
maintained up to the capacity of the water-power, delivering the
compressed air at the temperature of the water.

This drying of the air and the fact that practically isothermal
compression is secured, form the great advantages of this system
of air compression. ‘The initial cost need not be excessive, and as
the cost of attendance is slight, for certain purposes the method is
ideal. Its field of operation is quite broad, as a high fall of water
is not essential, for any desired pressure can be obtained from any
fall, the capacity being determined by the power available in the
water fall.

Losses of Hydraulic Compression.—The losses inherent in this
method of compression are: (1) The head expended in impregnating
the water with air. This usually amounts to about 1 ft.

(2) A loss which may be called the slip due to the velocity with
which the bubbles tend to rise. It is obvious that the rise of bubbles
during the decent of the water is a lost motion which deducts from

HYDRAULIC COMPRESSION OF AIR 139

the efficiency of the system and in addition there is a head con-
sumed in friction.

(3) A loss due to the increasing solution of the air in the water
with the increasing pressure as the water and air descend.

This air does not separate from the water in the lower chamber but
is eliminated in the ascending shaft in the same order that it is
dissolved in the descending shaft. The escaping air in the ascending
shaft aids the movement of the water and this partly balances the
loss in the descending column.

There are on the market to-day small hydraulic air compressors
for furnishing compressed air in small quantities for dental and other
purposes. They can be operated by water pressure from any water-
works supply and on this account are particularly adapted for such
purposes.

CHAPTER XIII
EFFECT OF ALTITUDE AND COMPRESSOR TESTS!

As the density of the atmosphere decreases with the altitude, a
compressor located at a high altitude will take in a smaller weight
of air at each stroke, that is, if the compressor is located at a high
altitude, the air is taken in at a lower pressure and in consequence
the early part of the compression stroke is occupied in compressing
the air from this lower density up to a standard atmospheric pressure
at the sea-level. The reduction of pressure at the inlet would, of
course, affect the power expended in compressing the air, but the
decrease in power required does not vary in the same ratio as the
decrease in capacity. For this reason compressors to be used at
high altitudes should have the steam and air cylinders properly .
proportioned to meet the varying conditions at different levels.

Effect of Altitude on Capacity.—This matter is of special im-
portance in connection with mining operations, because of the large
number of mines situated in elevated mountain regions. The rated
capacities of compressors, in cubic feet of air, as given in the makers’
catalogues, are for work at normal atmospheric pressure, and due
allowance must be made for decreased output at elevations above
sea-level. This reduction in output, which is usually also tabulated
in handbooks and catalogues, should receive due consideration in
order to avoid serious errors. For example, the volume of
compressed air delivered at 60-lb. pressure, at 10,000 ft. elevation is
only 72.7 per cent. of the volume delivered at the same pressure by
the same compressor at sea-level. In other words, a compressor
which at sea-level will supply power for 1o rock-drills, will at an
elevation of 10,000 ft. furnish air for only 7 drills.

Effect of Altitude on Power.—The foregoing statement relates
only to the volumetric capacity of the compressor. It must be re-
membered that the heat of compression increases with the ratio of the
final absolute pressure to the initial absolute pressure. As this ratio
increases with the altitude, more heat will be generated by compres-
sion to a given pressure at high altitudes than at sea-level. This
additional heat temporarily increases the pressure of the air in the
cylinder while under compression, and more power is therefore re-

1Peele, Compressed Air Plant.
140

HHPECIOCOFr ALTITUDE AND. COMPRESSOR TESTS ~ 141

quired to compress and deliver a given quantity of air. The cor-
responding loss of work, due to the subsequent cooling of the air in
receiver and piping, also increases with the altitude.

Relation between Altitude and Volume.—Contrary to a common
impression, the volume of air delivered by a given compressor does
not bear a constant ratio to the barometric pressure, but at different
altitudes this volume decreases slower than the barometric pressure.
This relation may be shown as follows: Two ideal indicator cards are
represented in Fig. 115, one of a compressor working at sea-level with

=
_
-—-

=<-
“=
oa”

4
Ce

P.
P.

—----——

eth eet be as ee

{

<
4 re

Fic. 115.—Effect of altitude.

an initial pressure Py, the other at an altitude with a lower initial
pressure P2. The initial volume V and the final gage pressure P
are the same for both compressors, P3 and P, being the respective
final absolute pressures. V1 and Ve are the final volumes, corre-
sponding to the dotted isothermal curves, these volumes being
taken as the basis because they are those to which the compressed
air will eventually shrink on losing the heat of compression. From
the theory of air compression,

VAeP
VP\—= Vik; or Wire, (1)
Vee:
and VPo=VoPa, or V7, (2)

But since P3=P,+P, and Ps=P2+P, equations (1) and (2) may
be written:

ary Cae if je (3)

and is ena. i (4)

142 AIR COMPRESSION AND TRANSMISSION

Dividing equation (3) by equation (4)

li
i+
Vogue 1 F Bae i : ge
ae po Vie Vee ioe pease (5)
ari

This gives an expression for the ratio between pressure and volume
at sea-level and for any altitude above sea-level, of which the corre-
sponding barometric pressure isP2. Thus, let Pps=10lb.,P=golb.,
and V;=0.1404 cu. ft. By substituting these quantities in equa-
tion (5), V2 is found to be 0.0999, or approximately o.1 cu. ft.

In Table XII, columns 4 and 5, are given the relative volumetric ©
outputs, at gage pressures of 70 and go lb. of a compressor working
at different altitudes, the figures being percentages of the normal
output at sea-level. These percentages have been derived by
Mr. F. A. Halsey from equation (5), a constant loss of initial pressure
of c.75 lb. being assumed to allow for the resistance presented by
the inlet valves, to which reference has been made in another
chapter. That is, for practical purposes the sea-level atmospheric
pressure is taken as 14, instead of 14.7 lb. The other columns
show the mean effective pressures and indicated horse-powers,
corresponding to different altitudes, up to 15,000 ft., which will be
found convenient for reference. It should be noted from the figures
in columns 4 and 5, which are for the ordinary range of pressure
employed in mining, that, though there is a difference of 20 lb.
between the two gage pressures, yet the outputs at different altitudes
vary only by a few thousandths and may often be neglected.!
Wide differences, however, occur in the columns of mean effective
pressures and horse-powers.

Owing to the increase of piston displacement per indicated
horse-power, as shown in columns 8 and 9 of the table, some
builders make the air cylinders of compressors for mountain work
of larger diameter for the same size of steam cylinder than those
for sea-level service. As against the losses of the air end of the
compressor at high altitudes, there is some gain in mean effective
pressure of the steam cylinders, because the exhaust takes place
against lower atmospheric pressure. The same is true in part
of the air exhaust of machines using the compressed air. But the

1 Attention may be called to the fact that for this reason, in compressor-

builders’ catalogues, no account is taken of the gage pressures in tables of
compressor capacities at altitudes.

EFFECT OF ALTITUDE AND COMPRESSOR TESTS 148

resultant of these gains is small and cannot be given much weight
in offsetting the losses.

TABLE XII
. ‘ f ,
Barometric poke eee Cubic feet
Pressure pistons compressed
.7 Relative out-| M.E.P. for placement : ene
: ce. oi air per indicated
put for gage gage pres- per indicated
horse-power
Altitude, pressure sure horse-power foreraze
feet Pounds for gage
Taches ve préséitre pressure
mer-
square
cury ach
: 7.OnlDeiLOOm Delle Onl Den OO; bs 1770: 1bn00) Lbs 1.70 1b. | 90 1b.
I 2 3 4 5 6 i. 8 9 Io II
fo) 30.00 17 5S OOO LE NOOOU Sale So Neal OOS 5,00 et. LAAN On sor
I,000 28.88 Te? ONO TOO 7aLOR OOO IES 2e Ollme 37> Olle yn O3 ale O,. 00m i. k23 nl cOm7 a7
2,000 27.80 TA OVMMOMOS Su LOTOS ses 2a. MesOnOM) 7S Os2On) ele Toss OL773
3,000 26.76 U3 eLOe | OFOOA LO ROOOH EST a Sle 3053) I) Feo 71, Oc ZT Hl T.O84) 190.7750
4,000) 25.76 T22O7MOnS (nO eCOON, GEeO leas eOmle7 43004) Oo4 30 2. O05 a) 0.740
5,000 24.79 P22 OMNLOMGAG HOSS Ones On5 ues 5 Omiya 5 Loos 55 aie BOAO) Pron 733
6,000 23.86 Tey Ga ROMO Css LOMmOO Olle OOM s Aisin OS) al. Or. Oil ate O26) \kOe 720
7,000 22.97 Dees On One Sun Ole OON ee Oud aime 37 lee 1S Onl O47 ON ie ON Le sOn 70'S
8,000 PAP) Su TOMS 7 OMT SOULOM 7S Lm2o Oma sa ale 704 O02 OL OOd i) On005
9,000 PA PAO, LOMAOM MOR 72 TMPOM 7232 Seow S2n Silo 5OOMs 7: OONl O1O76 ale 0.083
I0,000 20.49 TORO 7a OMT OSL OnOOO! E276 5. lest Sil (Oo 224 071.204) OLOSOMI T0070
II,000 Lon 72 On OmmFOF OSOML ORO Mae 2 7 Amie Gl so 3O ls 7134 aOLOAZE I O2058
I2,000 18.98 Onset LO sO5010 50473 |620,.0)1 30.0.9) 8554 107240, 1) 0.025 5) 0.046
13,000 Qua] SAOSMLOVOS2IOmO 232003 sO MOMS. 7k wn OAM OOO Ga 1041035
I4,000 D750 865 04008) 02.000)) 25q8 e204 oe S58) 7.800) On8Org 02624
15,000 16.93 SES 2 MOMS osNLOM S57 OlmMe Se sale COOL OOM: 7 (OO O.o 75 moe OLS

The relation between compressor output and barometric pressure
may be expressed simply in another way. Take the case of two
compressors of the same size, one operating under an atmospheric
pressure of, say, 14 Ib. and the othe at 10 lb. (corresponding approxi-
mately to an altitude of 10,000 ft.). If the first compressor is
producing 6 compressions, the final absolute pressure will be 146
= 84 lb. or about 70 lb. gage pressure. To produce the same gage
pressure, the other compressor must work to an absolute pressure
of 7o+10=80 lb., the number of compressions corresponding to

ep ibe 2
which is ert From each cubic foot of free air the compressor

will produce 1/6 of a cu. ft. of compressed air, and the second
complessor, 1/8 cu. ft. Hence, the ratio of the respective outputs
of the two compressors will be 1/8+1/6=3/4 or 0.750. As com-

144 AIR COMPRESSION AND TRANSMISSION

pared with this, the ratio of the respective barometric pressures

5 10
iS ade

COMPRESSOR TESTS

To indicate the observations required to secure the data for
the complete test of a compressor, together with the deductions
from the observed data, the following record of the test of a com-
pound, two-stage Nordberg compressor, at the mines of the Tennes-
see Copper Co., will be found useful.’ It will be noted that items
28, 29 and 32 to 35 are necessary in this case, because the boiler
plant supplied steam for the hoisting engine and an independent
condenser, as well as for the compressor. Though the hoist was
not running, steam was passing continuously to the jackets of the
cylinders. The same conditions would often be met in other tests.
The boiler-feed water was taken from a wooden tank, and during
the run this water was supplied from two barrels on scales set
temporarily over the tank. The water of condensation from steam
jackets and reheater was drawn off continuously and also weighed.
The calorimeter tests were made with a Peabody throttling calori-
meter. Eight sets of indicator cards were taken during the 8-hour
test, at hourly intervals.

ITEMS OF COMPRESSOR TEST

Altitude, 1,800 feet

1. Date of test, February 16, 1902.

2. Dutation of.test. hours<,6) a 8
3. Diameter of high-pressure Sank (aindee
(steam jacketed); inches im wen ee eee 14
4. Diameter of low-pressure steam cylinder
(steam, jacketed), inches.) v)j.s.42 4am. 28
5. Diameter of low-pressure air cylinder, inches... 24 1/2
6. Diameter of high-pressure air cylinder, inches. 15 3/8
7estroke ol allpistons, Inehesa 4. ase eae ee 42
8, Diameteriol piston-rods<inchés.. eee 201/16
g. Revolutions of engine, average per minute.... go
10. Piston speed perminute. ect. .n) al eee 630
11. Steam-gage pressure, average, pounds........ 145.9
12. Temperature of steam in bees average
degrees Fahrenheit. . ee 364

1 Abstracted from an article by I Parke aera Mines and Minerals, May,
1905, P- 475-

EFFECT OF ALTITUDE AND COMPRESSOR TESTS

rat
14.
15.
16.
17.
THe)
IQ.
20.

22.

Steam pressure in Spee Seeitee average

POUDCSe eaten he ean 8
Vacuum in Pordencer™ average inate BA aes oer
Air pressure in intercooler, average pounds... . Dee
Air pressure in receiver, average pounds...... 70.
Temperature of air at intake, average degrees

Fahrenheit. . foes 65.
Temperature ie air sieving Noaenresetce eyline

der, averave degroesahrenneit, .,...40.. 201"
Temperature of air leaving intercooler, average

deprecsalreinrenhiel ie sie teh tee ies 78.
Temperature of air leaving high-pressure cylin-

gderrvaverave cevreesm abrenneth: j) vo... ss , 240.
Indicated horse-power in high-pressure steam

CYMNGCE AAVCLACG at an eae, situs ne. oh 140.
Indicated horse-power in low-pressure steam

CyUnder Aa Vera ge meme ieN pe teen, an abel. oath. 153.

. Indicated horse-power in both steam cylinders,
PVCEAC Owe an ore eae 203.
. Indicated horse. power in plows -pressure air ee
ITC CUM VEL APC ARP ete et C00 vay Ci ddacs ae a 743"
. Indicated horse-power in high-pressure air
CVILGe Lercay CLAP ern city a euels awe da nets eh a Tage
. Indicated horse-power in both air cylinders,
VOU Omget, pcre Men Sole a ede eS eue LARP LS 278.
. Feed-water weighed to boilers, pounds........ 43,343

. Re-heater and jacket water from compressor,

Were edt OUh Co Auer ee Peat ides Ade gee 4,081

. Average temperature of re-heater and jacket
water, degrees Fahrenheit....'...... 350.
. Total heat in 1 lb. of steam ae 200d ve F., bheat

LARUE ete Ea fers eR ly rele acct Sea at a cup 1,190
. Total heat in 1 lb. of water at 356.7° F., heat

units. ids 320.
3 amvalcat edie tis re- rieiter anh Ayaas

WaterepOUNdSr vate tear. Sai? Rae Le
. Water weighed tron Rosideiseen orict in Renee

engine acket, DOUNUS#. fe. ele ae 1,781.
. Steam used to run condenser, pounds......... 4,320.
. Total credits to feed-water, pounds.......... 7.228).

. Total feed-water charged to engine, pounds... 36,115.
. Moisture in steam shown by Peabody calor-

imeter, per cent.. MR att ee or on Coes hae eh en iy
. Credit for ribistaret in eecenn sounds ey crea & Aus
. Total steam charged to engine, pounds....... 35,642.

. Dry steam per hour charged to engine, pounds.. 4,455.
. Steam consumption per indicated horse-power
Dev MIOUL MPOUNGS an eal. saath al OM neckdye sf. 5,

66
63

O02

oOo
oOo
OO

30
OO
OO
0O

145

146 AIR COMPRESSION AND TRANSMISSION

42. Guaranteed steam consumption per indicated
horse-power per hour, at 92 revolutions per

Taintite pounds.) aa en. 14.00
43. Excess of steam SORE TRD TORN per cued

horse-power per hour over guarantee, pounds 1.19
44. Theoretical delivery of free air per minute at

oo revolutions; cubic teet.|. 2)... ke ee O59 oO
45. Slip of air (percentage)... sok gates els)
46. Actual slip of air per fnute. vcabic tenes As HAE A Gilet
47. Actual delivery of free air per minute, average

CUDICH ECL aay cre: etLs07 0077

48. Theoretical terse -power Prenrede teh compress

and deliver actual delivery of air at receiver

pressure by adiabatic compression. . as 306.53
49. Theoretical horse-power required to compress

and deliver actual delivery of air at receiver

pressure by isothermal compression. . ee 229.00
50. Actual horse-power shown by air rican

cards.. eee 278.81
51. Actual eee -power howe bis steam i eevee

Candseeeak teu 293515
52. Actual Ree -power teonoumed iy fockon of

Celie eee : 14.34
53. Efficiency reo betw eens steam Pend air Rasatie

ders, petscel tye ae ee OSes
54. Efficiency ratio henween tear cae air veoiee

ders guaranteed by builder, per cent. 87.00

55. Efficiency of steam, or ratio of steam indicated
horse-power to theoretical air indicated horse-
power, isothermal compression, per cent.... Yow f

One of the combined indicator cards, from which the averages in
items 21 to 26 were calculated, is shown in the upper part of Fig. 116.

A series of tests were made in 1909 by Richard L. Webb, consulting
engineer, of Buffalo, N. Y., on a large number of compressors in a
well-known Canadian mining district. In conducting these tests, Mr
Webb had access to plants which have been in operation for a year
or more under normal working conditions, and his results are of
value not only to users of air compressors, but also to the manufac-
turers. As a rule, the plants tested were in the care of competent
machinists and in good running order, so that the results obtained
may be taken as representing a fair average of current practice in
the United States and Canada. The results of a few of these tests
are given here to show the importance of determining the actual
efficiency of air compressors when working under the conditions pre-
vailing in most mines.

EFFECT OF ALTITUDE AND COMPRESSOR TESTS 147

Discharge Pressure

x
GA
ZO" C163 %

Scale 20

Fic. 116.—Combined cards from two-stage compressors. Upper cards from
Nordberg compressor. Lower cards from Ingersoll-Rand “Imperial Type 10”
electrically driven compressor. Air cylinders 23’ and 14” X20”.

ROU tONS Per NINE Ca ete eer MED eye eee re Bane. atr een) BOT
Histon specd tect per minute c.f... ee ole we. wes es O23 23
Discharge air pressure, pounds...................--.. 93
Tntercodher pressure, DOUNdSs, o atau ustincy ts ok te es 24
Volumetric efictency (from catd)og cata. iss... | 9563%
Teeter ol low-Dressitercy Gehan ie tet he es Marg E32
etl Peonich-pressure cylinder. 22 40 te. et) T2078
BERL Ci teeter ee ae eae ee SASS a aden ASA. 8
Free air delivered per minute, cubic feet (from card)... 1706
HiMClenCY: COMP areas WitOyAGIAD pie a: oo sues Sis ee << O72 0
Pecieney COMarER WIL ssOtUae we crus a aaa fares nue oo) OFM RVS

Mode of Conducting the Tests.—The following plan was employed
in each case. First, a boiler test was run for not less than two weeks,
the coal being carefully weighed, the boiler feed-water measured,

148 AIR COMPRESSION AND TRANSMISSION

and the total revolutions of the compressor recorded by a revolution
counter. From these data the cost per boiler horse-power and the
average speed of the compressor were determined. Second, the com-
pressor was operated at various speeds over its entire range. By
means of a meter installed in the steam-pipe near the throttle, the
total steam consumed, in pounds per hour, was measured. Indi-
cator cards were taken on all cylinders, together with temperatures
at the air inlet, intercooler, and discharge. To measure the actual
volume of air delivered, a meter was placed in the discharge pipe
outside of the receiver. A number of simultaneous readings on all
instruments were taken at each speed. From these were calculated
the total horse-power of the steam and air cylinders, the steam
consumption, and the total piston displacement per minute.

The air and steam meters were of the Dodge type, as modified
by the General Electric Company, and were operated by their ex-
pert sent for this purpose. The indicators were of the Roberts-
Thompson and the American-Thompson make, which are well
known and generally accepted as standard. Their springs were
calibrated by a standard gage.

Results of the Tests.—As was to be expected, the friction loss was
found to be only a small item in the total. The other losses, which
are frequently overlooked or disregarded, played a large part in
cutting down the efficiency. The capacity of air compressors is
usually rated according to the volume of the cylinders. On this
basis, the mechanical efficiency only is given. For example, if the
horse-power of the air cylinder is 100 h.p. and the horse-power of the
steam cylinder 110, the efficiency of the compressor is rated as g1
per cent. This rating disregards the losses due to adiabatic com-
pression, heating of the cylinder and friction of the inlet and delivery
valves. The tests show the friction loss of the engine itself to be
usually not less than to per cent., and often considerably larger.
Losses from the other causes mentioned were found to range from
20 Per cents up:

As Mr. Webb is not at liberty to disclose the identity of the partic-
ular plants at which the tests were made, each test has been desig-
nated by a number.

Test of Plant Number One.—This consists of three 125 h.p.
return tubular boilers (one being held in reserve), supplying steam
for a cross-compound condensing air compressor of standard make.
The steam cylinders have Meyer valve gear and are 16 in. and 28 in.
diameter by 24 in. stroke. The two-stage air cylinders are 28 in.

EFFECT OF ALTITUDE AND COMPRESSOR TESTS 149

and 18 in. by 24 in. From a two weeks’ run the following results
were obtained:

Fas esas al ie ee el

‘
= RIDES aaa
Se ES
| aR Ne .
HEBRUSY
Eide tases s
ramen '
(aie iasltviesl ave i
Rene ase :
honwoeZes :
Cet Aone
ZZ caie win
Wa :
0 10 20 30 40 100 110

R. p.M.

Fic. 117.—Test results of compressor plant No. tr.

Potaltcoalspurneds pounds s255as. easels en «2 2 204,300
Total iced-water scubiCyeGtsm. fort en err ss oy. 37,459
Total feed-water, pounds.............. . 2,335,508
Average temperature of feed-water, degrees

Hahrenhiel tage seman on wg ers eee Mach 131
Average evaporation per pound coal con-

SUTMCC MDOUMCS ae tue Sor s Thee, weoah te mnce. ak S72
Average revolutions per minute............... 03 aE
Indicated horse-power of steam end, corre-

sponding; to 6321 t.p.1, af 2 5 yom sean - 161
Corresponding indicated horse-power of air end. 123
Average steam pressure, pounds.............. £15
AN OLAGesVaCUUIh,, DOUDCS, 028, 74: ajacteere ss sock TO.5
Averageyail pressure, pounds, 6:5 hte ness 96
Average temperature of outside air, degrees

PATI TORLELE Roe Patan RRO cases ocala: ohe/4 24
Average air piston displacement at 70° F.,

CUDIC LECL ante ou eee Wet Cats hen iat pe poh wye.as Tia

Average metered output corrected to 70° F.,
CU DEGHLOOL See mea ibe td direst beets Sat, eb wnat sl 758

150 AIR COMPRESSION AND TRANSMISSION

The average evapoiation of 8.72 lb. of water from 131° F. to an
average steam pressure of 115 lb., is equivalent to 9.83 lb. of water
evaporated from and at 212° F. per pound coal consumed. At
the average compressor speed of 63.1 revolutions per minute, the
metered output was equivalent to 758 cu. ft. of free air per minute,
the piston displacement being 1,172 cu. ft. per minute. Fig. 117
presents some of the principal data of the test run on this compressor.

A -CostofCoal perf. HP perAnnum...

Tanaris B- * and Labor-

C-\ 9° "| Labor, int and Dep.
D-Total Cost per Annum.

Ace Ae
Poe mol BOd BEE ees Ge ee ee

IPAS

WAS VAs

ONO TB fea etc | cee

ieee] SS esl | a} [SY Aas ee ae]
BEES ibaa Boees

oe Vee fens |e ea
i enamel Ys Pa a Soe |

ied a Ls Fe eS
seh eel oe Feta i) Sl tla ay ete
Fea og a i a ee

Fic. 118.—Test results of compressor plant No. 1.

To find the average operating results, the curves at 63 revolutions
should be followed, at which the indicated horse-power of the steam
cylinder was 160.8, and that of the air cylinder, 123, showing the >
mechanical efficiency to have been 76.5 per cent. ‘The theoretical
horse-power required to compress isothermally 1 cu. ft. of free air
per minute to 96 lb. (the average pressure) isc.129. The theoretical

HEP ECT OF ALTITUDE AND COMPRESSOR TESTS 151

useful work done by the compressor is, therefore, 758 X0.129 or
97.8, and the net total efficiency of the compressor is 97.8-+161 or
60.8 per cent.

The cost data were furnished by the owner and are based on one
year’s operation. In Fig. 118 these costs are plotted, showing
how the cost per steam horse-power, per year is affected by the

i plea ep
Nea SNCS Goooeeeeee
EN Saale beets
tas pam melanie taal
eal Magee menienal ik
ied eesti ri
elie roa ees estan Poe |
Sh ele FUG a
ait Smo e les iaia
ue SS CEA ee a
on Se eeeeee res
mises
1a a
Stl Recieve:
ac ela RRA REM
2 ail Mieisiehs elalelshlal
Se lise: Remmi alsin
SHER Ft
eu ia Baer DEL
Beaeld seroeiile lab ieee
ia ies Melosh |
abd Song Taam
i Hloe a oP ST
iia is ce Pe InP JN
fafa a eae a ae IT
esa Lo |
R. pM. >

Fic. 119.—Test results of compressor plant No. 1.

average running speed of the compressor. The curve of Fig.
119 shows the operating costs in another way. ‘These costs may
be read in terms of 1,000 cu. ft. of free air compressed to too |b.
or 1,000 cu. ft. of compressed air at too lb. gage pressure.

Test of Plant Number Two.—The plant consisted of three
150 h.p. return tubular boilers, supplying steam for a Corliss engine,

152 AIR COMPRESSION AND TRANSMISSION

the air compressor, and steam heating. To determine the boiler
horse-power, a meter was placed on the steam-pipe to the compressor
during the test run, so that only the portion of steam actually used
by the compressor was charged to the same. The compressor was
duplex, with Meyer valve gear, simple steam cylinders 14X22 in.

real ats ee | | a pe
of TT A a
ce a A
STA an ie
a] ee | ea
BB 02 oo
wo EEA g
A 5 0g
TSE 4 caer:
A A A 81s
le 1 a lays relies
oe in EE AAD, ail 2
BURN a bales
i MPV per
x 5 x
- 4 fe 70 e a
HEME BAVOGGn ears
BG MOO Anime. i.e
RIGA SIRE, a
HOPE He tat
h/ 3
IM WAU/ ApS Cle E ee). in
Bar
30 100
20 0

0 20 40 60 80 100 ~=3=—-:120
R. p.M.

Fic. 120.—Test results of compressor plant No. 2.

and two-stage air end, 14 and 22 X22-in. stroke, rated by the manu-
facturer at 1,050 cu. ft. of free air per minute at 105 revolutions.
At this plant the test lasted over a month, with the following
results:

EFFECT OF ALTITUDE AND COMPRESSOR TESTS 153

otal coal consumed apounds.. <n. a. ti. 05,2 e0rau 450,250
LOCAL Leeds Water epOUunUs tea oo eyes suctod wt 2,496,000
Average evaporation per pound coal consumed,

POUNGS 5 ie eee eee atte take Sheets 5.46
Average revolutions per minute............... 36
Corresponding average Micatec meaner

CFrOMIECUEVG) to, Warren ian, ee ee ead 53

HHatHt+ee 5 daborn Gs Pacts in
2 ee cate [kN lew RAS Ob Le |
= SCENE EEE
> NIC e ANG eee ee
: ACO Ce ao ae
E aaah e GRanee cnn dea
>
ena | pel
a eilghel
Serta iat)
adit a bess Chad OE
ia) Be Wa S2S0RR an
eo Les Siete
= LS De aes es
: Cease
5 aba Ca a el ae
a a ex eesti ee Ves ees
= 100
a

0 10 20 30 40 50 60 70 80 90 100.
R. p. M.

Fic. 121.—Test results of compressor plant No. 2.

Hourly readings of the revolution counter were taken, showing
an average speed of 36.05 revolutions. At this speed the steam
consumption was 51 lb. per indicated horse-power hour, as measured
at the throttle, the air meter showing a delivery of 275 cu. ft. of free
air per minute. The total efficiency was 67 per cent. Taking the
ordinary method of computing the mechanical efficiency only at the
same speed, there would be 48 air h.p., divided by 54 steam h.p.,
giving an efficiency of 89 per cent.

154 AIR COMPRESSION AND TRANSMISSION

The coal consumption per indicated horse-power per year, as
shown by the books of the company, amounted at the average speed
to about 56 tons. Figs. 119, 120, 121, and 122 present details of the
test on this plant, which was conducted similar to that on plant
No. I.

—152
[eg a sale
4 Ra Be KI de vali)
ai
PEELE EEE}
VP)
o
i
So
S
7 Se
S
El aN IS ICS
7, SRR RA GR BRAS EAS maee
Qo.
=
eh EERE EERE 3
rd UE EMI ASE lL
2 SSCS ese eee
oO
SII F i ae Sa ies eee
We #2)
ea aw Reo lalae
ea ea BAG elaciealees alla
5 [aa ae aa | elo
seein 2
oa eae BS WMEIDILI Le
EE Se

AY)

Fic. 122.—Test results of compressor plant No. 2.

Test of Plant Number Three.—This plant consisted of two 125
h.p. return tubular boilers, supplying steam for a non-condensing
cross-compound air compressor of standard make; steam cylinders
18 and 35 X24 in., air cylinders 14 and 28X24 in. A two weeks’ run
gave the following results: |

Total-coalbutned, pounds j-\..) eee 221,190
Total fced-water, cubic feetiay a. sue nas fae 7a
Total feed-waters pounds .f20.7) eee 2,094,057

Average temperature feed-water, degrees Fah-
Tenhel tie. chaos cue OM Pe wih ee ar 154

EFFECT OF ALTITUDE AND COMPRESSOR TESTS 155

Average evaporation per pound coal con-

SUMMER SOU Ge wager net yc ents ira ster at. wis 9.48
Avetage boiler horse-power.. 0...’ s05 fag ees 208
Average revolutions per minute: ..55 2.0. 06 cc. 66
Average indicated horse-power of steam end, at
OUT DAMM aloe Cury Gy maa ons teat ok avg es 210
_ Average indicated horse-power of air end (from
CULEVG) RANE eRe Ie, ore haan Naa T2005
Ayeracersteam preseureayasue nadine 22s ee eos oon 97
Average Gi Pressure sje. wet seta tae yas plas ee 07
Average outside temperature, degrees Fahren-
eT teat ete Cee emer nr AE cred SM cee Lig 3 . 23
Average air piston displacement at normal
SPecd cubic resteati goa Hea: af ects Wes Tey
Metered output in cubic feet corrected to 70° F. 734
4 CHES EE<a
[SIE eas aes a a
Pa CEES PAE Pe CS ae
Ee] ee -eA--RER-- H+ VAAN head |
Five MeMowmMGmmlei7icitee sk
PERE Aga .
pe Ie ea vac =
Beet acgU lace 5
0s el ES Sa FaRURSEs &
oh [Bil SBE I Sala Laeadicl| alles
eet (ea aT a a |
ee iS A) gel ats ®
BERS Zcn eee alee
daa Es Ca is wa he
ott} | A ea ore, patel | | bos As00°
YN eh cet +} °
ara eee aa oe
aaa CS ESP aU A
PEAR be ke sae 7
re Le | YA Pe Like ae fk CUE aE ie
0 10 20 30 40 oat be ce 70 80 90 100

Fic. 123.—Test results of compressor plant No. 3.

The average evaporation of 9.48 lb. of water per pound of coal,
from 154° F. to an average steam pressure of ¢7 lb., is equivalent
to 10.4 lb. of water evaporated from and at 212° F. At the average
speed of 66 revolutions, the displacement was 1,240 cu. ft. of free
air per minute, while the metered output was 734 cu. ft., showing a
net volumetric efficiency of 59 per cent.

156 AIR COMPRESSION AND TRANSMISSION

To determine the conditions in average operation, the curve at 66
revolutions should be followed (Fig. 123), at which the indicated
horse-power of the steam cylinders was 210, and that of the air
cylinders, 128. This shows the efficiency to be 61 per cent., the
friction loss being 81.5 h.p. or 39 per cent. of that delivered by steam

Be GR
se [Pa Ba Sa ce. saya arid Depreciation:

+ ,Depn, Repairs & Supplies.

\
<£)oE) Sas an aa
af a a
BPS EI apart
Re en — 2 PARSE RRP |
FE ete falc ec
aE} ee
BSE Vin OM EORGLE sae
SA a CPL

aia Nate ste [art aiat aaa a
oT =aTaN Neal aa
BUEN CERRORESE ROLES
he al
BERGEN ce Sec ce
fa] Nes nT ISIS a ea a
HN BRNSSEE

Pi STASIS IOASS
Mars Ee

es naa ST

0 10 20 30 40 50 60 70 80 90 = 100
R. p.M

per ILH.P per Hour

Dollars

Nae
ie
NN
ESS

NI
Ss
EI

Fic. 124.—Test results of compressor plant No. 3.

end. This extremely high friction loss was due to the fact that the
compressor shaft was out of line, and the plant could not be shut
down long enough to rectify it. The details and results of this test
are interesting in exhibiting the inefficiency that may be caused
by a purely mechanical defect. (Figs. 123, 124, and 125).

Test Number Four.—The results of a test on another plantare
given in Fig. 126, the details of the boiler test and of the costs
being omitted. In this case the compressor was of the tandem
compound non-condensing type, with Corliss valve gear for the

PEP ECTAOR ALPIRODE AND COMPRESSOR: TESTS Se157

steam cylinders. The test shows that, at a low speed, the steam
consumption increases more rapidly than with the Meyer type of
valve.

Summary.—The results of these tests are enlightening, in showing
the actual amount of the losses occurring in the compression of air,
particularly when the compressor is operating under the unfavorable
conditions of varying air consumption necessarily obtaining in

26 208

2 92 8
a

AG 76s
2 2
=20 160.=
2 Fs
£18 144 >
: ae {
Fig eas OE 128 8
RCT Na a
Sig AN alee tl el he a
8 Baca 2
Sie %6 3
: ia :
5 0 80
me eB 2

8 — ene
[as ae aSs = 5

AGES Shae ee vas
Hae

32
100

Bt
ES
Goer 10 20 30 40 R Oe 60 10 80 90

Fic. 125.—Test results of compressor plant No. 3.

mining and other work in which machine drills play an important
part. These losses are always recognized as existing by compressor
builders and by intelligent users, and it is clearly desirable that
properly conducted tests should be made more frequently.

Again, compressor plants generally develop less power than their
full rated capacity. It should be remembered that an air compressor
is essentially a variable speed machine, its speed being regulated by
some form of throttling governor, connected with the air-pressure
regulator. The machine is therefore called on to run only as fast

158 AIR COMPRESSION AND TRANSMISSION

as the demand for air may require. It may be suggested that it
would be well for compressor builders to give in their catalogues
the actual horse-power rating at different speeds, with a table of

ol RADSRGE Ea mE me

SB laaaeimiee ieee aera ma a
Bo A

1600

S
oS
Oo

>
oOo
@o
c=)
o

o
35 © 4 600

0 “10 pave 80 30 i000

Fic. 126.—Test results of compressor plant No. 4.

efficiencies at different loads and speeds, just as is done by some of
the manufacturers of electrical machinery. Catalogues might also
include some definite data respecting the cost per horse-power
delivered by the air end of the compressor at different working
speeds.

CHAPTER XIV

RECEIVERS. MEASUREMENT AND TRANSMISSION OF
COMPRESSED AIR

RECEIVERS

The purpose of installing a receiver is four-fold: First, to equalize
the pulsations in the air coming from the compressor; second, to
collect the water and grease held in suspension by the compressed
air as it leaves the compressor; third, to reduce the friction of air
in the pipe system; and fourth, to cool the air as thoroughly as
possible before entering the transmission system.

It does not act primaiily as a reservoir of power, for in order to
accomplish this its size would become impractical. However, in

8 ‘ For 4'Satety
=3 = ff Lhe | Valve ~,
8 5 Lg) _ foe ». .

‘|
WTAE
4 i]
|
AWN AH
HI Ht Hk Hl ie
ay Hua H '
fa, pest SISISISESIS SISESISESISLSS 3
ptt V9 its
ghee at pS eR

N
a,
eH

LZZ

ae, 4
Lary)

Fic. 127.—Receiver aftercooler.

compensating for the air pulsations it maintains constant pressure
in the pipe line and in that way reduces friction.

In order to facilitate the removal of water from the compressed
air, it is frequently equipped with a coil of pipes (Fig. 127) filled with
cooling water, in this way serving as an “after-cooler,” as it is
called. When so equipped the difficulty with water in the trans-
mission line and frost at the exhaust pipe of a compressed-air motor
is reduced.

When the pipe line is very long, receivers are placed at both ends
of the pipe; this increases the effectiveness of the receiver and reduces
materially the pipe friction.

159

160 AIR COMPRESSION AND TRANSMISSION

As manufactured, these 1eceivers are usually supplied with a
pressure-gage, safety-valve, blow-off cock and frequently a man-hole.
They are made either horizontal or vertical and of cubical contents
varying usually from 30 to 400 cu. ft. For exceptional cases as
for compressed air-pressure water systems, they are frequently
made much larger.

THE MEASUREMENT OF AIR AND GASES

‘The measurement of compressed air and gas in the commercial
distribution and sale of these commodities and in testing com-
F pressors has attracted a great deal of

Seen te attention in recent times and excel-

Re m ten pane Fe ul lent articles’ are to be found in the
aa ci technical press. The material here
given has been gathered from these
sources and includes some interest-
ing results of tests made in the
Steam and Gas Engineering Labora-
tory of the University of Wisconsin.
“Standards of Measurement.—
In making measurements it is usually
necessary to ascertain the number
of ‘standard cubic feet’ passing in a
given time. The contents of a stan-
dard cubic foot are determined by
the assumed standards of tempera-
ture and pressure used in defining
the unit of measurement. Scientific
data on gases are usually referred to
the freezing temperature of water
: and to the mean barometric pressure.

Fic. 128.—Wet displacement Common commercial standards of
meter. temperature and pressure in gas
measurement are 60° F. and 30 in.

jini

a uideaidtidtin nn : y
it i ce Md

of mercury, respectively.

‘““A quite general classification of meters includes two main types:
volumetric meters and velocity meters.

‘Volumetric Meters.—Volumetric meters include what are known
as “dry meters,’ operating on the general principle of a bellows, and
‘wet meters.’ The latter are built in large sizes for use at gas works

1The Measurement of Gases, Prof. Carl C. Thomas, Jour. Franklin Inst.,
Nov., 1911. Measurement of Nat. Gas, Thos. R. Weymouth, Jour. A. S. M. E..,
Nov., 1912. Flow of Gas through Lines of Pipe, Forrest M. Towl, Lecture
Columbia Univ., rort.

MEASUREMENT OF COMPRESSED AIR 161

in measuring the gas, as made, before being passed for storage to the
holders (Fig. 128). These meters are known as station meters,
their construction is, in general, that of a drum revolving within a
cylinder or tank which is more than half filled with water. The
revolving drum consists of a shaft carrying three or four partitions
arranged in a spiral form. These partitions emerge in turn from the
water as the shaft revolves, and each forms with the water a water-
sealed compartment, which alternately receives and delivers gas.
The drum receives its motion from the pressure of the gas itself and
the number of revolutions of the shaft when properly calibrated give
an index of the quantity of gas passing through the meter.

In testing air compressors, volumetric methods of measuring the
air compressed are sometimes used by installing three tanks. The
compressor is arranged to discharge constantly into one of these
at a constant pressure. This tank in turn discharges alternately
into either of the other two. It fills one tank while the other is being
discharged to the atmosphere and when the pressure approaches
that of the compressor the discharge is turned into the empty tank.
By noting the temperature and pressure and having the volume of
the two tanks it is possible to calculate the volume of air which each
has received from the compressor.

“Velocity Meters.—Volumetric methods of measurement, how-
ever, are not always feasible nor very satisfactory, and other methods
of measurement depending on the velo-
city of flow of the air or gas have been
developed and made use of in commercial
work. These methods may be separated
into three types: the orifice or Pitot-tube
type, which depends for its operation upon
fundamental principles of hydraulics; the
Venturi meter, which depends upon
thermo-dynamic principles involved in the
adiabatic expansion of the gas or air as it
flows through the reduced cross-sectional
area of the Venturi tube; and the heat
meter, of which the Thomas electric meter,
manufactured by the Cutler-Hammer Co., fy¢. 129.—Simple form of
Milwaukee, Wis., is the best example, in Pitot tube.
which the temperature of the gas or air is
increased through a known range by a measurable amount of heat.
From a knowledge of the specific heat of the gas and air, the weight
of gas or air flowing through the meter is automatically determined
and recorded.

“Ditot Tube.—The Pitot tube (Fig. 129) affords a means of
11 .

162 AIR COMPRESSION AND TRANSMISSION

measuling the velocity of air or gas through a pipe at any given
point in the pipe section. In its simplest form it consists of two
small tubes inserted in the pipe line—one having an opening pointed
up-stream and communicating to one end of a U-tube the pressure
due to velocity head in addition to the static pressure in the pipe;
the other having an opening at right angles to the direction of flow
and communicating to the opposite end of the U-tube the static
pressure only. The difference between these two pressures is the
pressure due to velocity alone, and from this, velocity of the gas or
air in the pipe can be computed by means of the formula v?=2¢h
where h is the static head necessary to give to the air or gas a veloc-
ity of vft. persecond. From a knowledge of the cross-sectional area

as Slots in Sides
or Outer tke wane Z
‘ ‘ LY

Fic. 130.—Modern form of Pitot tube.

of the pipe and the density of gas at observed pressure and tempera-
ture, the quantity passing per unit of time can be computed. Fig.
130 shows a modern type of Pitot tube.

“The velocity of gas flowing through a pipe is not the same at
all points in the section. It falls off gradually from the center out-
ward and very rapidly near the inner skin of the pipe. In order to
obtain accurate results with Pitot tubes, without exploring the pipe
at several different depths, it is necessary to ‘standardize’ the tube
and pipe together and find the depth at which the tube will indicate
the mean velocity; that is, a Pitot tube will not necessarily give
consistent readings if placed in a given position in pipes of different
sizes, different conditions of surface, etc. The tube must be located
with special reference to the size, shape, and condition of pipe with
which it is used. Great care must be taken that the openings

MEASUREMENT OF COMPRESSED AIR 163

through which the pressures are communicated to the U-tube are
properly placed with respect to the direction of flow, and they must
be kept free from deposits.

“The general formula for the Pitot tube and the orifice is derived
from the law of falling bodies. Let

T = absolute temperature of flowing gas, degrees Fahrenheit.

P = absolute pressure of flowing gas, pounds per square inch.

w = weight per cubic foot of flowing gas, at P and T.

G = specific gravity of flowing gas, (air 1.0).

v = actual velocity of flowing gas, feet per second.

h; = height in feet of homogeneous column of gas at P and T

producing 2. 3

h = corresponding height of water column in inches.

W. = weight per cubic foot of water, 62.37 lb. at 60° F.

P, = absolute storage pressure base, pounds per square inch.

[, = absolute storage temperature base, degrees Fahrenheit.

Wa = weight per cubic feet air at 32° F. and 14.7lb.=0.08073 lb.

d = actual inside diameter of pipe or orifice in inches.

E = efficiency of Pitot tube or orifice.

Q = flow in cubic feet per hour at P, and T,.

Then
V=A/2gh; =|2 8 a
ie
W=WaG aT
vy ae pez
OES ita Pal
O= 218.44 ka? a

“Prof. S. W. Robinson who was probably the first to use the
Pitot tube in connection with the flow of natural gas has developed
the following formula which has heen used by natural gas men for
a number of years: |

OQ =1,462,250 d? es) 0.29 os de

This was derived from the formula for adiabatic flow

n—1
oo 2g X44 Po hf Pi\ on ie in which
G1) w |G) geal

164 AIR COMPRESSION AND TRANSMISSION

v = velocity of flowing gas, feet per second.

Py) = absolute pressure of the atmosphere, pounds per square
inch.

n = ratio of the specific heats.

w = weight per cubic foot of gas at pressure P,.

G = specific gravity of gas, air I.

P, = absolute pressure shown by Pitot tube, pounds per square
inch.

d = internal diameter of well mouth, inches.

= open-flow capacity of well, cubic feet per 24 hours.
“Prof. Robinson has computed tables from the above formula
which have been used for years. The computations are based on
the following:

nm =I1.408
2g =04.3
Ps =14.6
Gir=0:0

IE i i Se Rito Ne.

To = absolute temperature of melting ice.

T = absolute temperature of flowing gas.

I’, = absolute temperature of storage.

“Thos. R. Weymouth in his article in the Journ. A. S. M. E.
points out that for natural gas the ratio of the specific heats is more
nearly equal to 1.266 and by using

Le OO nel.
(le ree NE
Vt ae ay)

P, = storage pressure 14.65
the formula becomes:

if 0-21 a
O=1,758,560 ay (44) Sh }*
el AA

“In order to obtain a mean value of / for the use in Pitot tube
measurements Prof. G. J. Davis of the University of Wisconsin
devised the following method which is illustrated in Figs. 131 and
132 showing results of an actual test of a Pitot tube placed tandem
with a Venturi meter and a Thomas electric meter.

‘The horizontal represents distances from the center of the pipe
at which readings of # were observed. On the vertical a suitable
scale of values of \/h is laid off. Readings of ~/h are then plotted
and joined by radial lines to the point representing the center of the
pipe. The intersections of the slanting lines with the perpendiculars

MEASUREMENT OF COMPRESSED AIR 165

representing the positions at which the corresponding values of Vh
were read are points through which a smooth curve can be drawn.
The area under the curve may now be determined and from this
the altitude of a triangle having the same area and base as the
irregular figure will give the mean / to be used in computing the mean
velocity.

‘““The mean velocity V is determined from the formula V? =2¢h
after reducing the # determined to equivalent feet of air. The

pounds per hour will then equal 3,600 AVG where A is the area of
the pipe in square feet.

eee

Gee
A
NAG

\ EN

LAREN GE aces

ae
rey
: ! Val a aa BA
I gt LA
= ieee AA Sa VAC
SSR Zasi ap aaa
3 0.1 FAR =e 5 07 - Puta
A oe. [eal nla
Te ‘CCEA
: FA “OOO sA
} LG 04 J Zt
ott wis =o L EL
02 Eee 2a ae 02 Kt
ale area ae eine it KAZ ie
premanee Sent Center of bier eine Distance from Center oF Pipe, mohes,
Fic. 131.—Graphical method of Fic. 132.—Graphical method of
determining mean head. determining mean head.

V is the velocity in feet per second.

G is the weight of a cubic foot of air as it passed through the pipe.

“‘In measuring air by means of a Pitot tube it is necessary to take
into account the humidity of the atmosphere and make corrections
as indicated in the discussion on Humidity given in Appendix C.

‘‘In measuring large quantities of air in testing air compressors it
is quite a common practice to have the air escape through a suitable
orifice to the atmosphere. An apparatus of this kind is shown in
Fig. 22 and one for large installations in Fig. 133.

“The formula usually used for measuring air under these condi-
tions is

166 AIR COMPRESSION AND TRANSMISSION

ie
W =0.53A Tz where P; is greater than twice atmospheric

pressure.

7 TT, Tt UM, iets
Ui ;

/ “frien, "|

; rT 7 ia ili >

7 i 3 : eit ti! (Uy \
LT. TV. Ahi batons, Hil!

(iil AR aig i ; \

NTR NNN aii AA |i NN

a : a in i ' al
gu a + Ge wy gulls ei

Fic. 133.—Apparatus for measuring large quantities of air.

When P, is less than twice atmospheric pressure the formula usually

used is
W =1.060A A eeieee
163

“This last formula, however, is
A not entirely reliable (see ‘Air
Flowing into Atmosphere through
Circular Orifices’ by R. J. Durley,
Trans, ASS. MAES Y ol27
In the above formule
W = weight of air escaping in
pounds per second.
P. = pressure of atmosphere in
pounds per square inch.
P, = pressure before the nozzle
in pounds per square inch
absolute.

(NNhhawm vrssr aa seees.

4 T, = absolute temperature of
SI] Ss . °
NIN air entering the nozzle.
A = area of nozzle in square

foc
“Tn using a nozzle or orifice it
is also necessary to consider the
humidity of the atmosphere in
measuring air.
“St, Johns Meter.—A number of meters have been made making
use of an orifice for measuring the flow of air. Such meters are usu-

PRaveeaaaarvees

Fic. 134.—St. John meter.

MEASUREMENT OF COMPRESSED AIR 167

ally calibrated by means of a gasometer. The St. Johns meter,
Fig. 134, is in effect a variable orifice meter. The position of the
plug S determines the size of the orifice through which the air passes
and a graphical record is kept of the position of this plug on a drum
moved by clock work and by planimetering this chart the average
position can be determined and the consumption be calculated.
‘Venturi Meter.—The Venturi meter, Fig. 135, consists of a throat
or gradually contracted portion of the passage, which causes a de-

{AU IOOU MOOI

Fic. 135.—Venturi tube.

crease in pressure and increase in velocity of the gas flowing through

it.

area of the up-stream section in square feet.

Az = area of throat in square feet.

P, = pressure at up-stream side, pounds per square inch.

P., = pressure at throat, pounds per square inch.

G, = weight of gas at up-stream section, pounds per second.

n = ratio of specific heats, constant pressure to constant
volume.

V2 = velocity of gas at throat, feet per second.

te
O
o
aN
I

“By equating the loss in potential energy to the increase in kinetic
energy it is found that

N
bo
|
Pgs
a
Saas
ty]
Pt aa
to
oR
S
eer 2
NH
=
|
ST ee
|
a
| 3
|
Lon]
dole

pe 4a \7(£2)\2
ae AyIEN Ps
1

ce . ° ie ee °
The quantity flowing O=A2V.G1 (3) ” in cubic feet per second.
1

“Tt is frequently necessary to take small readings of pressures
with both the Pitot tube and Venturi meter, and in order to do this

168 AIR COMPRESSION AND TRANSMISSION

accurately the water columns should be read with a micrometer
gage or differential (inclined) water column.
‘“‘A similar formula for the flow expressed in cubic feet per hour

would be
Po ned
Ts n Jey ‘poe ; Gal
= O40 Ag) ee eee 2\ in
Cee PNG JT; (>,)

“Terms in this formula not appearing in the other are

Q1 = flow in cubic feet per hour.

T, = absolute temperature of gas at entrance.

I, = absolute temperature of storage base pounds per square inch.

P, = absolute pressure of storage base pounds, per square inch.

G = specific gravity of gas, air=1.

“Thomas Meter.—The Thomas electric meter is based upon the
principle of heating the air or gas through a known range of temper-

y L <p SS SS

PASS aN [ASS

SAA COLL ALLL TOTO ly Zz y

Gee we beer rr}

Fic. 136.—Sectional view of Thomas electric meter.

ature. This measured energy is proportional to the weight of the
gas flowing. Electric energy is used as a source of heat because
it can be so conveniently and accurately applied and measured.
Electrical resistance thermometers are used to regulate the tem-
perature range through which the gas is heated, because with ther-
mometers of such type very small differences of temperature can be
accurately and easily determined.

‘““A sectional view of a ro-in. meter and casing is shown in
Fig. 136. Fig. 137 shows the meter diagramatically. The electric
heater is placed within the casing between two electric resistance
thermometers, 7; and 7». The heater consists of spiral turns of
bare nichrome resistance wire wound around a conical frame and
supported by insulators, so that heat is dissipated evenly over the
section of the pipe. A rheostat is placed in the heater circuit for
regulating the direct current supplied. This energy is measured
by an ammeter and voltmeter.

“The thermometers consist of resistance wire wound on wooden-

MEASUREMENT OF COMPRESSED AIR 169

spindles and evenly distributed over the section of the casing. The
wire is of such material that its resistance increases with the tempera-
ture according to known laws. The two thermometers form two
arms of a Wheatstone bridge, the other two arms being fixed coils
of wire that have a zero temperature coefficient. A galvanometer
is connected across the Wheatstone bridge, and a small rheostat is
placed in series with one thermometer for balancing the bridge when
no heat is passing through the heater. A small resistance R; is
arranged so that it can be placed in or out of series with the entrance
thermometer. ‘This resistance is equal in value to the increase in

Balancir
rheosn

y Pipe
ie w
Hew r Ley
aS Se
peel le RS i\/| §38
Direction of 2 S ESS
Flow SSE Xss
S | Heater x
ie aN |
Water
( g Rheostay

Ammeter

S
>
3
\.
iS

Fic. 137.—Diagrammatic sketch of Thomas electric meter.

resistance of the exit thermometer for a rise in temperature of prac-
tically 2° F. The meter in the laboratory of the University
of Wisconsin is for 2.0152° F.

‘The operation of the meter is as follows: With gas flowing through
the meter but with no energy in the heater, and with R; out of cir-
cuit, the two thermometers are brought to the same balance by
means of the balancing rheostat and the galvanometer. Then the
resistance R: is put in circuit and sufficient electrical energy is
supplied to the heater to bring the galvanometer to balance again,
by bringing the exit gas to a temperature 2.0152°, with the meter
mentioned, higher than that of the entering gas. The measuring
instruments in the heater circuit then indicate the energy required

170 AIR COMPRESSION AND TRANSMISSION

to raise the temperature of the air or gas through a known range.
The quantity of gas flowing can be found by the equation

Ww a 3412k
ts

where W is the number of pounds of gas or air per hour, # the amount
of energy supplied in watts per hour, ¢ the rise of temperature in
degrees Fahrenheit, and s the specific heat at constant pressure of
the gas or air.

“With the laboratory meter the air flowing through the meter
per minute is given by the formula 0.028218 E;.

“In applying this meter to gases it is necessary to ascertain the
composition of the gases in order to obtain the mean specific heat
for use with the meter.

“The meter in commercial form is equipped with automatic
devices to regulate the flow of current through the heater so as to
maintain a constant difference of temperature between the resistance °
thermometers of 2°. The electrical instruments for measuring the
consumption of current in the heater are then calibrated to read either
weight or quantity of gas flowing and this reading is recorded
graphically.

“Meter Comparisons.—At the University of Wisconsin tests
were run by placing a Venturi meter, Pitot tube and Thomas meter

SAG To ‘
Sweet
Pitot TORE fom Atte

0
Driven Fan

Fic. 138.—Sketch of meters placed tandem for testing.

in tandem, as shown by Fig. 138. The results of these tests are shown
as Fig. 139. A remarkable similar set of readings were secured.
“In April, 1911, a Thomas meter was tested on a natural-gas line
by comparison with Pitot-tube measurements giving practically iden-
tical results. This meter had a maximum capacity of 750,000 cu. ft.
of free gas per hour and an accurate minimum capacity of 12,500 cu.

MEASUREMENT OF COMPRESSED AIR 171

ft. It gave a continuous graphical record and integrated values of
the gas directly in standard cubic feet at 15.025 lb. absolute pressure
and 60° F., although the pressure of the gas varies from 50 to 200 |b.
gage and the temperature varies with weather conditions. The
specific heat was calculated from an average analysis of the gas for
the standard conditions given above. This particular meter was
placed in a ro-in. line and located about a mile and a half from a very

ena, co
+ Scere ts

Gas Engineering ee
University of Wisconsin, June 2-/9/1

Electric Meter «
Venturi Meter x ———
Pitot Tube o—----

0
0 100 200 300 400 500 600 700 800 900 1000 1100
R.p.M. of Fan

Fic. 139.—Result of test.

complete Pitot-tube meter station. A 22-hour comparative test
showed a difference of 0.2 per cent. between the two meters and a
similar comparative test from April 17 to June 3, 1911, showed the
same difference.”’

_ PIPE LINES!

“The transporting of gas or air requires a line which shall be “air
tight.” It is much more difficult to make a line to hold gas or air
under pressure than it is to hold a liquid. Trouble has been expe-
rienced in almost all lines built for high pressure on account of the
leaking of the gas at the couplings. The first high-pressure lines
were laid with bell and spigot joints, caulked with lead. The lines

1 Forrest M. Towl, Lecture Columbia University, 1o1t1.

siz AIR COMPRESSION AND TRANSMISSION

might be tight when they were first laid but the movement in expand-
ing and contracting soon caused them to leak in large amounts.

‘The next lines used were of wrought iron or steel pipe with screw
joints. While these held much better than the bell and spigot type,
there was still enough leakage to make it desirable to have a more
perfect joint. The leakage on some of the earlier screw-joint gas
lines was such that by putting a rubber bag over the coupling, gas
could often be collected at the rate of from 20 to 50 cu. ft. per hour,
or enough to run a good-sized torch. This was true of lines up to
8 or ro in. in diameter. When the lines became larger, the leakage
increased so much that it was practically impossible to use large
size lines and get a large percentage of the product at the market.

‘‘As the demand for natural gas increased, it became necessary
to use larger lines, and a rubber packed stuffing-box was developed.
The first successful joint of this kind in the market was the Dresser
coupler, and it is due largely to this and other couplings that the
natural-gas industry has become so great. |

‘Dresser Coupler.—The Dresser coupler consists of a sleeve into
which the ends of the pipe are placed. There is a projection at the
center of the sleeve so that the ends of the pipe
will be each inserted into the sleeve the same
distance. This sleeve acts as a follower to
compress rubber in an annular space into the
end rings which are drawn together by bolts.
The rubber is surrounded on one side by the
pipe, on another by the body of the coupling, and
Fic. 140.—Dresser On the remaining side by the end rings so that

pipe coupler. there is very little of the surface of the rubber
exposed either to the gas on the inside or the

air on the outside of the line. It is found that these joints will last
for years. (Fig. 140 shows a cross-section of the Dresser coupler.)

‘‘“Hammon Coupler.—The Hammon coupler is a modification of
the Dresser, one of the principal features of which is that the pro-
jection at the center of the sleeve is made by lugs welded onto the
sleeve. When it becomes necessary to take apart one of these
couplers, the lugs can be broken off and the coupler slipped back so
as to allow the pipe to be easily removed. (Fig. 141 shows the Ham-
mon coupler.)

‘Lines of pipe can be built in almost any kind of country, but it
is necessary in some places to arrange to keep the lines from acting
as a Bourbon tube and expanding in one direction until the ends
of the pipe may be pulled out of the coupling. To avoid this trouble
it is customary in such places as river crossings to use screw pipe,
and to place over the collar a clamp which is constructed to make a
rubber joint between the ends of the collar and the pipe.

\
) } tt

iain

PS SS SSS SB

RS ca

ereveereeenrrnrnrernrrrrrr rr

MEASUREMENT OF COMPRESSED AIR 173

“For power-transmission lines or for temporary gas lines, where
the distances are short or the service temporary, it is not considered
necessary to bury the pipe, it will be found that the screw-joint pipe
is satisfactory, but for other
natural-gas or air service, the
rubber coupling has many things
to recommend it, and when the
capacity requires large pipe, it is
almost absolutely necessary to use
this type of coupling. These
coupling have been used for manu-
factured gas, but it is found that
the condensation from the gas
collects in the coupling and soon
causes a leak in the rubber joint.
Work is now in progress to per-
fect a material which will not be
acted upon by the condensation
in the gas and which will make a .
gas-tight joint. Fic. 141.—Hammon pipe coupler.

In using screwed joints for air
it is necessary that the lead, litharge, or other material used at the
joint should be applied on the ends of the pipe and not in the coup-
lings, so that the surplus is brought outside instead of within the
pipe where it may cause a more or less serious obstruction.”

Pipe-line Formulz.—A very simple formula is often used for
calculating pipe lines for compressed air.

Deo a) or pee pi~ bs
Wy V/] Wi

in which

D =the volume of compressed air in cubic feet per minute dis-
charged at the final pressure.

¢ =a coefficient varying with the diameter of tke pipe, as
determined by experiment,

d =diameter of pipe in inches,

1 =length of pipe in feet,

pi =initial gage pressure in pounds per square inch,

p2 =final gage pressure in pounds per square inch,

1 The actual diameters of wrought-iron pipe are not the same as the nomina-
diameters for all sizes. This difference is small, however, except in the 1 1/4 in.
and 1 1/2 in. sizes, the actual diameters of which are 1.38 in. and 1.61 in.
respectively.

174 AIR COMPRESSION AND TRANSMISSION

w, =the density of the air, or its weight in pounds per cubic foot
at the initial pressure #1.
The second form of the formula, as given above, will be found
convenient for most calculations, as the facto1s can be considered in

groups.

In Tables XIII and XIV are given the values of c, d°, and
c\/d5, The values of c show some apparent discrepancy for sizes
of pipe larger than 9 in. but there would be no very material dif-
ferences in the results.

» TABLE XIII

Diameter of pipe, Values of Fifth powers of Values of

inches C d c\/ds
I 45-3 i 45-3

2 5200 a2 207

3 56.5 243 876

4 58.0 1,024 1,856

5 59.0 3,125 3,298

6 59.8 7,776 5,273

7 60.3 16,807 7,317

8 60.7 32,768 10,988

9 61.0 59,049 14,812

IO 61/32 100,000 19,480

i 61.8 161,051 24,800

12 62.0 248,832 30,926

TABLE XIV.—VALUES OF Wi FOR ET ACE EGS pene UP 100 LBS.PER SQUARE

Gage pressure, : | — Gage pressure, =

ap a Vw | pounds - Vw

° 0.0761 0.276 “is 0.3607 0.600

5 0.1020 0.319 60 0. 3866 0.622

ite) 0.1278 0.358 65 eg As 0.642

cf On 527 0.302 70 0.4383 0.662

20 0.1796 0.424 7 o.4042 0.681

25 0.2055 0.453 80 0.4901 0.700

30 O. 2313 0.481 85 0.5160 0.718

25 O.2572 0,507 go 0.5418 Of736

40 0. 2831 0.532 95 0.5677 0.753

A5 ‘0.3090 0.556 Too 0.5936 0.770
50 0.3348 O. 578.2 eva 2 yan a he he al cea OR te ee

MEASUREMENT OF COMPRESSED AIR 175

Mr. Frank Richards gives the following formula for determin-
ing the loss of pressure in pipes:

10,000D°a

from which

10,000D*a

In these equations
D=diameter of pipe in inches.
L =\length of pipe in feet.
V =volume of compressed air delivered in cubic feet per min.
H =head of difference of pressure required to overcome friction
and maintain the flow. |
a =constant depending on the diameter of the pipe.

TABLE XV.—VALUES OF a, D' AND Dia FOR WROUGHT-IRON PIPE.

Gane ; Ds Dia
pipe diameter
I in. 0135 I 023%
1% in fans RES Ta52s
Tz in 0.662 7.59 5,03
ain, 0.565 ex 18.08
anit 0.65 97.65 63.47
3 in 0.73 243. ETA
33 in O87 525. 413.2
4 in 0.84 1024. 860.2
5 in 0.934 2125. 2010075
6 in I .000 vie by doe Os
8 in Te ks 32768. 36864.
ro in pane 100000. 120000.
2-10 T20 248832. 213525.
16 in N34 1048575. I4O5001.
20 in 1.4 3200000. 4480000.
24 in I.45 7962624. 11545805.

For example, suppose it is desired to determine the loss in

pressure in transmitting 300 cu. ft. of compressed air per min.
through a 6-in. pipe one mile in length.

L =5280, D*a for a 6-in. pipe=7776
2
Nehey vind eb EBON Ne)

10000 X 7776
6.11 lb.

That is, the pressure drop will be

176 AIR COMPRESSION AND TRANSMISSION

As another example, suppose it is desired to ascertain the
proper size of wrought-iron pipe for transmitting compressed air
from a compressor of 1500 cu. ft. free air capacity per min. at
80 lb. gauge a distance of 2000 ft., with an allowable loss of
pressure of 5 lb.

The pressure at delivery will be 75 lb. gauge or. practically 6
atmospheres. ‘The volume of compressed air delivered per minute
will be:

1500+6=250 cu. ft. per min. =V

As H=5 the formula
Paes

tooo0oH
250” X 2000

10000 X 5 ere

From Table XV it is seen that D*a for a 5-in. pipe is 2918.75
and for a 6-in. pipe 7776. This would indicate the advisability of
selecting 5-in. pipe for the conditions of this problem.

The friction in pipe elbows may be expressed in terms of
equivalent lengths of straight pipe. Elbows having the largest
radius will naturally give the least friction and the accompanying
table as given by the Norwalk Compressor Co. gives the friction
effect of elbows in terms of the radius.

may be used with proper substitutions, from which

TABLE XVI.—FRICTION EFFECT OF ELBOWS IN TERMS OF PIPE LENGTHS

Radius of elbow in
pipe diameters

On
Ww
bo
eS
dle
eS
|
vol
ated
[<i

Equivalent lengths of straight

pipe in pipe diameters 7.85 |8.24 9.03 |10.36/12.72/17.51|35.09 leas

REHEATING

From a consideration of changes that take place during the
compression and expansion of the air, it is apparent that heating the
air just before expansion will raise its temperature and impart to it
an increase of energy which, if used immediately, will increase the
efficiency of the compressed air. In addition, this reheating will
increase the temperature at the end of expansion and prevent the
particles of moisture in the air from freezing.

It is not uncommon in the ordinary use of compressed air to find

MEASUREMENT OF COMPRESSED AIR aT

exhaust temperature varying from 5° to60° F. The lower tempera-
ture is very apt to cause trouble particularly in out-door work
during the winter months. It is quite probable that reheating was

5 FL =" *
1C) (|| 4Inler
Sen

p
Soe) —
ieneatareeeseneeeny = 5
4
Ns s
— iy
f S
a)
4 kay
i
(i

RRS ones
verti T3054
RRR Rr fae

POOR

SER
eres

SIV

PSR
H IRS
pe

Reheater Coil

Fic. 143.—Leyner air reheater.

first introduced to prevent the formation of frost in the exhaust pipe
and its advantages are so apparent that no economical use of com-
pressed air in large installations is complete without some system of
reheating.
Stoves.—In quarry work it is very common to make use of stoves,
12

178 AIR COMPRESSION AND TRANSMISSION

such as are shown in Figs. 142 and 143, for heating the air just before
its use in the drills. In locomotive work for mines and surface use,
hot water is frequently used to heat the air. <A recent locomotive
built for the government for handling cars of explosives has a com-
pound cylinder, expansion taking place in one to such a point as to
bring the temperature below that of the atmosphere. ‘The air then
passes to the low-pressure cylinder through pipes in contact with the
atmosphere and in this way its temperature and energy are increased.

Reheaters are usually capable of raising the temperature of the air
to from 300 to 500° F., although common practice shows temperatures
from 250 to 350°. In figuring on reheaters it is usual to assume that
t lb. of coal will give from 8,000 to 10,000 B.t.u. to the air. As the
specific heat is 0.0686 B.t.u., it is evident that 1 lb. of coal will heat
33 lb. of air or about 408 cu. ft. to a final temperature of 350°. As
reheating increases the volume as well as the temperature, the econ-
omy in its use expansively is quite evident.

CHAPTER XV

THE SELECTION AND CARE OF AIR COMPRESSORS

It is difficult, if not impossible, to dictate the type of compressor
that must be selected to suit certain conditions, as the importance
of the various factors influencing the selection vary greatly in differ-
ent localities; however, a brief statement of some of the factors to be
considered may be of assistance.

Available Power.—The character of driving power available will
usually determine the type of prime mover to be selected. The
principal manufacturers of compressors have on the market the
belt-driven, steam-driven, and electrical-driven machines, as well
as those driven by gas engines or water turbines.

For small and intermittent demands it is difficult to make use of
an economical steam engine for operating the compressor, and for
‘this reason when belt power is available and suitable for the condi-

tions of operation, the belt-driven compressor is to be preferred.

When the quantity of air to be compressed is large, it will usually
be found advisable to install a steam or motor-driven compressor as
this allows greater flexibility and economy of operation.

If this type of motive power is selected, a choice of different de-
signs must be made. The straight-line type, equipped with proper
energy-compensating devices to secure economy of steam consump-
tion, has the advantage of simplicity and high mechanical efficiency.
On the other hand, the duplex type appeals to many, for, if
single-stage, it may be installed in sections and in this way future
extensions can be made with minimum expense.

The pressures that are desired will determine whether the com-
pressor is to be of the one-, two- or three-stage system; the first being
usually selected for pressures below 80 lb. per square inch, the second
for pressures from 80 to 500 lb., and the third for pressures from
500 to 1,000 lb. per square inch.

Valve Gear.—The price of fuel will influence largely the type of
‘valve gear to be selected for a steam engine. If fuel is very cheap
and its consumption comparatively unimportant, the simpler forms
of valve gears are to be preferred. If, however, the economy of

179

180 AIR COMPRESSION AND TRANSMISSION

fuel is of importance, the more complicated and expensive types,
such as the Corliss, should be selected.

The plain slide valve, the independent cut-off valve, such as the
Meyer, and the Corliss give steam consumptions which decrease
in the order named. The last two types, particularly the Meyer,
are quite common for steam-driven compressors.

In the mountainous sections of the country and in those parts
where the supply of water-power is abundant, water-wheels or
turbines are largely used as prime movers.

The distance.from the source of power to the place where the
compressed air is to be used will determine whether the water-
wheel is to be coupled directly to a compressor or to an electrical
generator which will generate current to be transmitted to an
electrically driven compressor.

Both types are in use in the mining regions, some of the largest
compressed-air installations being equipped with compressors
driven directly by water wheels.

Within the last few years the method of compressing air by means
of a waterfall without the use of any mechanical parts has increased
to such proportions as to demand the attention of engineers con-
nected with the installing of compressed-air equipments.

Air compressors driven by gas or gasoline engines are frequently
used in quarries and other places where it may be desirable to move
the compressing plant from point to point. Where gas can be
obtained cheaply it makes a most desirable machine, because of
the high efficiency of the gas engine. Within the last few years
gas engines operated by the gases from blast furnaces have been
developed to such an extent that many of the large blowing engines
for Bessemer converters are operated in this way, giving compressed
air to the converters and other places where used at a minimum
expense. .

In selecting any type of air compressor, particular attention
should be paid to the construction and design of the valves.
Mechavically operated inlet and automatically operated discharge
valves seem at present to represent the favorite practice, although
automatically operated inlet valves are preferred by many
because of the little attention they require.

All valves should be simple in construction, of large port opening,
durable and reliable in action, and easily removed for purposes
of examination and renewal.

The largest possible amount of surface, including cylinder heads,

SELECTION AND CARE OF AIR COMPRESSORS 181

should be water-jacketed on all piston compressors except those
discharging at very low pressures.

The nature of the work performed by the compressor will deter-
mine the advisability of installing an unloading or governing device,
economy of operation usually demanding the installation of some
such apparatus.

Size and Type of Compressor.—In order to give an idea of the
data required when determining the size and type of a compressor
to be selected, the following is given, as published by the Sullivan
Machinery Company in their catalogue:

1. Purpose for which the compressed air is to be used (coal-
mining machines, rock drills, air lift, etc.).

2. Volume of free air required in cubic feet per minute.

3. Working air pressure.

4. Altitude at which compressor will work, if over 1,000 ft.
above the sea-level.

5. Number, size, and class of machines to be operated by the
compressed air.

6. If the air is to be used for pumps, give make, size, and speed
of pump and height to which water must be delivered.

7. If for raising water by the “‘air lift,”’ state desired flow per
minute in gallons, diameter and depth of well, and height to which
water must be delivered, measuring the average height of water
in the well.

8. Will the demand for air be constant or intermittent during
the daily time of operation?

9. Will the compressor be operated by steam or power?

10. If steam driven, state working steam pressure, kind and
average cost of fuel available, type of engine preferred and whether
it is to be run condensing or non-condensing.

tr. If power driven, state motive power (as water-power,
electricity, rope driven, or gasoline engine) and whether direct
connection, belt, or gearing is preferred.

12. If water-power is to be used, give horse-power available,
or head or fall of water in feet, also amount of water supply in cubic
feet per minute.

13. If belt drive is employed, give horse-power at belt.

14. State facilities for transporting compressor to destination.
If machine must be sectionalized, state means of transportation,
heaviest weight allowable for a single package and number of pack-
ages permissible of maximum weight.

182 AIR COMPRESSION AND TRANSMISSION

COMPRESSED AIR EXPLOSIONS

‘“‘Compressed air claims to be and is a safe power. Occasionally
we hear of a case of firing, which to some may appear to be a serious
objection to the use of air; but if the causes are known and under-
stood and due care is observed, firing becomes a matter of care-
lessness. Compressed air is not inflammable, but, during com-
pression by piston compressors it is necessary to use oil for
lubrication, and this oil or the gases from it form a combustible
mixture with the air.

In most cases firing may be traced to the use of poor oil, but in
others too much oil sometimes causes ignition.

Lubricating Compressors.—It is a common mistake of engineers
to feed oil too rapidly to the air cylinders. A drop now and then
is all that is required to keep the parts lubricated. The air cylinder
does not require as much lubrication as the steam cylinder, for there
is no tendency to cut and wash away the oil as there is in a steam
cylinder.

When too much oil is used, there is a gradual accumulation of
carbon which interferes with the free movement of the valves and
which chokes the passages, so that a high temperature may for a
moment be formed and ignition follow.

It is well to get the best oil and use but little of it.

There are cases where firing has arisen from the introduction of
kerosene or naphtha into the air cylinder for the purpose of cleaning
the valves and cutting away the carbon deposits. This is a very
effective way of cleaning valves and pipes, but it is a source of danger
and should be absolutely forbidden.

The inflammability of benzine, naphtha and kerosene is so
acute that it is a dangerous experiment to introduce anything of this
kind into an air cylinder.

Cleaning Valves.—Soft soap and water is the best cleanser for the
air cylinder and it is recommended even in cases where the best oil
is used and it is a good plan to fill the oil cup with soft soap and
water and feed it into the cylinder, as the oil is fed, at least once or
twice a week, or even oftener if necessary, in order to prevent the
carbon deposit from gumming up the valves.

A thick or cheap grade of cylinder oil should never be used in an
air compressor. Thin oil which has a high flash-point and which
is as free from carbon as conditions of lubrication will admit is the
best oil.

There may be considerable danger in a valve which is so gummed

PEEL CIIONGAND CAKE OPSATR COMPRESSORS, 188

as to be unable to close at the right time. When a piston has
compressed air and forced it through the discharge valve and then
starts on its return stroke, there is immediately a tendency for
the air just compressed to return to the cylinder, and if the valve
does not operate properly there will be some hot compressed. air
in the cylinder when the piston starts again on its compression stroke
with air at an initial temperature, 200° or 300° above the normal.
The final temperature at the end of compression will in consequence
be quite high and may even be above the ignition point of the lubri-
cating oil that is used.

To guard against this, care should be taken in the selection of the
type of compressor used and the valves and passages should be
thoroughly cleansed once a week by the engineer, who should also
investigate the valve seats to insure the valve fitting properly when
in place.

Inlet Connection.—The inlet should be closed in a cold-air box,
or some direct connection should be made to the outside air in order
to avoid taking in hot air to the compressor.

In compressors used in the coal-mining districts, care should be
taken to see that coal-dust is not drawn into the cylinder.

A thermometer should be placed in the discharge pipe close to the
compressor so that the operating engineer may note any change of
temperature and stop or slow down the compressor to avoid an
accident.

APPENDIX A

Common Logarithms.—The following table of common logarithms
will be found of assistance in solving work and power problems
dealing with compressed air.

The table shows the “mantissa”? only, the ‘“‘characteristic”’
depending on the location of the decimal point and being one less
than the number of figures to the left of the decimal point of
the given number. For example:

The logarithm of 529 is 2.7235

The logarithm of 52.9 is 1.7235

The logarithm of 5.29 is 0.7235

The logarithm of 0.529 is 9.7235—I10

The logarithm of 0.0529 is 8.7235 —Io.

The table of proportional parts is for the purpose of interpolation.
For example, if the log of 80.54 is desired it is found as follows: The
log of 80.5 from the table is 1.9058. The table of proportional parts
shows in the same line of figures the additional figure to be added
under the column marked 4 to be 2, making the required log of 80.54
1.9058 plus 0.0002 or 1.9060.

Logs of powers of numbers are found by multiplying the log of
the number by the given power or exponent. For example, suppose
it is required to find the value of 28.31".

The log of 28.3 is 1.4518, and this multiplied by 1.4 is 2.0325, that
1s, the logrol 2823" 2"se290375.

The antilog of this or the numerical value of 28.31:4.is found by
looking in the table for the number whose logarithm has a “‘mantissa”’
of 0325 and then pointing off three places from the left or one more
than the characteristic 2. The antilog of 2.0294 from the table
is 107, and as the given log is 0.0031 higher than 2.0294 the table
of proportional parts would indicate that the antilog of 2.0325 is
about 0.75 higher than 107. That is 28.31:4 is equal to 107.75.

The principal difficulty in handling logarithms of small numbers
with fractional exponents is met in dealing with the characteristics.
This may be treated as follows:

Suppose the value of 0.483°°* is desired.

The log of 0.483 is 9.6839—10

184

APPENDIX A 185

LOGARITHMS.

Proportional Parts.

8
12 3/4 5

0043|0086|0128
0453)0492/0531
0828/0864/0899
1173)1206)1239
1492)1523/1553

0212/0253
0607|0645
0969}1004
1303}1335
1614/1644

0334
0719
1072
1399
1703

-—
=

(SU SCOR at
& &~100
ht pe
BSS
— ee
bo ow

—
—

1790/1818) 1847
2068) 2095/2122
2330| 2355/2380
2577/2601) 2625
2810/2833) 2856
3032|3054/3075
3243/ 3263/3284
3444/3464/3483
3636|3655|3674
3820/ 3838/3856

1903/1931
2175/2201
2430/2455
2672) 2695
2900) 2923
3096/3118) 3139
3304/3324/3345
3902/3522/3541
3692/371113729
3874} 3892/3909

1987
2253
2504
2742
2967
3181
3385
3579
3766
3945

—
—_

bo bo bo 09
He OT OT OI Od
~I ~I ~1 00 00
—
o

OU D> OH HD

4048) 4065/4082
4216)4232|4249
4378/4393|4409
4533/4548) 4564
| 4683)4698/4713
4857
4997
5119/5132
5250/5263
5366!5378/5391

4116
4281
4440
4594
4742
4886
5024
5159
5289
5416

3997/4014
41664183
4330/4346
4487/4502
4639) 4654

m bo bo bO bO bo bo bo bo bo
wwwwo PRP PA

He Or Or Or Or

4786/4800|4814
4928) 4942/4955
5065/5079] 5092
5198}5211/5224
5328/5340) 5353

bo bo BO bO bo Co WO OO

5490|5502/5514
5611}5623/5635
5729 5740|5752
5843/5855|5866
5955/5966)5977
6064) 6075) 6085
6170/6180/6191
6274/ 6284/6294
6375)/6385|6395
6474/ 6484/6493

5539
5658
5775
5888
5999
6107
6212
6314
6415
6513

5453/5465|5478
5575/5587|5599
5694|5705|5717
5809| 5821/5832
5922/5933|/5944

Www, > eee PP

See et
CLD DAN
NIT I

6031)|6042|6053
6138/6149/6160
6243/6253/6263
6345) 6355/6365
6444/6454) 6464

a
Co 0) 09 09 OO

6609
6702
6794
6884
6972
7059
7143

6580)6590
6675/6684
6767|6776
6857|6866
6946/6955
7033|7042
7118|7126

6542|6551/6561
6637| 6646/6656
6730|6739|6749
6821/6830|/6839
6911/6920/6928
6998) 7007/7016
7084/7093'7101

ee
Drrwryrmp nwwnwpnywrn
09 09 Oo Oo 09

7168/7177
7251/7259
7332/7340

7185
7267
7348

7202
7284

7364

7210
7292
7372

7226
7308

7388

|
bo bo bo bo bo

C2 2 2 >

~I ~I ~1 00 00

186

AIR COMPRESSION AND TRANSMISSION

1 2

7419
7497
7574
7649
7723

7796
7868
7938
8000/8007
8069) 8075
8136/8142
8202/8209
8267|8274
8331)8338
8395)/8401

7412
7490
7566
7642
7716

1789
7860
7931

8463
8525
8585
8639/8645
8698/8704
8756/8762
8814/8820
8871'8876
8927/8932
8982/8987

8457
8519
8579

9042
9096

9036
9090
9143/9149
9196/9201
9248/9253
9299/9304
9350/9355
9400/9405
9450/9455
9499/9504

9552
9600
9647

9547
9595
9643
9689/9694
9736/9741
9782)9786
9827/9832
9872/9877
9917/9921
9961/9965

7427
7505
7582
7657
7731

7435
7513
7589
7664
7738

7810
7882
7952
8021
8089
8156
8222
8287
8351
8414

8476
8537
8597
8657
8716
8774
8831
8887
8943
8998

9053
9106
9159
9212
9263
9315
9365
9415
9465
9513

9562
9609
9657
9703
9750
9795
9841
9886
9930
9974

LOGARITHMS.

5 6 7 8 9
7443)7451 || 7459|7466)7474
7520|7528 || 7536|7543|7551
7597|7604 || 7612/7619)7627
7672)|7679 || 7686/7694|7701
7745)7752 || 7760|7767|7774
7818|7825 || 7832|7839)7846
7889|7896 || 7903/7910)\7917
7959|}7966 || 7973/7980) 7987
8028/8035 || 8041/8048/8055
8096/8102 || 8109)8116)8122
8162/8169 || 8176)8182)8189
8228/8235 || 8241/8248|/8254
8293/8299 || 8306/8312/8319
8357/8363 || 8370/8376|8382
8420/8426 || 8432)8439)8445
8482/8488 |) 8494/8500/8506
8543/8549 || 8555/8561/8567
8603/8609 || 8615)8621|8627
8663/8669 || 8675/8681/86386
8722/8727 || 8733/8739|8745
8779/8785 || 8791/8797/8802
8837/8842 || 8848/8854/8859
8893 8899 || 8904/8910)8915
8949/8954 || 8960/8965/8971
9004)9009 || 9015)9020)9025
9058/9063 || 9069/9074/9079
9112/9117 || 9122)9128/9133
9165)9170 || 9175)9180/9186
9217/9222 || 9227/9232/9238
9269/9274 || 9279)9284/9289
9320/9325 || 9330/9335/9340
9370/9375 || 9380/9385/9390
9420/9425 || 9430)9435/9440
9469/9474 || 9479)/9484/9489
9518/9523 |) 9528/9533/9538
9566/9571 || 9576|9581/9586
9614/9619 || 9624)/9628)9633
9661/9666 || 9671}9675|9680
9708/9713 || 9717|/9722/9727
9754/9759 || 9763|9768/9773
9800)9805 || 9809/9814/9818
9845/9850 || 9854)9859)/9863
9890/9894 || 9899}/9903)9908
9934/9939 || 9943)/9948)/9952
9978/9983 || 9987/9991/9996

1 2534) 4. 6 (Gave Ss

et et
— bo bo bO
bo bo bo bo bo bo bo bo bo bo

— a et ee

pt et ot

ee ee ee

wmnodnhynwlrynmynnmnyn NWwwwwl wowwwow wowweo

— at pt et

— a pe pp pe

SSS oCcoorrF

co eo CO CO
a aoe

'
tt pt pt
mer bo bo

See ep See et tt et ht tt ee
bo bo bo bo bo bo bo bO bO bo bo bo bo bo bo

bo bo bo bo bo

Oe ee a eG

Proportional Parts.

ee’

ee ee

bo bo bo bo bo

bo bo bo bo bo bo bo bo bo bo bo bo bo bo bo

ww We PP PP

Cow WO 0 Co Go GO GO

ww 0 0 Qo Oo OO GO OO

bo bo bo Go CO

bo bo bo bb bo bo bo bo bd bo

[antl alls lls ls EPP PP PRR RR oon

Oo GO GOD

wwwww ew O93 0) OO

WwWwwww

eo) 09 Co WO Ww

a WwWwwe He ee He ee He ee Ree eR PR OL OT or Ovr9or1ror1 or or Or Orv or or or

Or Or Or Or Or Or Ot Ot Or Od

Co tg cl a ls ll Leal als eels oll sd PPP LP Pp He He HR OT Or Or Or Or Or Or Or Or Ot Or Or Or Otic & DD 2 2

PORWOMOD Wsaivwsy

OD DD OD

eee PP He Ee BR Or Or Or Ot Or Or Or

CO ee

PP PP LP

APPENDIX A 187

The log of 0.483°°*? is found by multiplying the log by the exponent
or 0.42 (9.6839—10) which is 4.067238—4.2. It is difficult to get
the antilog of this directly, but the value of the logarithm is not
changed if a number be added to the first part and subtracted from
the second part to make this —1o. In this case add 5.8 to the first
and substract 5.8 from the second, making the log of 0.4839”
9.8672 —TIo.

The antilog of this is 0.736 plus 0.0004 or 0.7364.

That is, 0.483°°4? is equal to 0.7364.

APPENDIX B

Naperian Logarithms.—The natural, hyperbolic or naperian
logarithm of a number can be found by multiplying the common log-
arithm of the number by 2.3026 but the solution of problems involv-
ing this log or the loge as it is written will be facilitated by the
use of the following tables which read from 1 to 10 by increments
of hundredths.

For example, the loge of 4.36 is given directly as 1.4725.

Characteristics and mantissas are not handled in this table in the
same way as the common logs. But as the log of 43.6 is the same as
the log. of 4.36 X10 this may be found by adding the logse of 4.36
and to. In this case this is the sum of 1.4725 and 2.3026 or 3.7151.
That is, the loge of 43.6 is 3.7151. .

In the same way the loge of .436 is the same as the loge of (4.36
divided by 10) or the loge of 4.36 minus the loge of 10. In this
case it 1S 1.4725—2.3026 or —o.8301. That is the loge of 0.436
is —0.8301, a negative number. }

188

PEAY DIX, B ; 189

€ = 2.7182818 log e = 0.4342945 = M
0 1 2 3 4 5 6 7 8 9
1.0 {0.0000 |0.00995/0.01980,0. 02956/0. 03922\0. 04879|0. 05827/0. 06766|0. 07696/0. 08618
1.1 /0.09531)0.1044 )0.1133 |0.1222 |0.1310 |0.1398 |0.1484 |0.1570 |0.1655 |0.1739
1.2 |0.1823 |0.1906 |0.1988 |0.2070 |0.2151 |0.2231 |0.2311 |0. 2390 |0.2469 |0. 2546
1.3 |0.2624 |0.2700 |0.2776 /0.2852 |0.2927 |0.3001 |0.3075 |0.3148 |0.3221 |0.3293
1.4 |0.3365 |0.3436 |0.3507 |0.3577 {0.3646 |0.3716 |0.3784 |0.3853 » 720 |0.3988
1.5 0.4055 |0.4121 |0.4187 |0.4253 |0.4318 |0.4382 |0.4447 |0.4511 {0.4574 |0.4637
1.6 |0.4700 |0.4762 |0.4824 |0.4886 |0.4947 /0.5008 {0.5068 [0.5128 |0.5188 |0.5247
1.7 |0.5306 |0.5365 |0.5423 |0.5481 /0.5539 /0.5596 |0.5653 |0.5710 |0.5766 |0.5822
1.8 |0.5878 |0.5933 |0.5988 |0.6043 |0.6098 |0.6152 /0.6206 /0.6259 |0.6313 |0. 6366
1.9 |0.6418 |0.6471 |0.6523 [0.6575 |0.6627 |0.6678 /0.6729 |0.6780 |0.6831 |0.6881
2.0 |0.6931 |0.6981 |0.7031 |0.7080 |0.7129 |0.7178 |0.7227 |0.7275 |0.7324 |0.7372
2.1 (0.7419 |0.7467 |0.7514 |0.7561 |0.7608 |0.7655 |0.7701 |0.7747 |0.7793 |0. 7839
2.2 (0.7884 |0.7930 |0.7975 |0.8020 |0.8065 |0.8109 |0.8154 |0.8198 /0.8242 |0.8286
2.3 |0.8329 |0.8372 |0.8416 |0.8459 |0.8502 |0.8544 |0.8587 |0.8629 /0.8671 |0.8713
2.4 |0.8755 |0.8796 |0.8838 |0.8879 /0.8920 |0.8961 /0.9002 |0.9042 |0.9083 |0.9123
2.5 |0.9163 |0.9203 |0.9243 |0.9282 |0.9322 |0.9361 |0.9400 |0.9439 |0.9478 |0.9517
2.6 |0.9555 |0.9594 |0.9632 0.9670 |0.9708 |0.9746 |0.9783 |0.9821 |0.9858 |0.9895
2.7 |0.9933 |0.9969 |1.0006 /1.0043 |1.0080 {1.0116 |1.0152 |1.0188 /1.0225 |1.0260
2.8 |1.0296 {1.0332 |1.0367 |1.0403 |1.0438 |1.0473 |1.0508 [1.0543 |1.0578 |1.0613
2.9 {1.0647 |1.0682 |1.0716 {1.0750 |1.0784 |1.0818 |1.0852 |1.0886 |1.0919 [1.0953
3.0 {1.0986 |1.1019 (1.1053 |1.1086 )1.1119 |1.1151 {1.1184 |1.1217 {1.1249 /1.1282
13.1 (1.1314 |1.1346 )1.1378 |1.1410 |1.1442 |1.1474 )1.1506 1.1537 |1.1569 /1.1600
3.2 /1.1632 /1.1663 |1.1694 |1.1725 |1.1756 {1.1787 |1.1817 |1.1848 |1.1878 |1.1909
3.3 {1.1939 |1.1969 |1.2000 {1.2030 |1.2060 {1.2090 [1.2119 |1.2149 {1.2179 |1.2208
3.4 {1.2238 |1.2267 |1.2296 |1.2326 {1.2355 |1.2384 [1.2413 |1.2442 (1.2470 |1. 2499
3.5 /1.2528 |1.2556 |1.2585 |1.2613 |1.2641 {1.2669 |1.2698 |1.2726 |1.2754 |1.2782
3.6 [1.2809 |1.2837 |1.2865 |1.2892 |1.2920 {1.2947 |1.2975 |1.3002 {1.3029 /1.3056
3.7 {1.3083 |1.3110 |1.3137 [1.3164 |1.3191 {1.3218 |1.3244 |1.3271 |1.3297 |1.3324
3.8 {1.3350 |1.3376 |1.3403 |1.3429 |1.3455 |1.3481 |1.3507 |1.3533 |1.3558 |1.3584
3.9 {1.3610 /1.3635 |1.3661 |1.3686 |1.3712 |1.3737 |1.3762 |1.3788 |1.3813 |1. 3838
4.0 {1.3863 |1.3888 {1.3913 |1.3938 /1.3962 |1.3987 |1.4012 |1.4036 |1.4061 |1. 4085
4.1 {1.4110 |1.4134 |1.4159 |1.4183 |1.4207 |1.4231 |1.4255 |1.4279 |1. 4303 |1.4327
4.2 |1.4351 |1.4375 |1.4398 |1.4422 |1.4446 |1.4469 [1.4493 |1.4516 /1. 4540 |1.4563
4.3 |1.4586 |1.4609 |1.4633 |1.4656 |1.4679 /1.4702 |1.4725 |1.4748 |1.4770 |1.4793
4.4 |1.4816 |1.4839 |1.4861 |1.4884 |1.4907 |1.4929 |1.4951 |1.4974 |1.4996 |1.5019
4.5 |1.5041 |1.5063 |1.5085 |1.5107 {1.5129 |1.5151 |1.5173 [1.5195 [1.5217 [1.5239
4.6 |1.5261 |1.5282 |1.5304 |1.5326 |1.5347 |1.5369 |1.5390 |1.5412 |1.5433 |1.5454
4,7 |1.5476 |1.5497 |1.5518 |1.5539 |1.5560 /1.5581 |1.5602 |1.5623 |1.5644 /1.5665
4.8 |1.5686 |1.5707 |1.5728 |1.5748 |1.5769 {1.5790 {1.5810 1.5831 (1.5851 |1.5872
4.9 1.5892 |1.5913 |1.5933 |1.5953 {1.5974 |1.5994 {1.6014 |1.6034 |1.6054 |1.6074
5.0 {1.6094 [1.6114 |1.6134 |1.6154 |1.6174 {1.6194 |1.6214 |1.6233 (1.6253 |1.6273
5.1 |1.6292 |1.6312 |1.6332 |1.6351 |1.6371 |1.6390 |1.6409 |1.6429 |1.6448 |1. 6467
5.2 |1.6487 |1.6506 |1.6525 |1.6544 |1.6563 /1.6582 |1.6601 |1.6620 {1.6639 |1.6658
5.3 |1.6677 |1.6696 |1.6715 |1.6734 |1.6752 |1.6771 |1.6790 |1.6808 |1.6827 {1.6845
5.4 |1.6864 |1.6882 |1.6901 |1.6919 |1.6938 |1.6956 |1.6974 |1.6993 |1. 7011 |1.7029
5.5 |1.7047 |1.7066 |1.7884 |1.7102 |1.7120 /1.7138 [1.7156 {1.7174 |1.7192 |1.7210
5.6 (1.7228 |1.7246 |1.7263 |1.7281 |1.7299 |1.7317 [1.7334 [1.7352 |1.7370 |1. 7387

190 AIR COMPRESSION AND TRANSMISSION

NAPERIAN LOGARITHMS.

eS et pe ——— a =" — pat pe
ai hee! ve eee 8. . or ceils

bO bO bO bo bo bo bo bo bo bo et _— — eee — eee —oe — aay

bo bo bo bO bo bo bo DO bo bo bo bo bo bo bo bo bo eet oe Rt pet ——a— — ee — ee

Nowy bobo bo bobo bo bo DNMp MMH MHP WH WH HYD Bee Se Eee

Ddb MNMN Whwrp WO HWWHY HWwry

APPENDIX C

HYGROMETRY'!

Hygrometry is the measurement of the amount of water vapor in
the atmosphere. There is always more or less water vapor in the
atmosphere depending onits temperature and its degree of saturation.
The study of hygrometry is of increasing importance. It has been
found by experience that the moisture in air has a marked effect on
many industrial processes, suchas the spinning of cotton, the smelting
of iron in blast furnaces and the ventilation of factories and other
buildings. It isalso necessary to know the amount of moisture pres-
ent in all measurement of air or gases and in tests of machin-
ery for handling the same.

According to Dalton’s law, when a mixture of two gases fills a space
of, say, 1 cu. ft., the pressure in the space is the sum of the two pres-
sures that would be produced by a cubic foot of each of the gases alone
at the same temperature. In the same manner a mixture of air and
vapor hasa pressure which is the sum of the pressure of an equal vol-
ume of dry air, and of vapor alone, each at the given temperature
of the mixture. Air and vapor occur in mixtures varying from prac-
tically dry air to a state of saturation such that any addition to the
mixture of vapor at the same temperature causes a portion to con-
dense. To every temperature there corresponds a certain water-
vapor pressure or partial pressure which may be found in steam
tables such as ‘“‘ Marks and Davis” or ‘‘ Peabody’s.”’

Air in actual practice rarely contains vapor with 1oo per cent. sat-
uration and the weight of water vapor present is less than the maxi-
mum for that temperature of air. The air is then said to be only
partially saturated, and the degree of saturation is expressed by the
ratio of the weight of water vapor actually contained in a given space
to the maximum weight that the space can contain under the condi-
tions of absolute pressures and temperatures existing at that time.
This ratio is known as the “Relative Humidity.”

Absolute Humidity.—The absolute humidity is the weight of water

1 Christie’s and Kowalke’s Steam and Gas Engineering Laboratory Notes.
191

192 AIR COMPRESSION AND TRANSMISSION

vapor that 1 cu. ft. actually contains under the given pressure and
temperature conditions.

Relative Humidity.—Relative humidity is usually determined by
means of psychrometers or wet -and dry-bulb thermometers. These
consist of two thermometers fastened to a frame and placed in a
current of air. The bulb of one thermometer is kept covered with
cotton wick and is kept thoroughly wet with water at room tempera-
ture. Ifthe airis not saturated, evaporation will take place from the
wet bulb and its temperature will be lowered by the abstraction of
the latent heat of the water. This lowering of the temperature
has been found to be a measure of the relative humidity.

Psychrometers.—Psychrometers are made in two types, stationary
and sling. In the sling psychrometer the wick is moistened and the
whole frame whirled around by a handle for 15 or 20 seconds. The
wet bulb thermometer is read immediately after stopping. By the
use of Chart A of the accompanying diagram the relative humidity
may be obtained from the readings of the two thermometers. For
example, if the dry-bulb thermometer shows a reading of 72° and the
wet bulb 61°, or a difference of 11°, the relative humidity is 52.6 per
cent., if the atmospheric pressure is 30 in. of mercury.. Ifthe atmos-
pheric pressure is 28 in. the chart reading should be increased 2X 1/100
of 52.6 or 1.05, making the corrected relative humidity 53.65 per cent.

In making accurate measurements of air it is necessary to deter-
mine carefully the weight of moisture present in the atmosphere and
the volume occupied by this vapor. In order to do this, reference is
made to data regarding the density of vapor at various temperatures
and pressures. This information is given in most steam tables and
the following figures have been taken from Marks and Davis tables.
The density or weight of the vapor per cubic foot isshown graphically
in Chart C of the large diagram as the line marked 100 per cent.
This same chart also shows weights of the vapor to be used in calcu-
lations with air of various humidities, as shown in the example fol-
lowing’ the tables.

— ee ee

“= Le ee
SS te — < “4 : ‘ Te

Bs we 2 7 =

een ae

mba

\
. >
pet aren

~~) es mee

5 ae

ee a

ahd pes myc ge: «

ie, i eal: EE WE

Sree

Rony ons

i
-

we
*
4

4
,

} c
« oe

‘

}

oes

‘

shed

1
$
5,
}

‘
,
J

ee ri
!
i
ed?” be
th ee
r
te te ogee
She a

- nes a

Snes eae! batcs
4
1
- Ne
wh
ert earl
, “to”

“iS

2 tg eg the

: ‘
| mbes wauecs Saaees ta, See eceee eae
ng ;

ae ep

ARs

ee ; Ce BNA Ol oss eS

“hd ?

“y ronment 4 ee, «=,

Pea |

AK
¥

~
- ye be ’ * :
- 4 ys 2 , ha . oT a's
> S id =e w ts ee iS a- 3. ~*~ Py a *
~ ’ ¥ 5 . ASA, . f
. : ?
ra . > ;

é
is

sv
.
rT ae,

-
Line OF
obras!

< as
+32 § ape IN ie ner ; no eM ie Ut Oh OSes ey Sis se eS
a & aes im 2% > ‘ .*
a" o> fa: x 4 Er “t-i-* 4 S > ; '
“Fy dee he ee res ~ ¥ a - % - oe a + Anes —
ie. waice OL , ‘ ; 4 S . eH ' r ' : i ¥
*% Oe ela She Se PT : _ ee : bigs ek Poe be) ee
aK % al i er Os: es CSAP Ly anes Se SRS CRN ete er Cone =
= fs . ‘ i , t I oo
oer BEN, t en — SOE ee a Sy ere 1
SESS 2 Soto eae
a ss = = “ e = \ ~ aa a
SSFP a) SIN 1) sear WES COA) SEY SN NNER a Saeco (OOS ESET eS Re Boe a Ee

‘

Temperature of Air or Dry Bulb Thermometer, Degrees Fahrenheit
coy o fo} o So

100 F = = = =:

5 |
= >! pe

= 2

= 3
mod

2 1 4
=

ry 5
S Fe

”

= 7
3

E a
5 aS
&S = v
eer 0 = > §
Pe =<

xk = 10-5
of =
Ee ~v nea
oS —
ae ~ >
3 Fy es
+ Dil

5 eze : 88
EES z
pee bans = = 4 5
o gt =< 2
+fo = cy
© Sts = Ibi
ae =. ©
aS Pte les
£ 3g 50 Ss
oo? ~ 7 3
eg = =
Ss -
-o¢ 18 @
PA See A a
Les 19
mo] rT)
E2555 a:
282" 71 Ps "s
rae y~ (er
oo = o
= 230
ees 1 = al >
Ena So ~ 4 8
-—-> - 3

Po aa 33
at = 134

(=) =

eee tes eae 35
&z> ~F-L17 36
Ea ld e (oa 3
ow 38
=k f

S 39

Yo Oo

Lo ss 40
ce 4l
ov ;’ 42
vo

ge Bi
-=3 0

CHART A

rey re ° 2
is) Ss Ss So So is
105 of, 6
is F : | | TTT y | ourune | | [ Lop
; | OUTLINE OF CALCULATIONS ;
Specific Heat of Air ot 14.5 Lb. per Sq. Ih: + Weight of Water Vapor in Pounds} per Cubic!Foot+ = 2 2 ae. ae + i al Teoh Zz
P 7B. TU; perPound PSb-oe Ieatcvintea from Valves givenin Marke a Dowie Steam Tables i i : 4 das roe ees ieee al a, hu
oO w uw ‘a = : . i e A
Bal ae | lies i Ss S S S | S 3 By Sturated vapor pressure in lbs. per sq. in.(Marks &Davis Stearn Tables used) | we 6 lo or AS? ASD Ay AS 40% 0 fo coe or
S = B ° = Ss S Ss Ss Ss xP r Cent Hurniality | | | | | hae 00045
100 = f <Teynperature of Air in degrees Fahrenheit fel alee 12 =
W= Weight of one cu.ft: of a mixture of alrand water vapor at t-termp, press. ana xR humidity 7-s 2g
3 = Wa- Weight of one cu ft of dry airata pressure of ('40-xp,) las. per Sq-171- t$—}— S = 48
S We = Correction tobe added to ve for pressure above 40 /bs. per sq.in. | = + Lit ce
= ely 2 Jest | Wy Weight of water vapor contained in ane cu.ft of saturated air | | z
ion S\3 eS S\S Ss S88 = S =Specific heat ofa mixture of air and water vapor(B.tu//b. ) a =f Le
S Man Vink Wa a Sa Smecitic heat of dry alr=Q242+0.000009 t: Harvey M Davis, FarnsASME, Vol50,e 750/908) 3 r B 0040 &
Sw-Spyecitic heat of water vapor-04423+0 00018 WilisH Carrier Jour ASME, Vol33 2328/91) Ss = fu
= R-SRIS | | fe 3 it £
T 459.644 —| | te 2 a
ifeel Fo Dry Alc: + 25 - => Se] a Ea
1 Pala |RT 5 Wee pei W= gt Worx My pe 2 — S ie 33
ge GED AOA oud) 3 8 0.0035 ++ <
90 F5S5|4596+7 ilies! $ [2 =
+ =i =\5 ~—t oO a
nee 448 14.0) x £ o£.
(OB S5\45S9.6 +4) - | a haa a ion ib s 3
L aoa Wo + We) Ser) + Wy Sw) a S + S fe S s
= Way? WeFX 3 Ss
ie i+ ‘cal at VE PX My, Ss >
p + = 0.0030 i 2
£ 85 3 a alae 2
& 3 S ¢ > £
4 i 5 S = co 4° ae
8 2 4 | o wo oe
2 ey a 1 cS 8 tg = ~
+= aH S
S. | Ta r Cor re + x r3 = $
a I SSS AOS ee Se c 0.0025 S18 lst! 3
fe, | LT LY ree |S Tr T || oe s+ rt ‘
= AO = E <> OF ue
S ie | | =a x | ‘a & | AS esc ty Son a > 8
2 | |_| les we ae i 1 fe af e Ss
g = + =H e Fe ii 164 te oe. Re
= & Mo &
Bas a ee Sat ret = a 7 ao020 "4 ‘$ &
: a | : Hts bs Hee EHH Z
£ a Sales = 2 Sy rete re = i
& + | “ 2 ae jee ee 13 RG #3 : sleet &
RG S ~
Rs S| tt Ie Lt + o & a = 5
a {its SLE re mS I &A8 g
70 eS rast i Nes ee 2
cs Heer 1 « AS Lag
is Teo S | es | | ae ~Oy eS | 4
: at = Tl Ss 7 hey * rae Se a e
S 7 y & %S
pee anee fooes Betts HEE
) SEE ee eet Cees . s we S
65 x 8 | - Ly | =e elms} _ &
5S ; x S)
aE Ss 4—S ie ey
[ : t +. ttt eet { :
Ct ae POPE ae a et et 5 ;
= = T ir + + Al | as roe | 3 oa <
= 2 | ©) =
| EEC | aS 7 Tt ;
60 ] leaatssye | ie a CI > = | : L S
« <
a t * es " =
Error in Specific Heat Curves: ; = oe es ee
4100 °F, 100% hurnidity anid pressure of #0 /é or- a at S = e: TI] gs ima iS |
/5.0 |b per sg.in. these vizlves are in errorgof orre %. Bil ie Js iS | [ ze '
At 60°F, 00% humidity arid pressure of 140 /b or + ie [ [ | i © H
18.0 Ib. per sg.in. these values are in errorg ofone % = tn : +—|—+ “tt Ve | et | 2
55 S w ne % .
CHART B CHART C zg : GHART.D * CHART E 8 CHART F g CHART 6
sg s 3 s

Humipiry, WEIGHTS PER CuBIC Foor AND SPECIFIC HEATS oF

W,. C. RowsE, Mapison, Wis.

TIXTURES OF AIR AND WATER VAPOR

= 105

ao
a

s
erature of Air, Degrees Fahrenhe

Temp

APPENDIX, C 193

Tables showing the temperature, pressure, specific volume and density of
steam or water vapor from 32° to 219° F., condensed from Marks’ and Davis’
Steam Tables by permission of the publishers, Longmans, Green & Co.

pO Specific vol., Density,
Temperature :
cea cubic feet pounds per
Pounds per Inches per pound cubic foot
square inch mercury

32 0.0886 0.1804 3,204 ; O.000304
2 2A Be 0.0922 0.1878 ; pL Om ee 19 OC. 0003T0
34 0.0960 0.1955 3,052 0.000328
35 0.0999 0. 2034 2,938 ©.000340
36 0.1040 Oar rn? 2,829 0.000353
27, 0.1081 - 0.2202 ayn 0.000367
38 O.1T25 O.2290 2,626 0.000381
30 Oni170 0.2382 2,530 0.000395
40 OLT227, OV2477 2,438 0.000410
41 0.1265 0.2575 pe Ke. 0.000425
42 0.1315 0.2677 2,266 0.000441
43 0.1366 On 2782 2,185 0.000458
44 0.1420 0. 2890 2,107 0.000475
45 OfT 47s ©. 3002 2,033 0.000492
46 OLDE 32 0.3118 1,961 ©.000510
47 0.1591 0.3238 1,892 ©.000529
48 0.1051 0.3363 1,826 0.000548
49 0.1715 0.3492 1,763 0.000567
50 0.1780 0.3625 1,702 0.000587
51 0.1848 0.3762 1,643 0.000608
52 O.1Q17 0.3903 1,586 0.000630
53 0.1989 ©. 4049 1,532 0.000653
54 0.2063 0.4201 1,480 0.000676
55 0.2140 0.4357 1,430 2.000700
56 0.2219 0.4518 1,381 0.000724
i) 0.2301 0.4684 335 0.000749
58 0.2385 0.4856 1,291 0.000775
59 Os 2472 0.5034 1,249 0.000801

194 AIR COMPRESSION AND TRANSMISSION

Pate Specific vol., Density,
Temperature :
re tometer cubic feet pounds per
Pounds per Inches per pound cubic foot
square inch mercury
60 0.2562 0.522 1,208 0.000828
61 O2054 0.541 1,168 0.000856
62 0.2749 0.560 1,130 0.000885
63 0.2847 0.580 1,093 ©.O000Q15
64 ©. 29049 0.601 1,058 0.000946
65 0.3054 oO. 022 1,024 0.000977
66 0.3161 0.644 ggI ©.O001009
67 0. 3272 0.667 959 ©.001043
68 0.3386 0.690 928 O.001077
69 0.3504 On7iA 899 O.OO111I2
70 0.3626 0.739 871 0.001148
71 Bgat 0.764 843 0.001186
7 0. 3880 0.790 817 O.001224
Fes O.4012 0.817 792 0.001263
74 0.4148 0.845 767 ©.001304
ws 0.4288 0.873 743 0.001346
76 0.4432 0.903 720 0.001389
ot 0.4581 0.03% 698 0.001433
78 0.4735 0.964 677 0.001477
79 0.4893 0.996 657 0.001523
80 0.505 I.029 636.8 ©.001570
8I O15 22 ip O0s O17 5 0.001619
82 0.539 1.098 598.7 0.001670
83 OL587 TLi134 580.5 O.001723
84 OVS75 ie 7a 562.9 O20017 77
85 0.594 I. 209 545.9 0.001832
86 O01? 1.248 526:.,5 0.001889
87 0.633 1.289 Bay ©.001947
88 0.654 1 B31 498.4 ©.002007
89 C2075 Daas 483.6 0.002068
go 0.696 TgA17 469.3 O.002131
QI °.718 1.462 455-5 0.002195
92 OL7AT 1.508 442.2 0.002261
93 0.765 T2556 429.4 0.002320
04 0.789 1.605 417.0 0.002398

APPENDIX C 195
Pressure . :
Specific vol., Density,
Temperature 2
Fahrenheit cubic feet pounds per
Pounds per Inches per pound cubic foot
square inch mercury

95 0.813 1.655 405.0 0.002469
96 0.838 1.706 303-4 0.002542
97 0.864 Le 7S0 20242 0.002617
98 0.891 1,813 27 LMA 0.002693
99 0.918 1.869 360.9 0.002771
100 0.946 1.926 350.8 0.002851
IOI 0.975 1.985 GAO 0.002933
102 1.005 2.045 ZZ 1s 0.003017
103 TROZS Siew B2242 0.003104
104 1.066 py hs Sy iew 0.003192
105 1.098 22230 BO4m 7) 0.003282
106 Pers 26303 2960.4 0.003374
107 1.165 gus72 288.3 0.003469
108 I.199 e443 280.5 0.003565
109 re235 2eSLs 27250 0.003664
IIO Lear 2.589 265.5 0.003766
III 1.308 2.665 2522 0.003871
Lis 1.346 2740 oa Ret) 0.003978
by he 14386 2,822 244.7 0.004087
II4 1.426 2.904 238.2 0.004198
II5 1.467 2.987 2270 0.004312
116 I.509 3.072 22573 0.004429
ray 14552 Be ror 219.9 0.004548
118 1.597 2-252 ZrAe TL 0.004671
I1g 1.642 3.344 208.5 0.004796
120 1.689 3.438 203.1 0.004924
I2E 1.736 aoe 197.9 0.005054
22 e705 3.635 192.8 0.005187
123 1.835 By Ta 187.9. 0.005323
124 1.886 3.841 183.1 0.005462
125 1.938 3.948 178.4 0.005605
126 1.992 4.057 17300 0.005751
127 2.047 4.168 169.6 ©.005900
128 21O 4.282 16553 0.006052
129 2.160 4.399 POT SI 0.006207

196 AIR COMPRESSION AND TRANSMISSION

Pena Specific vol., Density,
Temperature ;
ea Nes cubic feet pounds per
Pounds per Inches per pound cubic foot
square inch mercury

130 2270 AasZ TS 7% 0.00637
131 221276 4.64 a5 a2 0.00653
132 2.340 4.76 149.4 0.00669
133 2.403 4.89 145.8 0.00686
134 2.467 5.02 142.2 0.00703
135 2.533 5.16 13027 0.00721
136 2.600 5.29 135.4 0.00739
£37 2.669 5-43 Poza 0.00757
138 R740 5.58 128.9 0.00776
139 2.812 ew ke 125.5 0.00795
140 2.885 5.88 122.3 0.00814
141 2.960 6.03 119.9 0.00834
142 Bk027 6.18 fi721 0.00854
143 Seni as 6.34 114.3 0.00875
144 BOs OF 55 IIr.6 0.00896
145 Bee 77 6.67 109.0 0.00918
146 Be ZOT 6.84 106.5 0.00940
147 3.446 702 104.0 0.00962
148 23:3532 7220 101.6 0.00985
149 3,023 7.38 99.2 0.01008
150 3.714 2A57 96.9 O,O1L0s2
I51 3.809 7 270 04.7 0.01056
152 3.902 7.95 92.6 0.01080
153 3-999 8.14 9005 = O.O1I05
154 4.098 8.34 88.4 O;OEI3I
155 4.199 Gas5 86.4 O.OII57
156 4.303 8.76 84.5 0.01184
r57 4.408 8.98 82.6 O.OI2I1
158 4.515 Q.20 80.7 0.01239
159 4.625 Q.42 78.9 0.01267
160 4.737 9.65 7702 0.01296
161 4.851 9.88 7S uS 0.01325
162 4.967 10.12 fc tates O.01355
163 5.086 10. 36 FORD 0.01386
164 5.208 10.61 70.6 O.OI417

Temperature
Fahrenheit

APPENDIX,C 197
Pressure : ;
Specific vol., Density,
cubic feet pounds per
Pounds per Inches per pound cubic foot
square inch mercury

5-333 10.86 69.1 0.01448
5-460 LLere 67.6 0.01480
5-589 T1245 66.1 O.OI1513
cana LI.65 64.7 0.01546
5.855 II.Q2 63.3 0.01580
5.992 12520 62.0 0.01614
OL 135 12.48 60:7 0.01649
09273 iy | 59.4 0.01685
6.417 13.07 58.1 O-On7 21
6.564 E3437 56.9 0.01758
6.714 E3307 Bey 0.017096
6.867 13.98 54.5 0.01834
i; O23 14.30 53-4 0.01873
Teth2 14.62 cos 0.01912
7-344 14.95 51.2 0.01953
fhe wt 15.29 5O. 15 0.01994
7.68 15403 49.12 0.02036
TEs 15.098 48.12 0.02078
8.02 16.34 47.14 O-O2T2T
8.20 £570 46.18 0.02165
8.38 17.07 Aseas 0.02210
8.57 17.45 44.34 0.02255
8.76 L7os 43.45 0.02301
nOE05 10.22 42.59 0.02348
9.14 18.61 ALTA 0.02396
9.34 19.02 40.91 0.02444
9.54 19.43 40.10 0.02493
9.74 19.83 39.31 0.02544
9.95 20.27 38.54 0.02505
TOL 7 20n7t 37.78 0.02647
10.39 Zia Ds 37.04 0.02700
10.61 21.60 3632 0.02753
10.83 22505 acyo? 0.02807
II.06 PUNE FAnOS 0.02863
71.20 22.99 34.26 0.02919

198 AIR COMPRESSION AND TRANSMISSION

Eres Specific vol., Density,
Temperature ’

Pahreoheit cubic feet pounds per
Pounds per Inches per pound cubic foot
square inch mercury

200 tT ah 2 225A 7 33.60 0.02976
201 11.76 23.95 32.96 0.03034
202 12.0% 24.45 3 2e a3 0.03093
203 12520 24.96 2272 0.03153
204 12555 25.48 Gratz 0.03214
205 T2477 26.00 20,53 0.03276
206 13203 20.53 29.95 0.03339
207 i3430 27.08 29.39 0.03402
208 Tig 554 2763 28.85 0.03466
209 13505 28.19 Baa 0.03531
210 T4512 28.76 27.80 0.03597
Bit 14.41 20.33 27-20 0.03664
Bee I4.70 29.92 26.79 Of02732
213 14.99 202 26.30 0.03802
214 15.29 ar 1s 25202 0.03873

Partial Pressures.—Suppose, for example, the barometers read
29.214 in. of mercury at a temperature of 78° F. Chart F of the
diagram shows that at this temperature 1 in. of mercury corresponds
to a pressure of 0.4889 Ib. per square inch. That is, the barometer
reading of 29.214 in. of mercury corresponds to an absolute pressure
of 14.2827 lb. per square inch. If the air is saturated with moisture
at 78° F., the pressure exerted by this vapor is, as shown from the
tables of Marks and Davis, 0.4735 lb. persquareinch. The pressure
of the dry air present would then be 14.2827—0.4735 or 13.8092
lb. per square inch.

Suppose the psychrometer shows a relative humidity of 40 per
cent. As the vapor pressures are proportional to the absolute
weights, the pressure exerted by the moisture in the air will be
40 per cent. of 0.4735 or 0.1894 lb. per square inch. In this case
the pressure due to the dry air present will be

14.2827 —0.1894 or 14.0933 lb. per square inch.

If it is necessary to find the weight of a cubic foot of this moist

air, this can be found by adding the weight of the cubic foot of dry

ALPEN DIX'G VES})

air at its pressure and temperature to the weight of the vapor
present.

The weight of vapor present is found by multiplying the weight
of a cubic foot of vapor at the given temperature by the relative
humidity. The tables show that 78° F., the weight of a cubic foot
of vapor, is 0.001477. The weight of the vapor present in the
example is 40 per cent. X0.001477 or 0.000501 lb.

The weight of dry air present is found from the formula

BiVa 144 X 14.0933
Sh Sys == - =0.070
53-311 53-3(460+78) Nie

The weight per cubic foot of the air and its accompanying vapor
is

== le

0.000591 +0.070773 =0.071364.

This calculation can be made quite simply by referring to the
various charts of the large diagram. By referring to Chart D it
will be seen that the weight of air at 4o per cent. relative humidity
and 78° F. is .o6992 lb. per cubic foot if the pressure of the atmosphere
is 14 lb. per square inch. In the example given the pressure is
14.2827 lb. per square inch. By referring to Chart E it will be
seen that for the pressure of 14.2827 and temperature of 78° F. a
correction of 0.00144 should be added making the weight per cubic
foot of this mixture

0.06992-+0.00144 or .07136 lb. per cubic foot.

When it is desired to measure air with a Thomas electric meter,
the mean specific heat of the mixture of air and water vapor must
be known. W. H. Carrier in his paper ‘‘ Rational Psychrometric
Formule,” Journal A. S. M. E., Nov., 1911, gives the following
values which represent the results of the more recent investigations
on the specific heat of air and water vapor. Instantaneous specific
heat of air

C pa =0.24112-+0.0000001

where ¢ is the temperature in degrees Fahrenheit; and the instan-
taneous specific heat of water vapor as approximately

C ps =0.4423 +0.0001 8

where ¢ is the temperature in degrees Fahrenheit.
Applying these formule to the example given with temperature
of 78, C pa IS 0.241822 and C ps 1S 0.45634.

200 AIR COMPRESSION AND TRANSMISSION

The mean specific heat can then be found by multiplying the
weight of each substance in the mixture by its specific heat, adding
the products, and dividing the sum by the weight of the mixture.
Thus

For the air, 0.070773 X0.241822 =0.017114
For the moisture, 0.000591 X0.45634 =0.000270

0.017384

Mean specific heat is 0.017384 -+0.071364 or 0.2436.

The mean specific heat may also be obtained by referring to
Chart B of the large diagram. This shows that for the given temp-
erature of 78° F. and a relative humidity of 40 per cent. the mean
specific heat may be taken as 0.2435.

The above principles are applied commercially in testing steam
condensers. An accurate thermometer is placed in the suction
to the dry air pump and a mercury column attached to the same.
In a condenser the conditions are such that the mixture is always
saturated. Hence the pressure due to water vapor passing to the
air pump will equal that due to its temperature as given in the
steam tables. Then the difference between this pressure and that
shown by the mercury column will equal the pressure due to the
dry air in the mixture. If the volumetric efficiency of the air
pump is known, the amount of air pumped can be computed, and
this gives a means of readily checking the condensing equipment
- for air leakage.

The large diagram containing Charts A, B, C, D, E, F and G
was prepared by W. C. Rowse, Instructor in the Steam and Gas
Engineering Department of the University of Wisconsin.

INDEX

Absolute humidity, 191
temperature, 5
zero, 6
Action of piston compressor, 70
Actual card of piston compressor, 78
compression, 75
Advantage of isothermal compressor,
25
of multi-stage compressor, 90
AIT UE
at low pressures, 38, 68
at pressures below the atmosphere,
26,—68
composition, 1
characteristics, 1-4
and energy equations,
10-17
compressor cards, 75
discharge valve, 102
density at various pressures, 174
dry, 4
for cupolas, 39
for forges, 39
for ventilation, 39, 40
free, 2
humidity, 2-6
internal energy, 6, 7, 16
in water, 29
inlet valve, ror
measurement, 160-171
pump, Edwards, 31
U.S. Navy, 30
supply for various buildings and
rooms, 40
Allis Chalmers fan, 65
Altitude effect, 140-144
Anemometers, 43
Apparatus for measuring large quan-
tities of air, 166
Apparent specific heat, 8
volumetric efficiency, 77
Area of inlet valves, 100
of discharge valves, 1or
of fan blast, 43, 62

Arrangements for coupling turbo-
blowers, 125

Arthur compressor, 132

Arthur, Thomas, 132

Automatic valves, 100

Available power, 179

Axial discharge fan, 41

thrust, balancing, 121

Balancing axial thrust, 121
Rateau impellers, 122
by balancing piston, 123
by counter position, 121
by diminishing back area, 122
Baloche and Krahnass compressor,
eWay Bed
Belt regulator, 105, 107
Blast area, fans, 43, 62
Blower capacities, 50
cross section, 50
definitions, 42
efficiency, 81
losses, 81
mixing, 127
Parsons, 114
pressures, 50
Rateau, 114
Blowing engine, 41
Blowers, 41
Boyle’s law, 10
Brake horse-power for fans, 58, 60, 66
Brauer’s method of constructing ex-
ponential curves, 19
Brown, Boveri and Co. turbo-com-
pressor, 117
British thermal unit, 6
Buildings, air required, 40

Calculated and actual horse-power
required for single stage com-
pression, 74

Capacity of blowers, 50

of fans, 42
of intercoolers, 93, 94

201

202

Capacity of receivers, 160
Card of piston compressor, actual, 78
ideal, 77
Cards, combined two-stage, 147
clearance unloader, 112
from air compressors, 70, 75
showing adiabatic and isother-
mal compression, 73
Carrier, W. H., 199
Centrifugal fans, 38-65
Channing, J. Parke, 144
Characteristic and energy equations
fOtedit, 20-87
equation for perfect gas,
IO
Characteristics of air, 1-4
Christie, A. G., 101
Classification of fans and blowers, 41
of valves, 98
Clayton governor, I09
Cleaning valves, 182
Clearance effect, 70, 71, 96, 97, 990
methods of reducing, 71
unloader, 110, 112
> Cards 142
Coefficient of contraction, 43
of efflux, 43, 56
of velocity, 43
Combined cards, two-stage compressor,
147
governor and regulator, 109
Common logarithms, 184-186
Comparative effect of altitude on out-
put, 143
Compensator, hydraulic, 83
lever, 83, 84
weight, 83
Composition of air, 1
Compressed air explosions, 182
Compression, actual, 75
isothermal, 25
line, 73
wet and dry, 74
exponential, 23
Compressor, direct-acting steam, 82
low pressure, 38
tests, 144, 158
Computation of internal or intrinsic
energy. 16

INDEX

Concentration of liquors, 34
Condenser pumps, 27
Cone wheel fans, 65, 66
Constants for pipe formule, 174,175
Construction of equilateral hyperbola,
18, 19

of exponential curves, 19

of isothermal curves, 18
Contraction, coefficient of, 43
Cooling capacity, 93

devices, 117

surface, 93

turbo-compressors, II5
Cost of Taylor compressor at Ains-_

worth, B. C., 136

Coupling compressors, 124
Cross-section, standard blower, 50

piston compressor, 69
Cupolas, air required, 39
Cutler-Hammer Co., 161
Cylinder efficiency, 80

D’Auria system of energy compensa-
tion, 83
Dalton’s law, 191
Davis, G. J., 164
Definitions, fundamental, 5—9
for fans and blowers, 42
Density of air for various pressures, 174
of water vapor, 193-108
Description of fans, 58
Design of fans, 58, 67
of turbo-compressors, 113
Details of piston air compressors, 98-
110
Developed section of Parsons blades,115
Devices, cooling, 117
Diagram, three stage piston compres-
sor, 116
turbo-compressor, 116
Diagrammatic sketch of Thomas elec-
tric meter, 169
Diagrams, graphical, 18-25
Difference between isothermal and
adiabatic compression, 22
Direct acting steam compressor, 82
Disc fan, 58
Discharge from a fan, 57, 59, 66
valve, 102

INDEX

Discharge, area, or
Draft measurement, 43
Dresser coupler, 172
Dry air, 4
pump, 27
Duplex compressor, 86
cross compound steam, two-stage
air compressor, 88
belt driven compressor, 87
steam driven compressor, 87
Durleys Ref. 266

Economic efficiency, 81
Edwards air pump, 31
Effect of altitude, 140-144
of clearance, 70, 71, 96, 97
of changing discharge pressure, 99
of early closing of inlet valve, 73
of pressure on temperature, 4
Effects of heat, 6
of outlet on capacity, 55

Effects of pressure on tempera-
ture A
Efficiency, apparent volumetric, 77
blower, 81

cylinder, 80

economic, 81

of compression, 80

of fans, 45

of Taylor compressor, 134
Efficiencies, 77-82

true volumetric, 80
Efflux, coefficient of, 43, 56
Electric meter, diagram, 169
Energy, 5

compensation, 82-88

in air, 6
Engineering Magazine, 113
Equalizing steam pressure and air

resistance, 82

Equilateral hyperbola, 18, 19
External energy changes, 6
Expansion of casing, 118
Explosions, compressed air, 182
Exponential compression, 23

curve construction, 19

Fan, blast or steel plate, 60
capacity, 42

203

Fan, centrifugal, 38-65
cone wheel, 65-66
definitions, 42
design, 58-67
description, 58
discharge, 58, 59, 66
efficiency, 45
losses, 45
mechanics of, 52
pressure, 42
proportions, 41, 61
radial wheel, 58
speed, 62, 67
Fans, axial, 41
classification of, 41
or blowers, 41
Flow of gas through an orifice, 45, 46
Forges, air required, 39
Forms of poppet valves, ror
Free air, 2
discharge, 42
Friction effect of elbows, 61, 176
Frigells |5.P:.120
Frizell’s compressor, 129
Fundamental definitions, 5-9

Gases in air, I
Governor and regulator combined, 109
Clayton, 109
for electric driven compressors,
107
Nordberg, 109, 110
Grains, vapor per cu. ft. saturated air, 2
Graphical construction of exponential
curve, 18, 19
of isothermal curve, 18, 19
diagrams, 18-25
method of determining
head, 165

mean

Halsey, F. A., 142
Hammon coupler, 172, 173
Heat. 5
added or taken away for iso-
thermal change, 21
for exponential change, 21
etrects.6
taken away during compression,
22

204

Hero’s device for opening temple doors,
VII
fountain, VII
Horse-power, brake for fans, 58, 60, 66
single-stage compression, 74
Horizontal-vertical arrangement of
cylinders, 86
Housing for fans, 42-63
Humidity, absolute, ror
OL aipmtone 1s &
Hydraulic air compression, 129-139
air pump, 26 .
compensator, 83° *
compression losses, 138
compressor, Arthur’s, 132

Baloche and Krahnass, 131, 132.

Taylor’s, 133-137
Hygrometry, 191

Ideal card, piston compressor, 77
Impellers, rotary blowers, 49, 50
Improved cooling, turbo-compressors,
118
Indicator card piston compressor, 70
cards, condenser pumps, 30
Industrial uses vacuum, 32
Ingersoll Rand Co., 103, 111, 112
compressor, 147
Inlet connection, 183
for blowing fan, 61
for exhaust fan, 61
valve, Ior
area, 100
setting, Ior
Intercoolers, 90
capacity, 93
Nordberg, 92
pressure, 93
surface required, 93
types, 92
tubes, 92
with separator, 92
Internal energy changes, 6
or intrinsic energy of air, 7
computation of, 16

acver. (5 Heerco
Jaeger’s turbo-blower, 119
patent impeller, 120

INDEX

Kennedy blowing engine valve, 105
Kowalke, O. L., 191
Krahnass, A., 131

Labyrinth bushing, 120
Law, Boyle’s, 10
of Charles, ro
Leakage past turbo-stages, 120
Lecture by H. deB. Parsons on fans,
41-68
Lever compensation, 83, 84
Leyner air reheater, 177
Liquors, concentration of, 34
Logarithms, common, 184-186
Naperian, 188-190
Loss of capacity due to clearance, 79
of head due to friction in ducts, 47
Losses of blower, 81
of hydraulic compression, 138
Low pressures, compressors, 38
Lubricating compressors, 182

Marks and Davis condensed steam
tables, 193-198
Measurement of compressed air, 160-
171
of draft, 43
of large quantities of air, 166
Measuring vacuums, 27
Mechanical efficiency, 81
valve of Corliss type, 104
valves, 98
Mechanically operated discharge
valve, I0o
Mechanics of the fan, 52
Mercurial air pump, 26
Meter comparisons, 170
test results, 171
Methods of reducing clearance, 71
Mines and Minerals, 144
Mixing blower, 127
Mode of conducting tests, 147
Modern form of Pitot tube, 162
Moisture precipitated from air, 3
Mt. Cenis tunnel, VIIT
Multi-stage compression, 97
advantages, 90

Naperian logarithms, 188-190

INDEX

Net efficiency, 81
Nordberg compressor test, 144
governor, 10g—110
intercooler, 92
Mfg. Co., 109
Norwalk compressor, 84
regulator, 108
Notation of symbols for fan formule,

47
Numerical value of R, 10

Orifice, flow of gas in, 45, 46
Oxygen in air, I
in hydraulic compressed air, 137

Parsons, H. deB., 41-68
blower, 114
blades, 115
Partial pressures, 198
Peele, Robert, 140
Perfect gas, characteristic equation,
IO
intercooling, 93
Peripheral speed of fans, 62, 67
Phenomena of hydraulic air compres-
sion, 137
Pipe couplers, 172, 173
formule, constants, 174, 175
lines, 171-176
line formule, 173
losses, ducts, 48
Piston, balancing, 123
-balanced turbo-compressor, 122
compression, hydraulic, 72
three-stage diagram, 116
compressor action, 70
cross-section, 69
details, 98-112
compressors, 69-77
controlled by multiplicator, 126
-inlet valve, 102, 108
Pitot tube, 161, 162
Pounds of water precipitated per cu.
ft. cooled air, 3
Power, 5
available, 179
consumed by rotary and piston
compressors, 52
for rotary blowers, 51

205

Pressures, blower, 50
oz. per sq. in. in water head to
inches, 44
used for various stages, 90
water column in inches to oz.
per sq. in., 44
Proper receiver pressure for multi-
stage compression, 96
Propeller fan, 58
Proportions of fans and housing, 41, 61
of rotary blowers, 50
Psychrometers, 192
Pump, dry air condenser, 26
Pumps, condenser, 26-30

R, numerical value, 10
Radial wheel fan, 58
Railway and Engineering Review, 130
Rand Imperial unloader, 111
Rateau blower, 114
multiplicator, 125
turbo-compressor, 128
Ratio of air cylinder to low-pressure
steam cylinder, 29
of air cylinder to volume of con-
densed steam, 29
of port to cylinder area, 100
Real specific heat, 8
Receiver aftercoolers, 159
intercoolers, 92
capacity, 160
Receivers, 159
Regulator, belt, 105, 107
and governor combined, 109
Norwalk, 108

Regulators and unloading devices,
105

Relation between altitude and volume,
141

specific heats, 10

Relations between P, v and T for
adiabatic and _ exponentia
changes, 16

Relative humidity, 192

Restricted discharge, 42

Results of meter tests, 171

of tests, 148
Richards, Frank, 175
Right-angle bend resistance, 49

206

Robinson, S. W., 163

Rotary blowing machines, 49
blowers, proportions, 50

Rooms, air required, 40

Rowse, S. W., 200

Runners, 119

Salt evaporating effects, 32
Sangster, Wm., 39
Schmidt, Henry F., 81
Sectional view of Thomas
meter, 168
Selection of air compressors, 179-182
Semi-mechanical valves, 103
Shape of fan blades, 58, 61, 66, 67
Simple form of Pitot tube, 161
Single-stage compression, horse-power
required, 74
Sirocco double inlet fan, 68
Size and type of compressor, 181
of water and air pumps, 28
Sketch of meters placed tandem for
testing, 170
Sommeiller’s compressor, IX
Southwork blowing engine valve, 104
Specific heat, 7
apparent, and real, 8
at constant pressure, 7
at constant volume, 7
at various pressures
peratures, 8
volume of water vapor, 193-1098
Speed of fans, 58, 62, 67
of turbo-compressors, 113
Sperr, aa Ws E36
Sprengle air pump, 26
St. John’s meter, 166
Standards of measurement, 160
Steam cylinder size, 30
Steel plate fans, 61, 64
Straight line compressor, 84
Stuffing boxes, 123
Suction line, 73
Surface of intercoolers, 93
Sullivan air reheater, 177
Mch) Cota 153
Summary of tests, 157

Syphon, 37
bulk head, 131

electric

and tem-

INDEX

Taylor, Charles H., 133
compressor, 133
efficiency, 134
at Ainsworth, B. C., 135
at Magog, Quebec, 134
at Victoria mine Michigan, 136
Temperature, 5
absolute, 5
Temperatures due to adiabatic com-
pression, 22, 23
Test curves, Jaeger’s turbo-blower, 124
of hydraulic compressor, 136
of plant No. 1, 148-151
of plant No. 2, 151-154
of plant No. 3, 154-156
of plant No. 4, 156-157
Tests, mode of conducting, 147
Thomas, Cy Ci rr60
meter, 168
diagram, 169
Three-quarter housed steel plate fan,
64
Tightness between stages, 120
Towl, Forrest, M., 160, 171
Trompe, 129
True volumetric efficiency, 80
Turbine blast or Sirocco fan, 67
Turbo-blower coupling arrangements,
125
ol2s 000 Cutt. 121
of 140,000 cu. ft. capacity, 128
Turbo-compressor cooling, 115
design, 113
diagram, 116
for mixing air and gas, 128
Jaeger’s, 119
Parson’s, 114
Turbo-compressors, 113-128
Two-stage compressor cards, 147
Types of blading, 68

U. S. Navy pump, 30
Uncovering port to release clearance
pressure, 71
Unloader, clearance, 110, 112
Rand Imperial, 111
Unloading devices, 110
Uses of air at low pressures, 38
Usual velocity in ducts, 47

Vacuum cleaners, 36

concentration of liquors, 34, 35

manufacture of salt, 32
measurement, 27
Valve area, 100
of discharge, 101
gear, 179
in cylinder head, 102
mechanical, 98
poppet, Io1
piston-inlet, 102
setting, 101
semi-mechanical, 103
Valves, area of inlet, 100
automatic, 100
classification, 98
cleaning, 182
Vapor in air, 1
Velocity, coefficient of, 43
of air through ports, ror
meters, 161
Ventilation, air required, 39, 40
Venturi meter, 167
vacuum pump, 26
Volumetric efficiency, 77
apparent, 77
true, 80
meters, 160

Water, air in, 29

INDEX 207

Water-cooled turbo-compressor, 117,
118
Water measurements, hydraulic com-
pressor tests, 137
percipitated from compressed air,
3
present in saturated air, 2, 55
required for intercooler, 94
Webb, Richard, L., 146
Weight compensation, 83
of air, 10
Westinghouse air pump, 85
governor, 106
Wet air pump, 27
displacement meter, 160
and dry compression, 74
Weymouth, Thos. R., 164
Wheeler combined pump, 27
condenser pump, 28
Work, 5
done by a compressor, 23
of adiabatic change, 15
of exponential change, 14
of isothermal change, 12
required to move a volume of gas
56

Zero, absolute, 6
Zur Nedden, Franz, 81, 113

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