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THE WILEY TECHNICAL SERIES

FOR

VOCATIONAL AND INDUSTRIAL SCHOOLS

EDITED BY

JOSEPH M. JAMESON .,

'*• GIRARD COLLEGE

THE WILEY TECHNICAL SERIES

EDITED BY

JOSEPH M. JAMESON

TEXT BOOKS IN
MECHANICS, HEAT AND POWER

NOW READY

Steam Power. By C. F. HIBSHFELD. Professor of
Power Engineering, Sibley College, Cornell Uni-
versity, and T. C. ULBRICHT. via +420 pages,
5J by 8. 228 figures. Cloth, $2.00 net.

Gas Power. By C. F. HIRSHFELD, Professor of Power
Engineering, Sibley College, Cornell University,
and T. C. ULBRICHT, Formerly Instructor, Depart-
ment of Power Engineering, Cornell University.
x+209 pages, 5J by 8, 60 figures. Cloth, $1.25
net.

Heat: A Text Book for Technical and Industrial
Students. By J. A. RANDALL, Pratt Institute.
xiv+331 pages, 5£ by 8, 80 figures. $1.50 net.

Elementary Practical Mechanics. By J. M. JAME-
SON, Girard College, xii+321 pages, 5* by 8,
212 figures. Cloth, $1.50 net.

Continuous and Alternating Current Machinery.

By Professor J. H. MORECROFT, Columbia Univer-
sity. ix+466 pages, 5J by 8, 288 figures. Cloth,
$1.75 net.

Applied Mechanics Laboratory Manual. By J. P.

KOTTKAMP, M.E., Instructor in Steam and Strength
of Materials, Pratt Institute. 8 by 10£. Loose
Leaf. 55 exercises with numerous cuts and tables.
Complete, paper cover, $1.00 net.

Exercises in Mechanics. By J. M. JAMESON, Girard
College, Formerly Pratt Institute. 8 by 10£.
Loose Leaf. 52 exercises. Complete, paper covei ,
85 cents net.

Exercises in Heat. By J. A. RANDALL, Pratt Insti-
tute. 8 by 10$. Loose Leaf. 13 exercises. Com-
plete,.paper cover, 26 cents net.

For full announcement see list following index.

5000 8-1-16

STEAM POWER

C. F. HIRSHFELD, M.M.E.

PROFESSOR OF POWER ENGINEERING, SIBLEY COLLEGE, CORNELL UNIVERSITY

AND

T. C. ULBRICHT, M.E., M.M.E.

FORMERLY INSTRUCTOR, DEPARTMENT OF POWER ENGINEERING, SIBLBY COLLEGE,

CORNELL UNIVERSITY; ASSOCIATE MEMBER AMERICAN
SOCIETY OF MECHANICAL ENGINEERS

FIRST EDITION

FIRST THOUSAND

NEW YORK

JOHN WILEY & SONS, INC.

LONDON: CHAPMAN & HALL, LIMITED

1916

•f

PRESS OF

BRAUNWORTH & CO.

BOOKBINDERS AND PRINTERS

BROOKLYN, N. Y-

PREFACE

THE following pages represent the results of an attempt
to collect in a comparatively small book such parts of the
field of steam power as should be familiar to engineers
whose work does not require that they be conversant with
the more complicated thermodynamic principles considered
in advanced treatments. The experience of the authors
has led them to believe that a book of this sort should give
a correct view-point with regard to the use of heat in the
power plant even though it does not enter deeply into the
theoretical considerations leading up to that view-point;
that it should supply the tools required for the solution of
power plant problems of the common sort ; and that it should
give sufficient description of power plant apparatus to
make the reader fairly familiar with the more common
types.

Mathematical treatment of the subject has been elim-
inated to the greatest possible extent, and anyone familiar
with elementary algebra should be able to understand
readily such equations as it has been deemed necessary to
include.

Brief explanations of physical and chemical concepts
are given in every case in which the text required their use,
so that those who have not studied these subjects, and those
who have but have failed to crystallize and hold the neces-

iii

342913

iv PEEFACE

sary ideas, should have little difficulty in reading the text
understandingly.

It is hoped that the book may prove serviceable as a
text for steam power courses given to civil engineers in the
various colleges and that it may also meet the needs of those
instructing power plant operators in industrial schools.

C. F. H.

T. C. U.

JUNE, 1916.

CONTENTS

CHAPTER I

PAGE

PHYSICAL CONCEPTIONS AND UNITS 1

1. Matter. 2. Energy. 3. Units of Matter and Energy.
4. Work. 5. Mechanical Energy. 6. Heat. 7. Temper-
ature. 8. Measurement of Temperature. 9. The Unit of
Heat Energy. 10. Specific Heat. 11. Quantity of Heat.
12. Work and Power.

CHAPTER II

THE HEAT-POWER PLANT 20

13. The Simple Steam-power Plant. 14. Cycle of Events.
15. Action of Steam in the Cylinder. 16. Hydraulic Analogy.

CHAPTER III

STEAM 27

17. Vapors and Gases. 18. Properties of Steam. 19.
'Generation of Steam or Water Vapor. 20. Heat of Liquid,
q or h. 21. Latent Heat of Vaporization, r or L. 22. Total
Heat of Dry Saturated Steam, X or H. 23. Total Heat of Wet
Steam. 24. Heat of Superheat. 25. Total Heat of Super-
heated Steam. 26. Specific Volume of Dry Saturated
Steam, V or S. 27. Specific Density of Dry Saturated

Steam, — or 6. 28. Reversal of the Phenomena Just De-
scribed. 29. Generation of Steam in Real Steam Boiler.
30. Gauge Pressure.

CHAPTER IV

THE IDEAL STEAM ENGINE 43

31. The Engine. 32. Operation of the Engine. 33. Work
Done by the Engine. 34. Heat Quantities Involved. 35.
Efficiency. 36. Effect of Wet Steam. 37. Application to

vi CONTENTS

PAGE

a Real Engine. 38. Desirability of Other Cycles. 39. The
Complete Expansion Cycle. 40. The Incomplete Expan-
sion Cycle.

CHAPTER V

ENTROPY DIAGRAM 61

41. Definitions. 42. Temperature-Entropy Chart for
Steam. 43. Quality from TV-Chart. 44. Volume from T<£-
chart. 45. Heat from T<£-chart. 46. The Complete T0-
chart for Steam.

CHAPTER VI

TEMPERATURE ENTROPY DIAGRAMS OF STEAM CYCLES 72

47. Complete-expansion Cycle. 48. Area of Cycle Repre-
sentative of Work. 49. Modifications for Wet and Super-
heated Steam. 50. Incomplete Expansion Cycle. 51.
Effect of Temperature Range on Efficiency.

CHAPTER VII

THE REAL STEAM ENGINE » 77

52. Operations of Real Engine. 53. Losses in Real In-
stallations. 54. Clearance. 55. Cushion Steam and Cylinder
Feed. 56. Determination of Initial Condensation. 57.
Methods of Decreasing Cylinder Condensation. 58. Classi-
fication of Steam Engines. 59. Rotative Speeds and Piston
Speeds. 60. The Simple D-Slide Valve Engine. 61. Engine
Nomenclature. 62. Principal Parts of Engines.

CHAPTER VIII

THE INDICATOR DIAGRAM AND DERIVED VALUES 115

63. The Indicator. 64. Determination of I.h.p. 65.
Conventional Diagram and Card Factors. 66. Ratio of
Expansion. 67. Determination of Clearance Volume from
Diagram. 68. Diagram Water Rate. 69. TV-diagram for
a Real Engine. 70. Mechanical and Thermal Efficiencies.

CHAPTER IX

COMPOUNDING 141

71. Gain by Expansion. 72. Compounding. 73. The
Compound Engine. 74. Cylinder Ratios. 75. Indicator
Diagrams and Mean Pressures. 76. Combined Indicator
Diagrams.

CONTENTS vii

CHAPTER X

PAGE

THE D-SLIDE VALVE 159

77. Description and Method of Operation. 78. Steam Lap.
79. Lead. 80. Angle of Advance. 81. Exhaust Lap. 82.
The Bilgram Diagram. 83. Exhaust and Compression. 84.
Diagram for Both Cylinder Ends. 85. Piston Positions.
86. Indicator Diagram from Bilgram Diagram. 87. Limita-
tions of the D-slide Valve. 88. Reversing Engines. 89, Valve
Setting.

CHAPTER XI

CORLISS AND OTHER HIGH-EFFICIENCY ENGINES 196

90. The Trip-cut-off Corliss Engine. 91. Non-detaching
Corliss Gears. 92. Poppet Valves. . 93. The Una-flow En-
gine. 94. The Locomobile Type.

CHAPTER XII

REGULATION 213

95. Kinds of Regulation. 96. Governor Regulation. 97.
Methods of Varying Mean Effective Pressure. 98. Con-
stant Speed Governing. 99. Governors.

CHAPTER XIII

THE STEAM TURBINE 221

100. The Impulse Turbine. 101. Theoretical Cycle of
Steam Turbine. 102. Nozzle Design. 103. Action of Steam
on Impulse Blades. 104. De Laval Impulse Turbine. 105.
Gearing and Staging. 106. The Reaction Type. 107. Com-
bined Types. 108. Economy of Steam Turbines.

CHAPTER XIV

CONDENSERS AND RELATED APPARATUS 251

109. The Advantage of Condensers. 110. Measurement
of Vacuums. 111. Conversion of Readings from Inches of
Mercury to Pounds per Square Inch. 112. Principle of the
Condenser. 113. Types of Condensers. 114. The Jet Con-
denser. 115. Non-Contact Condensers. 116. Water Re-
quired by Contact Condensers. 117. Weight of Water
Required by Non-contact Condensers. 118. Relative Ad-
vantages of Contact and Surface Condensers. 119. Cool-
ing Towers.

viii CONTENTS

CHAPTER XV

PAGE

COMBUSTION 277

120. Definitions. 121. Combustion of Carbon. 122.
Combustion to CO. 123. Combustion to CO2. 124. Com-
bustion of CO to CO2. 125. Conditions Determining Forma-
tion of CO and CO2. 126. Flue Gases from. Combustion of
Carbon. 127. Combustion of Hydrogen. 128. Combustion
of Hydrocarbons. 129. Combustion of Sulphur. 130. Com-
bustion of Mixtures. 131. Temperature of Combustion.

CHAPTER XVI

FUELS 296

132. Commercial Fuels. 133. Coal. 134. Coal Analyses.
135. Calorific Value of Coals. 136. Purchase of Coal on
Analysis. 137. Petroleum.

CHAPTER XVII

STEAM BOILERS 305

138. Definitions and Classifications. 139. Functions of
Parts. 140. Furnaces and Combustion. 141. Hand Firing.
142. Mechanical Grates. 143. Smoke and its Prevention.
144. Mechanical Stokers. 145. Rate of Combustion. 146.
Strength and Safety of Boiler. 147. Circulation in Boilers.
148. Types of Boilers. 149. Boiler Rating. 150. Boiler Effi-
ciencies. 151. Effects of Soot and Scale. 152. Scale. 153.
Scale Prevention. 154. Superheaters. 155. Draft Apparatus.

CHAPTER XVIII

RECOVERY OF WASTE HEAT 375

156. Waste Heat in Steam Plant. 157. Utilization of
Exhaust for Heating Buildings. 158. Feed-water Heating.

CHAPTER XIX

BOILER-FEED PUMPS AND OTHER AUXILIARIES 382

159. Boiler-feed Pumps. 160. The Steam Injector. 161.
Separators. 162. Steam Traps. 163. Steam Piping.

STEAM POWER

CHAPTER I
PHYSICAL CONCEPTIONS AND UNITS

1. Matter. The universe is generally pictured as com-
posed of matter and energy. Matter is regarded as that
which is possessed of mass, or as that which is possessed of
inertia; i.e., which requires the action of force to put it in
motion, to bring it to rest or to change its velocity. These
definitions merely enumerate characteristics of matter; they
do not tell what it really is. In the present state of knowledge
it is, however, impossible to define matter in any other way.

No experiment has yet shown that matter can be created
or destroyed by man. It can be changed from one form to
another, it can be given certain physical and certain chem-
ical characteristics, more or less at will, but the actual
quantity of matter concerned is always the same after and
before such changes. It is customary to state this experi-
ence in the form of a law known as the Law of the Con-
servation of Matter, which states that the " total quantity
of matter in the universe is constant."

Matter is known to exist in several physical states or
conditions of aggregation. The three most familiar are (1)
solid, (2) liquid and (3) gaseous. In each of these states
matter is conceived as made up of minute particles called
molecules which in turn are apparently composed of still
smaller parts known as atoms. These atoms can also be
broken into parts, but for the purposes of this book it is not
necessary to consider such divisions.

2 STEAM POWER

Experiment and mathematical reasoning seem to indi-
cate that the molecules of all materials are in constant
motion and that there are neutralizing attractive and repul-
sive forces acting between them. In solids the molecules are
apparently bound together in such a way that, although they
are in constant motion, the external form or shape of the
body tends to remain constant; in fact it requires the
expenditure of force to cause a change of form. In liquids
the molecular attraction is so altered that practically all
rigidity disappears and the shape assumed by the liquid is
determined by that of- the surrounding surfaces, as, for
instance, the shape of the vessel containing the liquid. In
gases the molecules are still more free and actually tend to
move apart as far as possible, so that a gas will spread in
all directions until it fills any closed containing vessel.

2. Energy. Nearly everyone has a conception of what
is meant by the term energy, but no one yet knows what
energy really is. It is defined as the capacity for doing work,
or the ability to overcome resistance. A man is said to be very
energetic or to be possessed of a great deal of energy when
he has the ability to perform a great amount of work or
to overcome great resistances. Matter is said to be pos-
sessed of energy when it can perform work or overcome
resistance. Actually, matter is not known in any form in
which it is not possessed of energy.

There are many different forms of energy. A body in
motion can do work and is said to be possessed of mechani-
cal energy. A body which we recognize as hot can do work
at the expense of the heat associated with it and is said
to be possessed of heat energy. Light, sound and electricity
are all forms of energy.

Experiment and experience have never shown that energy
can be destroyed or created by man, but they have shown
that one form of energy can be converted into another form
under proper conditions. The first part of this experience
is stated as a law known as the Law of the Conservation of

PHYSICAL CONCEPTIONS AND UNITS

Energy. This law states that " the total quantity of energy
in the universe is constant."

3. Units of Matter and of Energy. When attempts are
made to measure the amount of anything, some unit of
measurement is adopted. Matter is measured in numerous
ways and many units are used. The common methods of
measuring matter are by volume and by weight. Engineers
in English-speaking countries use the cubic, yard, the cubic
foot or the cubic inch as units in measuring matter by vol-
ume and they use the pound, the ounce, the grain, etc. as
units in measuring matter by weight.

Energy is measured in many units and, in general, there
is a characteristic unit or set of units for each form in which
it occurs. Thus the foot-pound is very commonly used for
measuring mechanical energy; the British thermal unit for
measuring heat energy; and the joule for measuring electrical
energy. Some of these units will be defined and considered
in greater detail in subsequent paragraphs.

4. Work. Work is defined as the overcoming of
a resistance through a distance. Thus, work is done
when a weight is raised against the resistance
offered by gravity; work is done when a spring is
compressed against the resistance which the
metal offers to change of shape; work is done
when a body is moved over another against the
resistance offered by friction.

The unit of work is the quantity of work which
must be done in raising a weight of one pound
through a vertical distance of one foot. It is called
the foot-pound. Thus, one foot-pound of work
must be done in raising one pound one foot; two
foot-pounds of work must be done in raising two
pounds one foot or in raising one pound two feet.

If a weight of one pound were suspended from
a spring balance as shown in Fig. 1, the balance would in-
dicate a pull or force of one pound. No work would be

FIG. 1.

4 STEAM POWER

done by this force as long as the weight remained stationary,
because no resistance would be overcome through a distance.
If, however, the same weight were slowly or rapidly raised
a vertical distance of a foot, one foot-pound of work would
be done. A force or pull of one pound would then have
overcome a resistance of one pound through a distance of
one foot. In general:

Work in ft. -Ibs. = Resistance overcome in Ibs.X distance.
= Force in Ibs.X distance in ft.

so that if a force of 10 Ibs. pushes or pulls anything which
offers a resistance of 10 Ibs. while that something travels
a distance of, say, 5 ft., the work done will be given by the
expression,

Work = 10X5,

= 50ft.-lbs.

A body in falling a certain distance can do work equal to
its weight multiplied by the distance it falls because it could
theoretically raise an equal weight an equal distance against
the action of gravity, and the work done upon this second
body would be equal to its weight multiplied by the distance
through which it was raised.

It is very important to note that no work is done by a
force if there is no motion; resistance must be overcome
through a distance in order that work may be done. Thus,
a force of 1000 Ibs. might be required to hold something in
position, that is to balance a resistance, but no work wrould
be done if the body upon which the 1000-pound force acted
did not move. Again, a weight of 50 Ibs. held at a distance of
10 ft. above the surface of the earth would exert a downward
push or pull equal to 50 Ibs. on whatever held it in that
position; it would, however, do no work if held in that
position. If allowed to fall through the distance of 10 ft.
it could do 50X10 = 500 ft.-lbs. of work.

It is very convenient to represent graphically the action

PHYSICAL CONCEPTIONS AND UNITS 5

of forces overcoming resistances, that is, doing work. This
is done by plotting points showing the magnitude of the
force at the time that the body on which it is acting has
traveled different distances. Thus, suppose a constant
force of 10 Ibs. pushes a body a distance of 15 ft. against a
constant resistance of 10 Ibs. The force acting on the body
will have a value of 10 Ibs. just as the body starts to move,
a value of 10 Ibs. when the body has moved 1 ft., a value
of 10 Ibs. when the body has moved 2 ft., and so on. This
might be represented by points on squared paper as shown

Force in Pounds .

H-* tC Oi *" C< 05 -> ;x CO C

0 1 2 3 4 56 7 8 9 10 11 12 13 14 15

Distance traveled in Feet

FIG. 2.

in Fig. 2 or by a horizontal line joining those points as shown
in the same figure.

The work done by this force would be 10X15 = 150
ft.-lbs. according to our previous definition. But 10X15
is also the number of small squares under the line represent-
ing the action of this force in Fig. 2. The number of these
small squares then must be a measure of the work done,
but it is also a measure of the area under the line represent-
ing the action of the force, so that this area must be a measure
of the work done. Each small square represents 1 Ib. by
its vertical dimension and l.ft. by its horizontal dimension,

STEAM POWER

so that its area must represent 1 Ib.Xl ft. = 1 ft.-lb.
The total number of squares below the line equals 10 X
15 = 150, and since the area of each one represents 1 ft.-lb.
the total area under the line represents 150X1 = 150
ft.-lbs.

It is not always convenient to choose such simple scales
as those just used. Thus it might be more convenient
to plot the action of this force as is done in Fig. 3. Here
the height of a square represents 2 Ibs. and the width
represents 1 ft.; the area then represents 2X1 = 2 ft.-lbs.
There are 5X15 = 75 squares under the line and as each

.12

567 8 9 10 11
Distance traveled in Feet

FIG. 3.

12 13 14 15

represents 2 ft.-lbs. the total area under the line represents
2X75 = 150 ft.-lbs. as before.

This is a very useful property of these diagrams and the
area under the line representing the action of the force
always represents the work done, no matter what the shape
of that line.

Thus, assume a force which compresses a spring a distance
of 6 ins. Suppose that a force of 10 Ibs. is required to com-
press the spring 1 in., a force of 20 Ibs. to compress it 2 ins.,
and so on up to a force of 60 Ibs. to compress it 6 ins.
Starting with a force of zero, the force will have to gradually
increase as the spring is compressed, as shown by the line in
Fig. 4. The area of each of the small squares will represent

=— ft.-lbs. Under the line there is an area equal

1- ' -

PHYSICAL CONCEPTIONS AND UNITS

/> \/ £*

o - —=18 small squares, and the work done in compressing
A

the spring must then be 18X^ = 15 ft.-lbs.

5. Mechanical Energy. Any body which exists in such
a position or location that it could do work by dropping or
falling is said to be possessed of potential mechanical energy,
or of mechanical energy due to position. As long as it
remains in this position, it cannot do work at the expense
of this energy, but, if allowed to fall, it could do so. The

e in Pounds

g S

1" 2" 3" 4" 5" 6"
K2 Ft. % Ft. %z Ft.

FIG. 4. — Graph Showing Action of Spring.

work it could do would be equal to the product of its weight
by the distance it could fall and the potential energy it
possesses before starting to fall is measured by this work.
Thus, a body weighing 40 Ibs. located 10 ft. above the surface
of the earth could do 40X10 = 400 ft.-lbs. of work in falling,
and, therefore, it is said to be possessed of 400 ft.-lbs. of
potential energy before it starts to fall.

If in falling it raises a weight equal to its own (theo-
retically) through a distance equal to that through which
it falls (theoretically), it will have used up 400 ft.-lbs. of
energy in doing 400 ft.-lbs. of work upon the body raised

8 STEAM POWER

and will no longer be possessed of that amount of potential
energy. The body which has been raised will, however,
have an equal amount of energy stored in it and will in turn
be able to do 400 ft.-lbs. of work if allowed to fall a distance
of 10 ft.

If the body assumed above falls through a distance of
10 ft. without raising another body or doing an equivalent
amount of work in some other way, it acquires a high
velocity. When it arrives at the bottom of the fall of 10
ft., it certainly does not possess the 400 ft.-lbs. of potential
energy which it had before dropping nor has it done work
at the expense of that energy. Moreover, the energy could
not have been destroyed because it is indestructible. The
only conclusion is that it must still be possessed of this
energy in some way. At the end of the fall it has lost its
advantageous position, but it has acquired a high velocity,
and experience shows that if brought to rest it can do
work upon that which brings it to rest equal to what
it could have done in raising a weight as previously
described.

At the end of its fall and before being brought to rest,
the body is therefore said to be possessed of energy by virtue
of its velocity, and this form of energy is called kinetic
mechanical energy. The kinetic energy will be exactly equal
to the potential energy which disappeared as the body fell.

Any body which is moving is possessed of kinetic energy
because it can do work on anything which brings it to
rest. This energy is expressed by the equation,

1 W

Kinetic Energy in ft.-lbs.=-Xi^^X V2,

_. . >_ ._

in which

' W = the weight of the moving body in pounds.

F = the velocity in ft. per second, and
32.2 = a gravitational constant commonly represented by g

PHYSICAL CONCEPTIONS AND UNITS 9

6. Heat. One of the most familiar forms of energy is
heat, which manifests itself to man through the sense of
touch. In reality every body with which man is familiar
possesses an unknown amount of heat energy and it is
assumed that this heat energy is in some way associated
with the motions and relative positions of the molecules and
their constituents.

For this reason heat is often described as molecular
.activity and is regarded as energy stored up in a substance by
virtue of its molecular condition. Heat energy can be made
to perform work in ways which will be discussed later and
this is proof that it is a form of energy and not a material
substance, as was once supposed.

Heat is observed and recorded by its effects on matter,
producing changes in the dimensions or volumes of objects;
changes of internal stress; changes of state, as ice to water
and water to steam; changes of temperature; and electrical
and chemical effects.

Neglecting certain atomic phenomena not yet well under-
stood, the probable source of all heat energy appearing on
the earth is the sun. Heat, however, may be obtained
from mechanical and electrical energy; from chemical
changes; from changes of physical state; from the internal
heat of the earth.

7. Temperature. Man early realized that under certain
conditions bodies felt " hotter " than under other conditions
and gradually came to speak of the " degree of hotness " as
the temperature of the body. It was later realized that what
was really measured as the " hotness " or intensity of heat
or temperature of a body was the ability of that body to trans-
mit heat to others and that it had no connection with quantity
of heat.

Thus, if the temperature of two adjacent bodies happened
to be the same, one of them could not lose heat by trans-
mitting it to the other, but if the temperature of one
happened to be higher than that of another, the body at

10 STEAM POWER

higher temperature would always lose heat to the one at
lower temperature.

As a means of measuring temperature certain arbitrary
scales have been chosen. The centigrade scale of tempera-
ture, for instance, is based upon the temperatures of melting
ice and boiling water under atmospheric pressure. The tem-
perature difference between boiling .water at atmospheric
pressure and melting ice at atmospheric pressure is arbi-
trarily called one hundred degrees of temperature, and the
temperature of the melting ice is called zero, making that
of the boiling water 100 degrees.

Any body which has such a temperature that it will not
give heat to, or take heat from, melting ice is said to be at
a temperature of zero degrees centigrade, represented as 0° C.
Similarly, any body in such a condition that it will not give
heat to or take heat from water boiling under atmospheric
pressure is said to have a temperature of 100° C. A body
with a temperature exactly half way between these two
limits would then be said to have a temperature of 50° C.

8. Measurement of Temperature. The temperatures of
bodies could be determined by bringing them in contact with
such things as melting ice and boiling water and determining
whether or not a transfer of heat occurred, but this would
be a very cumbersome and unsatisfactory method. As a
consequence many other means have been devised for the
measurement of temperature.

One of the most common and convenient methods in-
volves the use of what are known as mercury thermometers.
These depend upon the fact that the expansion of mercury
with changing temperature is very uniform over a wide
temperature range. Thus, if mercury expands a certain
amount when its temperature is raised from that of melting
ice to that of boiling water, i.e., 100° C., it will expand just
half as much when its temperature is raised half as high,
and one-quarter as much when its temperature is raised one-
quarter of the range from 0° to 100° C.

PHYSICAL CONCEPTIONS AND UNITS

The thermometer is made by enclosing a small quan-
tity of mercury in a glass tube fitted with a bulb at
one end, as shown in Fig. 5. The lower end
of the thermometer is immersed in melting
ice and the point on the stem which is
reached by the top of the mercury column is
marked and labelled 0° C. The thermometer
is then immersed in the steam from water boil-
ing under atmospheric pressure and the point
reached by the top of the mercury column is
marked and labelled 100° C. The distance be-
tween the two marks is then divided into one
hundred parts and each represents the distance
which the end of the column of mercury will
move when its temperature changes one centi-
grade degree.

It is customary to extend this same scale
below 0° and above 100°, carrying it, on ex-
pensive thermometers, as far in each direction

, ! . , • • FIG. 5.— Mer-

as the approximation to a constant expansion cur Ther

on the part of the mercury and to constant m0meter.
properties of the glass justifies.

The temperature of a body can then
be found by placing the thermometer in
or in contact with that body and noting
the point reached by the end of the
mercury column. The division reached
gives the temperature directly.

The centigrade scale just described is
the one commonly used by scientists the
world over, but engineers in this country
more often use what is known as the
FIG ^-Comparison Fahrenheit scale. This is so chosen that
of Centigrade and ,, - ,,.

Fahrenheit Scales. the temperature of melting ice is called
32° F. and the temperature of water boil-
ing under atmospheric pressure is called 212° F. There are

c.'

17.8

STEAM POWER

thus 180° on this scale for the same temperature difference
as is represented by 100° on the centigrade scale. The
relation between the two scales is shown diagrammatically
in Fig. 6. It is apparent that the temperature of a body
at 0° C. will be 32° F. and that of a body at 0° F. will be
-17.8CC.

Since 100 centigrade degrees are equal to 180 Fahren-
heit degrees, it follows that

i o (^ AUV^ ^ o T7<

1 c Too 5 F'

and that

1 o -p 100 5 o n
1 F = l80 = 9 C'

(1)

(2)

Therefore, if tF and tc represent temperatures on the Fahren-
heit and centigrade scales respectively,

and

(3)
(4)

Cent .

Fahr.

100

-273

459.4

Ma-

is still another temperature scale of great impor-
tance. It is known as the absolute
scale and temperatures measured
on it are spoken of as absolute tem-
peratures. The zero on this scale
is located at -273° C. or 273 centi-
grade degrees below centigrade
zero, or, what is the same thing,
at -459.4° F., or 459.4 Fahrenheit
degrees below Fahrenheit zero. The
degrees used are either centigrade

FIG. 7. — Comparison of Ab- or Fahrenheit, as convenient, so
solute and Ordinary that there are absolute tempera-
tures expressed in centigrade de-

grees above absolute zero and there are absolute tempera-

459.4°

PHYSICAL CONCEPTIONS AND UNITS 13

tures expressed in Fahrenheit degrees above absolute zero.
The relations between the various scales are shown dia-
grammatically in Fig. 7.

It is apparent from this diagram that,

TV = ^+460 (approximately) . ...'.. (5)
and that

f ...... ". J (6)

if TF and Tc represent absolute temperatures and if the
number 459.4 is rounded out to 460, as is commonly
done.

9. The Unit of Heat Energy. The unit used in the
measurement of heat energy in the United States is the
British Thermal Unit (abbreviated B.t.u). It is defined as
the quantity of heat required to raise the temperature of one
pound of pure water one degree Fahrenheit. In order to
make the definition very exact it is necessary to state the
temperature of the water before the temperature rise occurs,
because it requires different amounts of heat to raise the
temperature of a pound of water one degree from differ-
ent initial temperatures. For ordinary engineering pur-
poses, however, such refinements generally may be omitted.

Many experimenters have shown that heat energy and
mechanical energy are mutually convertible, that is, the one
can be changed into the other. When such a change occurs
no energy can be lost since energy is indestructible, and it
follows that, if one form is changed into the other, there
must be just as much energy present after the change as
there was before.

As the units used in measuring the two forms of energy
are very different and as it is often necessary to express
quantities of energy taking part in such conversions, it is
desirable to determine the relations between these units.
This was first accurately done by Joule, who showed that one
British thermal unit of heat energy resulted from the con-

14 STEAM POWER

version of 772 ft.-lbs. of mechanical energy. Later experi-
menters have shown that the number 778 more nearly
expresses the truth than does the number 772 and the larger
value is now known as Joule's Equivalent.

Expressed mathematically, the relation between the units
is

lB.t.u. = 778 ft.-lbs ...... (7)

(8)

10. Specific Heat. The specific heat of a substance is
defined as that quantity of heat which is used up or recovered
when the temperature of one pound of the material in question
is raised or lowered one degree. Its numerical value depends
upon the specific heat of water since the quantity of heat is
measured in units dependent upon the amount required to
raise the temperature of water. The specific heat of water
is, however, very variable, as shown by the values given in
Table I., and it is therefore evident that exact numerical
values of specific heats can only be given when the definition
of the B.t.u. is exactly expressed.

The specific heats of all real substances vary with tem-
perature and the values commonly used are either rough
averages or are those determined by experiments at one
temperature. For most engineering purposes errors arising
from this source may, however, be neglected.

From the definition of specific heat it follows that :

in which

C = a mean or average specific heat over a range of tem-
perature from ti to <2, and

Q = the heat supplied to raise the temperature of W
pounds of material from t\ to £2.

PHYSICAL CONCEPTIONS AND UNITS

TABLE I
SPECIFIC HEATS OF WATER.*

(Value at 55° F. taken as unity)

Temp. F°.

Spec. Ht.

Temp. F°.

Spec. Ht.

1.0168

350

1.045

1.0098

400

.064

1.0045

450

.086

1.0012

500

.112

0.9990

510

.117

• 70

0.9977

520

.123

0.9970

530

.128

0.9967

540

.134

100

0.9967

550

.140

120

0.9974

560

.146

140

0.9986

570

.152

160

1.0002

580

.158

180

1.0019

590

.165

200

1.0039

600

.172

220

1.007

240

1.012

260

1.018

280

1.023

300

1.029

* Values taken from Marks and Davis, " Steam Tables and Diagrams," p. 68.

ILLUSTRATIVE PROBLEMS

1. Given: Sp. ht. of iron=0.113, of aluminum =0.211; Initial
temp. = 150° F. Temp, range (fe -ti) = 100° F.

If 1 Ib. of iron and 1 Ib. of aluminum are cooled through this
temperature range, how much more heat is lost in one case than
in the other?

Qlr

21.1 B.t.u.
1 X .113X 100 = 11.3 B.t.u.

Difference 9.8 B.t.u.

2. If the difference obtained in Prob. 1 were used to heat up
5 Ibs. of silver, with a specific heat equal to 0.057, what would be
the temperature range through which it would be raised?

Q =9.8 =5X0.057(«2 -fc) =0.285(fe-«i)

16 STEAM POWER

3. If the initial temperature of the silver in Prob. 2 were 150° Ft
what would be the final absolute temperature Fahr.?

t2 = ^+34.4° = 150+34.4 = 184° (approximately) .
r, =460+184 =644° F. Abs.

4. 100 Ibs, of water in a 20-lb. tank of iron, both at 60° F.,
are placed in salt brine at 0° F. The water becomes ice at 32° F.
and the temperature'of the ice is lowered to 26° F., the brine being
raised to 26° F. Sp. ht. water = 1.0; Sp. ht. ice =0.5; Sp. ht.
iron =0.113; Sp. ht. brine =0.8; and 143 B.t.u. per pound of water
must be removed to convert liquid water at 32° F. to ice at the
same temperature. What weight of brine is required?

100[l(60-32) + 143+.5(32-26)]+20X0.113(60-26)

= TFX0.8(26-0)
TF=840 Ibs. of brine.

11. Quantity of Heat. It is impossible to determine the
total quantity of heat in or " associated with " a substance,
because no means of removing and measuring all the heat
contained in any real material have ever been devised.
Since, however, the engineer is concerned with changes of
heat content rather than with the total amount of heat
contained, this fact causes him no difficulty.

For convenience in figuring changes of heat content,
it is customary to assume some arbitrary starting point or
datum and to call the heat in the material in question zero
at that point.

Thus, for example, if it were necessary to figure heat
changes experienced by a piece of iron weighing 5 Ibs. and
having a specific heat of 0.1138, and the temperature of this
iron never dropped below 40° F. under the conditions exist-
ing, this temperature might be taken as an arbitrary starting
point above which to figure heat contents. If the iron were
later found at a temperature of 75° F., •" the heat content
above 40° F." would be said to be

Q = CW (t2-ti) =0.1138X5(75-40) =2.27 B.t.u.

PHYSICAL CONCEPTIONS AND UNITS 17

This type of formula can only be used when the sub-
stance does not change its state between the limits of tem-
perature concerned. In the case of water which might
change to steam during such a rise of temperature, it might
be necessary to include other heat quantities in the cal-
culations, as shown in a later chapter.

12. Work and Power. Since steam engines are designed
for the purpose of converting the heat energy contained in
fuel into mechanical energy which may be used to perform
work, it will be necessary to consider the units used in
measuring work and power.

Work was defined in a previous paragraph as the over-
coming of a resistance through a distance, by the application
of a force; that is, a force expressed in pounds, multiplied by
the distance in feet through which the force acts, gives a product
expressed in foot-pounds.

The amount of work performed in a unit of time is termed
power, which may be denned as the rate of doing work.
Therefore,

Force X Distance

Power = =7- —^- -^. . . . (10)
Time (mm. or sec.)

The unit of power used by steam engineers is the horse-
power, which is equivalent to the performance of 33,000
ft.-lbs. of work per minute, or 550 ft.-lbs. of work per second,
or 1,980,000 ft.-lbs. per hour. Therefore, the horse-power
developed by any mechanism is

ft.-lbs. of work per min.

h'p-=

Since 33,000 ft.-lbs. of work can be accomplished only by
the expenditure of 33,000 ft.-lbs. of energy and since one
B.t.u. of energy is equal to 778 ft.-lbs., it follows that 33,000

oo
ft.-lbs. of work must be the equivalent of *Q =42.41 B.t.u.

4 4 O

It is customary to speak of power in terms of horse-

18 STEAM POWER

power-hours. One horse-power-hour means the doing of
work equivalent to one horse-power for the period of one
hour, or the doing of work at the rate of 33,000 ft.-lbs. per
minute for an hour. A horse-power-hour is therefore equiva-
lent to 33,000X60=1,980,000 ft.-lbs. As 33,000 ft.-lbs.
are equivalent to 42.41 B.t.u., it follows that 42.41X60 =
2544.6 or about 2545 B.t.u. are the equivalent of one horse-
power-hour.

The number 2545 should be memorized as it is very often
used in steam-power calculations. If an engine could deliver
one horse-power-hour for every 2545 B.t.u. it received, it
would be working without losses of any kind ; that is, all
the heat energy entering it would leave it in the form of
useful mechanical energy. It will be shown later that this
is impossible even in the most perfect or ideal engine.

REVIEW PROBLEMS

1. Express 32° F. in degrees centigrade.

2. Express 150° F. in degrees centigrade.

3. Express 250° C. in degrees Fahrenheit.

4. Express the results of problems 1, 2 and 3 in absolute values.

5. What is the heat equivalent of 233,400 ft.-lbs. of work?

6. Find the heat supplied 10 Ibs. of water when its temperature
is raised from 20° F. to 160° F., assuming the mean specific heat
over this range to be 0.997.

7. Find the temperature change of 2 Ibs. of lead (sp. ht. 0.0314)
when 20 B.t.u. are added.

8. How many B.t.u. must be abstracted to lower the tem-
perature of 15 Ibs. of water from 212° F. to32°F., assuming the
specific heat of water to be unity?

9. Find the weight of water which will have its temperature
tripled in value by the addition of 250 B.t.u., the final temperature
being 150° F. Assume specific heat unity.

10. The specific heat of a piece of wrought iron is 0.113 and of
a given weight of water is 1.015. 1 cu. ft. of water weighs approxi-
mately 62.5 Ibs. Find the increase in temperature of 4 cu. ft. of
water when a common temperature of 65° F. results from placing
in the water a piece of iron weighing 15 Ibs. at a temperature of
900° F.

PHYSICAL CONCEPTIONS AND UNITS 19

11. Find the final temperature of the mixture, when 100 Ibs.
of iron (sp. ht. =0.113), at a temperature of 1200° F. are immersed
in 300 Ibs. of water (sp. ht. 1.001) at a temperature of 50° F.

12. Five pounds of silver (sp. ht. = 0.057) at 800° F. are im-
mersed in water at 60° F., resulting in a final temperature of 85° F.
Assume Sp. ht. water = 1. What weight of water is necessary?

13. An engine is developing 10 horse-power. Express this in
ft .-Ibs. of work done per minute and find the amount of heat
energy equivalent to this quantity of mechanical energy.

14. A pump raises 1000 Ibs. of water 50 ft. every minute.
How much work is done? Find the equivalent horse-power.

15. An engine develops 1,980,000 ft.-lbs. of work at the fly-
wheel per minute.

(a) Find the horse-power developed.

(6) If this engine operated in this way for an hour, how many
horse-power hours would it make available?

CHAPTER II
THE HEAT-POWER PLANT

13. The Simple Steam-Power Plant. The various pieces
of apparatus necessary for the proper conversion of heat
energy into mechanical power constitute what may be
called a " Heat-Power Plant," just as the apparatus used
in obtaining mechanical energy from moving water is called
an hydraulic or water-power plant. Heat-power plants are
distinguished as "Steam-Power Plants"; "Gas-Power
Plants "; etc., according to the way in which the heat of
the fuel happens to be utilized.

The apparatus around which the plant as a whole centers,
that is, the apparatus in which heat energy is received and
from which mechanical energy is delivered, is termed the
engine or prime-mover. This heat engine may use steam
generated in boilers and may require certain apparatus, such
as condensers, pumps, etc., for proper operation; or it may
use gas, generated in gas-producers requiring coolers,
scrubbers, tar extractors and holders, depending upon the
class of fuel used and upon certain commercial considera-
tions. Again, the power-plant may simply contain an in-
ternal-combustion engine using natural gas, gasoline or oil,
a type of plant which is now very common.

But whatever type of plant is used, a general method of
operation is common to all. Heat energy in fuel is constantly
fed in at one end of the system and mechanical energy is
delivered at the other end. The steam-power plant will be
briefly described in the following paragraphs, showing the
cycle of events with the attendant losses through the
system.

THE HEAT-POWER PLANT

22 STEAM POWER

In Fig. 8 is shown a simple steam-power plant which con-
verts into mechanical energy part of the heat energy, origi-
nally stored in coal, by means of a prime-mover called a
steam-engine. The main pieces of apparatus used in this
type of plant are the steam-boiler; the steam-engine; the
condenser; the vacuum pump; and the feed-pump. The
energy stream shows the various losses occurring through-
out the plant. These losses cause the " delivered energy "
stream to be only a small fraction of the total heat sent into
the system.

14. Cycle of Events. 1. Fuel is charged on the grate
under the boiler, where it is burned with the liberation of a
large amount of energy. Air is drawn or forced through
the grates in proper proportions to support this combustion.
The hot gases resulting pass over the tubes, in a definite
path set by the baffle plates, so that the largest possible
amount of heating surface may be presented to the products
of combustion.

There are certain losses accompanying this operation,
such as radiation, loss of volatile fuel passing off unburned,
loss of fuel through the grate, and loss of heat through the
excess air which must always be supplied to insure com-
bustion.

2. That part of the heat in the gases which is not lost
by radiation from the boiler and in the hot gases flowing up
the stack passes through the heating surfaces of the boiler
to the water within. From 50 to 80 per cent of the
total heat energy in the fuel passes through the heating
surfaces and serves to raise the temperature of the water
to the boiling point at the pressure maintained, and to con-
vert this water into steam according to the requirements.

3. Having obtained steam within the boiler, it is led
through a system of pipes to a steam engine, where some of
the heat stored in the steam is converted into mechanical
energy by the action of that steam against a piston. The
steam is then discharged, or exhausted, from the engine

THE HEAT-POWER PLANT 23

at a much lower temperature and pressure than when it
entered.

From 5 to 22 per cent of the available heat in the
steam is converted into mechanical energy in the engine
cylinder, and because of frictional and other losses occurring
in the mechanism, only from 85 to 95 per cent of this
energy is turned into useful work at the fly-wheel.

4. In some plants, known as non-condensing plants,
the exhaust steam, which still contains the greater part of
all the heat received in the boiler, is discharged to the atmos-
phere and represents a complete loss. In others, known as
condensing plants, the exhaust steam is led to a condenser,
where it is condensed by cold water, which absorbs and
carries away the greater quantity of the heat not utilized
in the engine. The condensed steam or " condensate " is
then either discharged to the sewer or transferred by means
of a vacuum-pump to the hot-well, from which it is drawn
by means of the feed-water-pump, raised to the original
pressure of the steam, and returned to the boiler. Here
it is again turned into steam and the cycle of operations
outlined above is repeated. Naturally there is some loss
due to evaporation and leaks throughout the system, so
that " make-up " water must constantly be supplied.

The series of events just described constitutes a complete,
closed cycle of operations, wherein the water is heated,
vaporized, condensed and returned to the boiler, having
served only as a medium for the transfer of heat energy
from fuel to engine and the conversion of part of that
energy within the cylinder. The water in such a case is
known as the working substance.

It is often more convenient to discard the working sub-
stance after it leaves the cylinder, as suggested above in
the case of a non-condensing plant; or, as in the case of
a gas engine, where a new supply of working substance
must be supplied for each cycle, because the burned gases
of the previous cycle cannot be used again.

24 STEAM POWER

15. Action of Steam in the Cylinder. In order to pre-
pare for the more detailed discussion of the action of the
steam in the engine cylinder, to be taken up in a later
chapter, a brief outline of the events occurring within the
prime-mover will be considered at this point.

Steam enters the cylinder through some kind of an
admission valve, and acts upon the piston, just as the latter
has approximately reached one end of its stroke and is
ready to return. The heat-energy stored up in the steam
causes it to expand behind the piston, thereby driving the
latter out and performing work at the fly-wheel. At about
90 or 95 per cent of the stroke, the exhaust valve opens,
and the steam begins to exhaust, the pressure within the
cylinder dropping almost to atmospheric or to that main-
tained in the condenser by the time the piston has
reached the end of its stroke. On the next or return stroke
the remaining steam is forced out through the exhaust
port, until, at some point before the end of the piston
travel, the exhaust valve closes, and the low-pressure steam
trapped in the cylinder is compressed into the clearance
space so that its pressure rises. Admission then occurs, and
the cycle is repeated.

The diagram given in Fig. 9 illustrates the operation of

steam within the cylinder.
This diagram is plotted

Cut-off (Closing of Admission Valve)

between pressures 01 steam
within the cylinder as

Release (Opening of -• . •• j

Exhaust vaive) ordmates and correspond-
f\ ing piston positions as

abscissas,
piston positions The method of obtain-

Closing of Exhaust Valve

^ c, ,-, . ing such a diagram, known

FIG. 9. — Steam Engine Indicator '

Diagram, as an indicator-diagram,

will be fully described in a

later chapter. Since vertical ordinates represent pressure
in pounds per square inch, and horizontal abscissas renre-

THE HEAT-POWER PLANT

sent feet moved through by the piston, the product of
these two must be work. But the product of vertical by
horizontal distances must also give area. Therefore, by

Source of Water
at High Head

Energy Supplied

iver of Discharged
Water at Low Head

Head ft 2
above; datum

(3)

Source of Heat
at High Temp.

High Temp,
tj above datum

Energy Supplied

LowJTemp.
tg above datura

Useful Energy-
Made Available

Energy Discharged

eiver of Discharged
Heat at Low
Temperature

FIG. 10.— Hydraulic Analogy.

finding the area enclosed within the bounding lines of the
cycle and multiplying this by a proper factor, the foot-
pounds of work developed within the cylinder can be
determined.

16. Hydraulic Analogy. The operation of heat-engines
is analogous to that of water-wheels. A water-wheel de-

26 STEAM POWER

velops mechanical energy by receiving water under a high
head, absorbing some of its energy, and then rejecting the
fluid under a low head. Similarly, the heat-engine receives
heat energy at a high temperature (head), absorbs some of
it by conversion into mechanical energy, and then rejects
the rest at a low temperature (head).

The analogy can be carried still further. The water-
wheel cannot remove all the energy from the water, nor
can the heat-engine remove all the heat-energy from the
working substance. There is a certain loss in the material
discharged in both cases and this cannot be avoided.

This analogy is illustrated diagrammatically in Fig. 10
(a) and (6) in which the widths of the streams represent
quantity of energy.

CHAPTER III
STEAM

17. Vapors and Gases. When a solid is heated, under
the proper pressure conditions, it ultimately melts or fuses
and becomes a liquid. The temperature at which this
occurs depends upon the particular material in question
and upon the pressure under which it exists. Ice, which
is merely solid water, melts at 32° F. under atmospheric
pressure, while iron melts at about 2000° F. under atmos-
pheric pressure.

When a liquid is heated, it ultimately becomes a gas,
similar to the air and other familiar gases. If this gas is
heated to a very high temperature and if the pressure under
which it is held is not too great, it very nearly obeys certain
laws which are simple and which are called the laws of ideal
gases.

When the material is in a state between that of a liquid
and that in which it very nearly obeys the laws of ideal
gases, it is generally spoken of as a vapor. This term is
used in several different ways and with several different
modifying adjectives which will be explained in greater
detail in later sections.

18. Properties of Steam. Of the many vapors used by
the engineer, steam or water vapor is probably the most
important, because of the ease with which it can be formed
and also because of the tremendous field in which it can
be used. It is generated in a vessel known as a steam boiler,
which is constructed of metal in such a way that it can
contain water, and that heat energy, liberated from burning
fuel, can be passed into the water, converting part or all of
it into water vapor, that is, into steam.

STEAM POWEE

The properties of water vapor must be thoroughly under-
stood before the steam engine and steam boiler can be
studied profitably. Probably the easiest way of becoming
familiar with these properties is to study the use made of
heat in the generation of steam from cold water.

19. Generation of Steam or Water Vapor. For the pur-
poses of development, assume a vessel of cylindrical form,
fitted with a frictionless piston of known weight, as shown
in Fig. 11, (a) and (6), the whole apparatus being placed
under a bell-jar in which a perfect vacuum is maintained.

Illl

nil!,!

F "|

(b)

__:_/

iZE

c— c:

//////

//////,

/////

To Vacuum
Pump

FIG. 11. — Formation of Steam at Constant Pressure.

Assume one pound of water in the cylinder, with the piston
resting on the surface of the liquid. There will be some
definite pressure exerted by the piston upon the surface of
the liquid, and its value will be determined entirely by the
weight of the piston.

It is convenient in engineering practice to refer all
vaporization phenomena to some datum temperature, and
since the melting point of ice, 32° F., is a convenient refer-
ence point, it is used as a standard datum temperature, in
practically all steam-engineering work. Therefore, assuming
the water in the jar to be at 32° F., if heat is applied the
temperature of the liquid will rise approximately 1° F.

STEAM

for every B.t.u. of heat added, since the specific heat of
water is approximately unity.

Experiment shows that for each pressure under which the
water may exist some definite temperature will be attained at
which further rise of temperature will cease and the liquid will

480
440

•g 400
a

360

£240

i m

£ 1GO

120

0 40 80 120 ItiO 200 240 280 320 3UO 400 440 480 520
Temperature -F.°

FIG. 12.— Pressure-Temperature Relations, Saturated Water Vapor.

begin to change to a vapor, that is, to vaporize. The tem-
peratures at which vaporization occurs at different pressures
are called the temperatures of vaporization at those pressures.
The temperatures of vaporization of water are plotted
against pressure in Fig. 12. It should be noted that the
values of vaporization temperature increase very rapidly
for small pressure changes in the case of low pressures, but

30 STEAM POWER

that, for the higher pressures, the variation of temperature
is very small for enormous variations of pressure. This
fact is of great importance in steam engineering.

The temperatures of vaporization are tabulated with
other properties of water vapor in so-called steam tables and
are constantly referred to by engineers. An example of
such a table is given on pp. 392 to 399.

Returning now to the apparatus under discussion, as
heat is supplied, the temperature of the water will rise from
32° F. until it reaches the temperature of vaporization cor-
responding to the pressure exerted upon the water by the
piston. When this temperature is reached vaporization will
begin, and if sufficient heat is supplied, will continue without
change of temperature until the water is entirely converted
into vapor.

Up to the time at which vaporization starts the volume
of the water will change very little, so that the piston will
be raised only a negligibly small amount and practically no
work will be done upon it by the water. On the other hand,
when vaporization occurs the volume of the material will
change by a very large amount and the piston will be
driven out (raised) against the action of gravity. That is,
work will be done by the steam in driving the piston out
during the increase in volume which accompanies vaporiza-
tion.

It is found that a very great quantity of heat is used up
during the process of vaporization despite the fact that no
temperature change occurs. This is described by saying that
the heat which is supplied during this period becomes latent
in the steam formed, and the quantity of heat is therefore
spoken of as the latent heat of vaporization. It is assumed
to consist of two parts, that used for separating the liquid
molecules against their attractive forces and that used for
doing the work which is done upon the piston as it is
moved upward. The former is called the internal latent
heat because it is used for doing internal or intermolecular

STEAM 31

work; the latter is called external latent heat because it is
used for the doing of external work.

It is to be noted that the internal latent heat may be
assumed to be tied up in some way within the molecular
structure of the material and hence to be in the steam.
The external latent heat, on the other hand, is used up as
fast as supplied for the purpose of driving the piston out
against the action of gravity. When the piston has been
raised to any point, the energy used in raising it is not in
the steam, but is stored as potential energy in the piston.
To get it back the piston must be allowed to drop. The
term " external " is therefore well chosen; the external
latent heat is in no sense in the steam; it is stored in
external bodies or mechanism.

After the constant temperature vaporization is complete,
the further addition of heat will again cause a rise of tem-
perature and a gradual increase of volume. Such raising
of the temperature of steam already formed is called super-
heating and results in carrying the vapor nearer and nearer
to the condition in which it very nearly obeys the laws of
ideal gases. Since an increase of volume accompanies super-
heating, the molecules of the vapor must move farther and
farther apart as superheating progresses.

Vapor in the condition in which it is formed from the
liquid and which has the same temperature as the liquid
from which it was formed is called saturated vapor. This
term can be pictured as meaning that the maximum number
of molecules of vapor are packed into a given space; the
addition of heat to saturated vapor would cause superheating
and the separation of the molecules so that fewer could be
contained in a given space.

20. Heat of Liquid, q or h. Returning once more to
the start- of the process described in the preceding section,
heat was added to water initially at 32° F. until the tem-
perature of vaporization corresponding to the existing pres-
sure was attained. The heat added during this period is

32 STEAMj POWER

called the heat of the liquid, and is usually designated by the
letters q or h. If the mean specific heat of water at con-
stant pressure (Cp) for the temperature range under con-
sideration were constant, and equal to 1, then, since

in which tv is the temperature of vaporization, it would
follow that, for this pressure

q = tv-32 ........ (12)

Therefore, if water boils under a pressure of 50 Ibs. at
a temperature, read from the steam tables, of 281° F.; it
would follow that

But the steam tables (see p. 394) for this pressure (50
Ibs.) give 5 = 250.1 B.t.u., indicating, as was shown in
Chap. I., that the specific heat of water does not remain
constant, and for this case the mean value must have been
approximately 1.004 as indicated by the following calcula-
tion.

q = Cp (k-32) or
so that

^250.1
p 249

Hence it is always advisable to use the steam table
values of q, except for very approximate calculations.

21. Latent Heat of Vaporization, r or L. The heat
supplied during the period of vaporization has already been
referred to as the latent heat of vaporization and has been
divided into internal and external latent heats.

The internal latent heat is generally designated by p or
by 7 and the external latent heat by the group of letters
APu or by E. The group APu merely represents the prod-

STEAM 33

uct of pressure, P, by volume change during vaporization,
u, and by the fraction T-i¥ which is represented by A. The
product of the first two terms gives external work in foot-
pounds during vaporization, and dividing this by 778 (Joule's
Equivalent) converts it to heat units to correspond with the
other values. It should be noted that P in this expression
stands for pressure in pounds per square foot.

The total latent heat of vaporization is generally desig-
nated by r or by L, and it follows from what has preceded
that j

r = p+APu (13)

The value of r for atmospheric pressure, that is, for a
temperature of vaporization of 212° F., is very often used
in engineering and should be memorized. Its value is now
generally taken as 970.4 B.t.u., though recent work would
seem to indicate a value of about 972 as nearer the truth.

22. Total Heat of Dry Saturated Steam, X or H. The
total heat required to convert a pound of water at 32° F.
into a pound of saturated vapor at some temperature tv is
called the total heat of the steam or the heat above 32° and
is designated by X or by H. It is obviously the sum of the
quantities which have just been considered, so that

\ = q+r = q + p+APu. .... (14)

23. Total Heat of Wet Steam. In practical work the
engineer seldom deals with pure saturated steam, the satu-
rated vapor nearly always carrying in suspension more or
less liquid water at its own temperature. To distinguish
between saturated steam which carries liquid water and that
which does not, the former is called wet steam or wet
saturated steam, and the latter dry saturated steam.

The condition of dryness or wetness is described by what
is known as the quality of the steam. Dry saturated steam
is said to have a quality of 100 per cent while saturated
steam carrying 10 per cent by weight of liquid is said to

34 STEAM POWER

have a quality of 90 per cent. Quality expressed as a
decimal fraction is designated by the letter x, so that if x
is said to be equal to 0.8 in referring to a certain sample of
steam, it means that that steam sample consists of 80 per
cent by weight of saturated steam and 20 per cent liquid
at the same temperature.

Since the water in wet steam has the same temperature as
the steam, it contains all the heat of the liquid which it would
contain if it had been converted into steam, but it obviously
contains no latent heat of vaporization. It follows that the
total heat in a pound of wet steam (one pound of a mixture of
saturated steam and water) with quality equal to x is

Heat per pound = q -f- xr = q + xp + x A Pu . . (15)

The letter X should never be used in designating the total
heat per pound of wet steam, as it has been chosen as the
symbol of the total heat per pound of dry, saturated steam.

24. Heat of Superheat. When the temperature of
saturated steam is raised by the addition of more heat, that
is, when it is superheated, a very definite quantity of heat
is required. The quantity required per pound per degivcc
would, by definition, be the specific heat of the material in
question.

If the specific heat of superheated steam were reasonably
constant, the heat required to raise its temperature at
constant pressure from saturation temperature to some
higher value fe would be given by the expression

Heat required per pound = Cp(t2 — 10)

but superheated steam, as handled by the engineer, is
generally comparatively near the saturated condition, and
under these circumstances the values of the specific heat vary
rapidly with changes of pressure and temperature. The
extent of these variations is shown in Fig. 13. It will be
observed that for low pressures the specific heat is approxi-

STEAM

mately constant at a value below 0.5 for any given pressure,
but that for very high pressures it varies widely over a
comparatively small temperature range. Thus at 600 Ibs.
per square inch the specific heat changes from unity at
about 510° F. to 0.6 at about 550° F.

Practically, it is customary to use the type of equation
just given and to substitute a mean specific heat over the
required temperature range for the specific heat which can-
not be assumed constant without too great an error. The

0 100 200 300 400 500
Temperatures Deg-. Fahr.

FIG. 13. — Progressive Values of Specific Heat, Cp, Water Vapor.

equation for heat required to raise the temperature from
tv to t2 is then

Heat of superheat per pound =

(16)

in which Cpm stands for the mean specific heat at constant
pressure over the temperature range from tv to t%.

Values of mean specific heats of superheated steam are
given in Fig. 14, the values indicated by the curves giving
the mean specific heat between saturation temperatures and
Various higher temperatures at different pressures.

STEAM POWER

25. Total Heat of Superheated Steam. The total heat
required to convert one pound of water at 32° F. into
superheated steam at a temperature of fe° F. under constant
pressure conditions is obviously

Total heat per pound = q+r+Cpm(t2 -t9). . (17)

100 150 200 250

Temperatures above Saturation0!?.

300

FIG. 14. — Variation of Mean Specific Heat, Water Vapor.

and representing the degrees of superheat (fe — tv) by D, as
is customary, this becomes

Total heat per pound = q+r+CpmD. . . (18)

26. Specific Volume of Dry Saturated Steam, V or S. The

volume occupied by one pound of a substance is spoken of as
the specific volume of that material. In the case of dry
saturated steam there are as many specific volumes as there
are pressures under which the steam can exist. These
values are generally tabulated in steam tables and are
represented by the letter V or the letter S.

STEAM

The values of the specific volumes of steam at different
pressures are given in Fig. 15. It is important to note the
very gradual change of specific volume at high pressures
and the very rapid change and enormous increase at low
pressures. These facts have considerable influence on steam
engineering practice.

0 24 0 8 10 12 14 16 18 20 22
Specific Volume of Dry Saturated Steaua (Cubic Feet)

FIG. 15. — Pressure- Volume Relations, Saturated Water Vapor.

A curve giving properties of saturated steam is called a
saturation curve, so that this name may be, and often is,
applied to the curve given in Fig. 15.

The volume occupied at any pressure by half a pound
of dry saturated steam will obviously be half that occupied
by one pound of such material at the same pressure, and

38 STEAM POWER

the same statement can be made for any other fraction of
a pound. It follows that if the small volume occupied by
liquid water in wet steam be neglected, the volume occupied
by one pound of steam (mixture) of 50 per cent quality
can be assumed equal to half that occupied by an equal
weight of dry saturated steam at the same pressure. A
similar statement could of course be made for any other
quality and a corresponding fraction.

Hence if one pound of " wet steam " at a given pressure
is found to have such a volume that it would be indicated by
point b in Fig. 16, the quality of this material must be given

by the expression x = -— if the volume occupied by the

liquid water in the mixture be
neglected.

27. Specific Density of Dry

Saturated Steam, - or 5. The

weight per cubic foot of saturated
steam is spoken of as its specific
density .: The specific density
is obviously the reciprocal of

the' specific volume and is there-
FIG. 16. — Determining Quality

from Volume. fore —

28. Reversal of the Phenomena Just Described. If any

process which has resulted in the absorption of a quantity
of heat by a substance be carried through in the reverse
direction, the same amount of heat will again be given up.
It follows that a pound of dry saturated steam will give up
the total latent heat of vaporization when condensed to
liquid at the same temperature, and that the resultant pound
of hot water will give up the total heat of the liquid if cooled
to 32° F.

29. Generation of Steam in Real Steam Boiler. The
steam boiler is equivalent to a vessel partly filled with water

STEAM

and fitted with means for supplying heat to the water and
for carrying off the vapor formed. This is shown diagram-
matically in Fig. 17. At first glance this would not seem
to be at all similar to the cylinder and piston already con-
sidered, but it really is the exact equivalent so far as the
generation of steam is concerned. The flow of steam out
of the steam-pipe is restricted to the extent necessary to
maintain a high and constant pressure within the boiler, and,
when in regular operation, steam is formed within the

o r - :q a

1"
. • "•„ " " v> " •• «•' ' o .. ' • , -

r* " " "'/*

r- _-_-_— _^-^_-

l)'w/i
///'•v

jMiiiJi£^'

t=r-^— — — _ _— -l-pr-pr}

FIG. 17. — Formation of Steam in a Steam Boiler.

boiler under this pressure just as fast as necessary to replace
that flowing out.

By picturing the steam as flowing out in layers or lamina
these lamina can be imagined as taking the place of the
piston in the apparatus of Fig. 11, and each pound of steam
formed will then push a piston before it exactly as was
assumed in the previous discussion.

30. Gauge Pressure. The steam pressure in a boiler is
commonly determined by means of an instrument called a
pressure gauge. These instruments are almost always con-
structed about as shown in Fig. 18 (a) and (6) . The Bourdon
spring is a tube of elliptical section bent approximately into

STEAM POWER

the arc of a circle. One end of this tube is connected directly
to the pressure connection of the gauge and the other end
is closed and connected to a toothed sector as shown.

When the pressure inside a tube of this character is
increased, the tube has a tendency to unroll or straighten
out, and in so doing it moves the toothed sector in such a
way as to rotate the pointer or gauge hand and make its end
move over the scale in the direction of increasing pressure.
With diminishing pressure the tube again rolls up and
rotates the hand in the opposite direction.

(a) (6)

FIG. 18. — Bourdon Pressure Gauge.

Instruments of this kind are so made and adjusted that
the hand points to zero when the gauge is left open to
the atmosphere. Under such conditions the pressure inside
the tube is equal to that of the atmosphere and is not zero.
The gauge therefore only indicates pressures above atmos-
pheric on its scale, and the total pressure inside the boiler
is really that shown by the gauge plus that of the atmos-
phere.

Pressures as indicated by the gauge are called gauge
pressures. Pressures obtained by adding the pressure of

STEAM 41

the atmosphere to the reading of the gauge are known as
absolute pressures. Then

Absolute Pressure = Gauge Pressure-}- Atmospheric Pressure
and

Gauge Pressure = Absolute Pressure — Atmospheric Pressure.

In accurate work the existing atmospheric pressure
should be determined by means of the barometer, but for
ordinary, approximate calculations and for cases in which
no barometric data are available, it is customary to assume
the pressure of the atmosphere to be equal to 14.7 Ibs. per
square inch. This is very nearly true, on the average, at
sea level, but is generally far from true at higher elevations.

PROBLEMS

1. Determine by means of the steam tables the temperatures,
total heats, heats of liquid, internal and external latent heats, and
the specific volumes of 1 Ib. of dry, saturated steam under the fol-
lowing absolute pressures (Ibs. per sq. in.): 15, 50, 95, 180 and 400.

2. Determine the heats of the liquid, latent heats of vapor-
ization and total heats for 2 Ibs. of dry saturated steam at the
following temperatures in °F.: 101.83, 212 and 327.8.

3. Determine the volumes occupied by 2 Ibs. of dry saturated
steam under the conditions of problem 2.

4. Determine the heats of the liquid, latent heats of vapor-
ization and total heats for 1 Ib. of saturated steam with a quality
of 90% at the following absolute pressures: 25, 50, 75, 125.

5. Determine the total heat above 32° F. in 12 Ibs. of saturated
steam with quality of 97% at a pressure of 125 Ibs. per square inch
absolute.

6. AVhat space will be filled by 20 Ibs. of dry saturated steam
at a pressure of 150 Ibs. per square inch absolute?

7. What space will be filled by 20 Ibs. of saturated steam at a
pressure of 150 Ibs. per square inch absolute and with a quality
of 95% if the volume occupied by the water present be neglected?

8. How many pounds of dry saturated steam at a pressure of
75 Ibs. per square inch absolute will be required to fill a space of
10 cu. ft.?

42 STEAM POWER

X 9. How many pounds of saturated steam with quality 96%
and at a pressure of 1 10 Ibs. per square inch absolute will be required
to fill a space of 8 cu. ft.?

10. How much external work, measured in B.t.u., is done when
1 Ib. of water at the temperature of 212° F. is converted into dry
saturated vapor at the same temperature?

X 11. How much external work, measured in foot-pounds, is
done when 2 Ibs. of water at a temperature of 212° F. are converted
into 90% quality steam at the same temperature?

12. How much heat is required for doing internal work during
the vaporization of 1 Ib. of water under such conditions that the
total latent heat of vaporization is 852.7 B.t.u. and the external
latent heat is 83.3 B.t.u.?

13. What is the quality of steam containing 1000 B.t.u. above
32° F. per pound when under a pressure of 150 Ibs. per square
inch absolute?

)( 14. Heat is added to 1 Ib. of mixed steam and water while the
pressure is maintained constant at 100 Ibs. per square inch absolute.
The percentage of steam in the mixture is increased thereby from
50% to 95%.

(a) How much heat was added?

(b) How much internal latent heat was added?

15. How much heat is required to completely vaporize 1000
Ibs. of water at a temperature of 92° F. when pumped into a boiler
in which steam is generated at a pressure of 150 Ibs. per square
inch gauge? Note that heat above 32° F. in 92° F. water is given
as q in steam tables for a temperature of 92° F. //#j~- ^ ' /J^S

16. Find the amount of heat necessary to produce in a boiler
200 Ibs. of steam having a quality of 97% at a pressure of 100 Ibs.
gauge when the feed water has a temperature of 205° F.

17. What volume would be occupied by the material leaving
the boiler in problem 16, neglecting volume occupied by water?

CHAPTER IV
THE IDEAL STEAM ENGINE

31. The Engine. If the cylinder and piston assumed in
the discussion* of the last chapter be imagined as turned
into a horizontal position and fitted with a frame, piston

Fly wheel

FIG. 19. — Sirr.ple Steam Engine.

rod, crosshead, connecting rod, crank shaft and flywheel
as in Fig. 19, a device results which might be used as a
steam engine for the production of power. By adding heat
to, and taking heat from, the water and steam in the cylin-
der in the proper way and at the proper time, the water
and steam, or working substance, can be made to do work
upon the piston. The piston can transmit this work
through the mechanism to the rim of the flywheel, and it
can be taken from the rim by a belt connected to a pulley
on a machine which is to be driven.

STEAM POWER

To make the analysis easier, a simplified type of engine
will be assumed. It is shown in Fig. 20 and consists of the
same cylinder, piston and piston rod as just described.
A wire is fastened to the end of the piston rod and run back
over a pulley in such a way that a weight fastened to the
free end of the wire will be raised if the piston moves out.
The weight is made up of two parts, one large and one small.
When both are on the wire the pull which they exert causes
the piston to exert a high pressure upon whatever is con-

Cy'linder!

Volume

FIG. 20. — Simplified Steam Engine.

tained in the cylinder. When only the small weight hangs
on the wire, the piston exerts a much lower pressure upon
the material in the cylinder.

Imagine that the piston and the walls of the cylinder
are made of some ideal material which will not receive or
conduct heat. Imagine also that the cylinder is fitted
with a permanent head which is a perfect conductor of
heat. These conditions are of course ideal but are assumed
for the sake of simplicity.

Assume further that, when one pound of water is con-

THE IDEAL STEAM ENGINE 45

tained in the cylinder and the piston is driven into the
cylinder by the two weights until the space between the
piston and the cylinder head is just large enough to
contain the pound of water, the piston exerts a high
pressure equal to PI pounds per square foot against the
water. The volume of this water and the pressure upon
it can be represented by the point a of the PV diagram,
Fig. 20.

32. Operation of the Engine. With conditions as de-
scribed in the preceding paragraphs, imagine a flame or
other source of heat at high temperature to be brought into
contact with the conducting cylinder head and to pass heat
into the cylinder, raise the temperature of the water within
to the temperature of vaporization and ultimately vaporize
it. As the water vaporizes it will push the piston out of the
cylinder just as described in the last chapter and a hori-
zontal line such as ab in Fig. 20 will represent the increase
of volume (vaporization) at constant pressure. The point
b may be assumed to represent the volume of one pound
of dry saturated vapor at a pressure PI. Obviously the
steam, as it is formed, does work in driving out the piston
against the resistance offered by the weights which must be
raised.

If a stop is provided which will prevent the movement
of the piston beyond the position corresponding to the
point 6, it will be possible to remove the larger weight when
that point is reached and the high pressure steam will hold
the piston and rod hard against the stop. If now some
cooling medium is applied, such as a large piece of ice held
against the conducting head of the cylinder or water running
over that head, heat will be abstracted and a partial con-
densation of the steam within the cylinder will occur.
As condensation progresses the pressure will drop because
there will be less and less steam, by weight, in a given
volume. Such a process, would be indicated by the line
be which represents a drop of pressure, while the volume

46 STEAM POWER

contained within the cylinder walls between head and
piston remains constant.

When some point c is reached, the steam pressure will
have been reduced to a value equal to that exerted by the
small weight, and the piston will be driven in toward the
cylinder head while the heat absorbing medium continues
to remove heat from the steam and to cause further conden-
sation. The combination of piston motion and heat ab-
sorption will be so regulated that the pressure remains con-
stant at P2 during this process, because the weight will
move the piston inward just as fast as necessary to main-
tain a constant pressure. If sufficient heat is absorbed,
the pound of material within the cylinder will ultimately
all be condensed or liquefied and will just fill the volume Vd.

The heat absorbing body may now be removed and an
infinitesimal motion of the piston toward the head would
serve to raise the pressure on the liquid water from P% to
PI so that the volume Vd may be taken equal to the volume
Va and the line da may be assumed to be vertical. It
would then represent' an increase of pressure at constant
volume. This might be caused by hanging a weight of the
larger size on the wire when condition d was reached.

Having brought the material, or working substance,
back to the conditions originally shown at a, the high
temperature source of heat can again be brought in contact
with the end of the cylinder and the entire cycle carried
through once more. There is obviously no reason why it
could not be repeated as often as desired.

33. Work Done by the Engine. If the device just
described is to serve as a steam engine, it must actually
make mechanical energy available, that is, it must convert
into mechanical form, some of the heat energy supplied it.
It is now necessary to see whether it does so.

Water vaporizing and increasing in volume as from
Va to F& was shown in the last chapter to do work upon the
piston confining it. Work has been shown to be equal to

THE IDEAL STEAM ENGINE 47

(total force X total distance) and in this case if L repre-
sents the distance" in feet traveled by the piston, the work
done by the steam upon the piston while the latter moves
from a' to b' must be

Work done on piston = total force X distance ,

= PI X area of piston XL . . . ft.-lbs.

But the product of area of piston in square feet by
distance traveled in feet is equal to the piston displace-
ment or volume swept through by the piston, that is
(Vb— Va) cubic feet. Therefore

Work done on piston = Pi(Vb- Va) ft.-lbs. . (19a)

Pl(Vb-Va)

778

B.t.u. . (196)

The first form of this expression Pi(F&— Va) is very
obviously represented by the area under the line ab in
Fig. 20 and this area therefore represents the work done by
the steam upon the piston during the change of volume at
constant pressure represented by that line. While the
steam is supplying this amount of energy to the piston or
doing this amount of work upon the piston, the latter does
an equivalent amount of work upon the weights if friction-
less mechanism be assumed. In such a case the total
weight hung on the wire multiplied by the distance raised
would therefore give the same result in foot-pounds as
that just obtained.

It should be noted that Eq. (19) is merely an expression
of the external work done during vaporization, that is,
.•in expression of the amount of heat which is used for the
doing of external work. It is the exact equivalent of the
external latent heat previously discussed. In fact,' the
group of symbols APu is really a condensation of Eq. (19)

formed by putting A for — — and u for (F&— Fa).

7/8

48 STEAM POWER

The line cd also represents a change of volume at con-
stant pressure and the same type of formula as applied
to ab will express the work done during this process. In
this case, however, the piston is being pushed into the
cylinder by the small weight against the pressure of the
steam, and energy is being supplied to push the piston in.
This energy is equal to the weight of the small weight
(pounds) multiplied by the distance it falls (feet). The
piston is therefore doing work upon the steam, and the
amount is

Work done on steam = P2(Fc—Fd) ft, -Ibs. . . (20)
-^=^>B.t.u. . . (21,

The first form of expression also represents the area under
the line cd and this area therefore represents the work
done by the piston upon the steam mixture in the cylinder
during the process represented by cd.

No work can be done by steam on piston or by pistor
on steam during the processes represented by be or da
because both the weights and the piston are stationan*
during these changes and it has already been shown that
work involves motion.

The total work done upon the piston by the steam
is therefore represented by the area abef and this amount
of energy is used in raising the two weights through a
vertical distance equal to the piston travel. Some of
this energy, or its equivalent, will have to be returned an
instant later, however, in order that the piston may do the
work shown by the area cdfe upon the steam. It is returned
by the small weight dropping through a distance equal to
the travel of the piston. The net mechanical energy
made available by carrying through the series of processes
is therefore represented by the area (abef) — (cdfe) = (abed}
or the area enclosed by the four lines representing the

THE IDEAL STEAM ENGINE 49

pressure and volume changes experienced by the working
substance during one cycle of events. It is equal to the
work done in raising the larger weight a vertical distance
equal to the travel of the piston.

This net energy made available is obviously

Energy made available = PI (Vb- Va) -P2(Ve- Va}

= (Pi - P2)(Vb- Va) ft.-lbs. (22)

(Pl-P2}(Vb-Va)

778

B.t.u. (23)

Since this amount of energy is made available while one
cycle of events is being carried out and since the cycle
can be repeated time after time if sufficient heating and
cooling mediums are available, any quantity of mechanical
energy can be produced from heat energy by repeating the
cycle a sufficient number of times. This would correspond
to picking up a number of the larger weights which were
slid on to the wire at the lower elevation and slid off at
the higher.

This repetition of cycles would correspond, in a real
engine, to running at such a speed that the required number
of cycles would be produced in a given time to make avail-
able the amount of mechanical energy required.

Or, the power made available per cycle could be increased.
This is easily seen by an inspection of Eq. (22). Increas-
ing the value of either of the right-hand terms will obviously
increase the amount of energy made available. The
value of (Pi — Pz) can be increased by raising the initial
pressure PI or by lowering the final pressure PI. The value
of (F&— Fa) may be increased by using more than one pound
of material, thus increasing both the volume Vb of the satu-
rated steam formed and increasing the volume Va of the
liquid water, but getting a greater numerical value for
(Vb— Va). This would correspond in a real case to using
a larger cylinder and therefore a larger engine.

50 STEAM POWER

ILLUSTRATIVE PROBLEM

An engine of the type described is to work with a maximum
pressure of 100 Ibs. per square inch absolute and a minimum
pressure of 15 Ibs. per square inch absolute. The cylinder is to
be of such size that 1 Ib. of water is used and the steam is to be
dry and saturated at the point b of the cycle.

Find: (a) the amount of mechanical energy made available
per cycle; (b) the amount of energy made available per minute
if 150 cycles are produced per minute; and (c) the horse power of
the engine.

It will first be necessary to find the piston displacement required
and the space necessary between piston and cylinder head to
accommodate the pound of water in liquid form. The steam tables
give the volume of one pound of dry saturated steam at 100 Ibs.
per square inch as 4.429 cu.ft. and the volume of one pound of
water may be taken as 0.017 cu.ft. The values of the various
volumes and pressures will therefore be

Va = Fd =0.017 cu.ft.;

Vb = Vc =4.429 cu.ft.;

pa =pb = 100 X 144 = 14,400 Ibs. per sq.ft.;

Pc=Pd = 15X144 =2160 Ibs. per sq.ft.

(a) Using Eq. (22) the amount of mechanical energy made
available per cycle will be

(P, -TV) (Vb - Va) = (14,400 -2160) (4.429 -0.017)
= 12,240X4.412;
= 54,002.88 Tt.lbs.

(6) If 150 cycles are produced per minute, the total amount
of mechanical energy made available per minute must be

150X54,002.88-8,100,300 ft.-lbs.
(c) The horse power must then be

8,100,300
h*-- -33^00-

34. Heat Quantities Involved. It is a very simple
matter to determine the quantity of heat which must be
supplied to produce the process abt and the quantities of

THE IDEAL STEAM ENGINE 51

heat which must be removed to produce the processes
be and cd. This can be done by making use of the known
properties of water and steam as given in the steam
tables.

The water at d must be at the temperature of vaporiza-
tion corresponding to pressure PI since it has just been
formed by condensation from steam under that pressure.
It therefore contains the heat of the liquid corresponding
to that pressure. If it is to be vaporized at pressure PI, it
must first be raised to the higher temperature corresponding
to that pressure. The amount of heat required to do this
will obviously be the difference between the heat of the
liquid at the temperature corresponding to PI and the heat
of the liquid at the temperature corresponding to P^
These can be found in the steam tables.

The latent heat of vaporization at PI must then be added
to cause the increase of volume shown by ab. This can
also be found in the steam tables for any given case.

The quantity of heat which must be removed to produce
the processes represented by be and cd can be found sim-
ilarly from steam table values, although the exact method
of procedure is not quite as obvious as in the preceding
cases.

Assuming that it is possible to find the heat supplied,
Qi, and the heat removed, $2, it is obvious that the energy
made available in mechanical form, per cycle, must be
equal to (Qi — Qz) B.t.u., since this is the amount of heat
energy which has disappeared and since it cannot have
been destroyed. This may be put in the form of an equa-
tion, thus

Energy made available = Qi — Q.2. . . (24)

If the proper substitutions are made in this formula and it
is then simplified, it becomes

Energy made available =;(APu)Pi—xc(APu)Ps B.t.u., (25)

52 STEAM POWER

in which

p^ihe external latent heat at pressure PI;
p2 = the: external latent heat at pressure P%, and
xc = quality at point c, which can be found from
the ratio of dc to dc'\

Numerical substitution in this equation for any given
case will show that it gives exactly the same values as would
be obtained by the use of Eq. (23).

It is to be noted particularly that the energy made
available is actually less than the external latent heat at
the higher pressure, while the heat supplied must be equal
to the total latent heat plus some of the heat of the liquid.
An inspection of the steam tables will show that the exter-
nal latent heat for ordinary steam pressures forms a very
small fraction of even the total latent heat, and therefore
the mechanical energy made available for a given expendi-
ture of heat energy is very small in the case under dis-
cussion.

35. Efficiency. The term efficiency is used in engineer-
ing as a measure of the return obtained for a given expendi-
ture. It may be defined in any one of the following ways:

Useful result
Efficiency =

Expenditure made to obtain that result

Result
Effort

' ....... (26)

In the case of a heat engine, the useful result is the
mechanical energy obtained by the operation of the engine,
while the expenditure made is the heat which is supplied.
For this case efficiency may therefore be defined by the
expression

THE IDEAL STEAM ENGINE 53

Mechanical energy obtained per cycle
Engine efficiency = — ,.

Heat supplied per cycle

= .......... «>

in which

E stands for mechanical energy obtained,
Qi stands for heat supplied, and
Qz stands for heat rejected.

In the case of the type of steam engine just considered,
this efficiency would have a value between 6 and 8 per
cent for ordinary pressures. That is, the engine would
produce in mechanical form only 6 to 8 per cent of the
energy supplied it in the form of high temperature heat.
Moreover, these figures would hold only for a theoretically
perfect engine; a real engine built to operate upon this
cycle would probably give efficiencies of the order of 2 to
3 per cent. The reasons for this great discrepancy will be
discussed in a later chapter.

36. Effect of Wet Steam. In what has preceded, it
was assumed that the pound of steam was completely vapor-
ized along the line ab so that dry, saturated steam existed
in the cylinder at b. It might, however, be assumed that
vaporization was incomplete at the upper right-hand
corner of the cycle, so that this point occurred at a point
to the left of 6 and with a quality x at that point less than
unity.

Under such conditions, the cylinder would not have to
be so big, since the maximum volume attained by the steam
would be smaller than in the preceding case. The work
done per cycle would obviously be smaller in quantity,
because the area enclosed within the lines of the cycle would
be smaller. It can also be shown that the efficiency would
also be lowered by lowering the quality at b.

54 STEAM POWER

37. Application to a Real Engine. The engine which
has been described in the preceding paragraphs could easily
be converted into the counterpart of a real engine by sub-
stituting connecting rod, crank shaft and flywheel for wire,
pulley and weights as described in the first paragraph of
this chapter. It could then be made to do work in just
the same way as has been described; some of the energy
made available during the outstroke would be used for
overcoming resistance at the shaft, that is, doing useful work,
and some of it would be stored in the flywheel which would
speed up slightly. The energy which must be expended on
the steam during the return stroke would be obtained by
allowing the flywheel to slow down and thus deliver suf-
ficient kinetic energy to drive the piston back against the
low-pressure steam. The cycle and the efficiency would
thus, theoretically, be exactly the same as those just in-
vestigated .

Great difficulty would, however, be met in a real engine
if the steam had to be formed and condensed within the
cylinder, and another method which gives the same results
is therefore used. Steam is generated in a boiler and
allowed to flow into the cylinder and push out the piston
just as though it were actually being formed in the cylin-
der as previously described. When the piston reaches the
end of its outstroke the inlet valve is closed and the exhaust
valve is opened, allowing some of the steam to blow out
into a space in which a lower pressure exists. As the
piston stands still at the end of its stroke while the pres-
sure drops, the line be is produced as in the previous descrip-
tion, but by a different method. The piston then returns
and drives the remaining steam out of the cylinder at a
constant pressure theoretically equal to that of the space
into which the steam is being forced or exhausted. The
line cd is thus produced and the closure of the exhaust
valve and opening of the admission valve when d is reached
will start the cycle over again.

THE IDEAL STEAM ENGINE 55

In order to get more work out of a given size of cylinder
and to obviate the necessity of giving back energy which
has already been given out, engines are generally made to
take steam on both sides of the piston. They are then
known as double acting engines. In this case the steam
admitted on one side of the piston would supply the energy
necessary both for overcoming the resistance due to the
load and for driving out the low-pressure steam on the
other side of the piston. On the return stroke conditions
would be just reversed.

38. Desirability of Other Cycles. The cycle of opera-
tions described in preceding paragraphs is the most inef-
ficient of all those actually used, that is, it gives the small-
est return for a given amount of heat supplied. This
is because only the external latent heat supplied is con-
verted into mechanical energy and part of that energy
must be returned to complete the cycle. All of the internal
latent heat and all of the heat of the liquid supplied along
ab pass through the engine without conversion and are
exhausted.

Therefore, cycles which differ from that described
in such a way as to make it possible to convert into mechani-
cal energy some of the internal latent heat and possibly
some of the heat of the liquid should be highly desirable
as they ought to yield a larger return of mechanical energy
for the same total amount of heat supplied. Two such
cycles are commonly used; they may be described as the
Complete-expansion cycle and the Incomplete-expansion cycle.
The former is used in steam turbines, the latter in most
reciprocating steam engines. The rectangular cycle which
has just been described is used in duplex pumps and similar
apparatus.

39. The Complete-expansion Cycle. This cycle, which
is also known as the Clausius and as the Rankine cycle,
starts just the same as that already described. This is
shown in Fig. 21. The pressure on, say, a pound of water

STEAM POWER

is raised from P<z to PI and its temperature is raised from
that of vaporization at P% to that of vaporization at PI.
After this it is vaporized, giving the increase of volume

>Iled Spring

1 ' \C

. J .

1 \ /*\ ' • i

WW)

\ / » ' N / '• '

v' Vv' 'v v' '

Saturation Curve

,Work done by Steam
.during expansio 6-<?

Volume
FIG. 21. — Complete Expansion Cycle or Clausi^y Cycle.

shown by ab. The supply of heat is then stopped. The
cylinder of the engine is made larger than in the preceding
type so that when the point 6' is reached the piston can
travel still further, and it is allowed to do so, that is, the

THE IDEAL STEAM ENGINE 57

high-pressure steam is allowed to push it further out. This
can be pictured by imagining the steam to act like the
compressed spring shown in the figure and to push the
piston in much the same way as does the spring. The line
b\c\ shows the decreasing pressure exerted on the piston by
the spring as the latter expands so as to get longer and
longer. Because of the properties of a spring this is a
straight line. The line be shows the decreasing pressure
exerted on the piston by the steam as the latter expands
so as to occupy greater and greater volumes. Because of
the properties of steam this line is curved instead of straight.

Work will be done on the piston by the expanding
steam during the process be and the amount of this work
will be indicated by the area under the line be as shown
in the figure. This work must have been done by the
expenditure of energy on the part of the steam and since
no energy was added after the point b was reached the work
must have been done at the expense of heat energy contained
in the steam at 6. It has already been shown that the
heat above 32° in the steam at b is equal to the sum of the
heat of the liquid and the internal latent heat, and some
of this heat must obviously be used for the doing of work
along be instead of being entirely rejected to the cooling
medium as in the preceding cycle without " expansion."

The expansion of the steam continues until the " back
pressure " P^ is reached. The cooling medium may then
be imagined to be brought into use and to abstract such
heat of vaporization as may remain in the steam besides
absorbing the equivalent of the work done on the steam
by the returning piston, thus giving the process shown by
the line cd.

If the expansion line be of the cycle just described
could be carried out within walls constructed of such mate-
rial that it would not give heat to nor take heat from
the steam, it is obvious that any heat energy lost by the
steam during the expansion could be lost only by conver-

58 STEAM POWER

sion into mechanical energy. An expansion of this kind
is called an adiabatic expansion.

In the figure, the curve of adiabatic expansion is shown
in its correct position with respect to the saturation curve
and it is obvious that for an adiabatic expansion, starting
with dry, saturated steam, the quality decreases as the expan-
sion progresses.

Comparison with Cycle without Expansion. The heat
supplied is the same in both of the cycles just considered
when they operate between the same two pressures, but
the mechanical energy obtained in the case of the complete
expansion cycle is much greater. In Fig. 21, for instance,
the mechanical energy obtainable with the cycle first
described is represented by the area abd'd while that obtain-
able with the complete expansion cycle with the same
heat supply Q\ is represented by the same area abd'd plus
the additional area bed'. The efficiency of the complete
expansion cycle is therefore very much higher than that
of the cycle without expansion.

For conditions similar to those giving a theoretical
efficiency of about 6 per cent without expansion, the com-
plete expansion cycle will give a theoretical efficiency of
about 12 per cent and this figure can be doubled by
expedients which will be considered later.

The cylinder required for the production of the com-
plete expansion cycle would be much larger than that re-
quired for the other cycle if both used the same weight of
steam per cycle. The proportion would be in the ratio
of the volume shown at c in Fig. 21 to the volume shown
at 6. But the complete expansion cycle would make avail-
able much more energy per pound of steam than would
the other, so that the difference in the size of cylinders
would not be so great if both were required to make avail-
able the same amount of mechanical energy per cycle.

40. The Incomplete-expansion Cycle. The shape of
this cycle is shown in Fig. 22. It is just like the complete

THE IDEAL STEAM ENGINE

FIG. 22. — Incomplete Expansion Cycle.

expansion cycle down to the point c. The cylinder in which

it is produced has a smaller volume than that used for the

complete expansion cycle so that the piston arrives at the

end of its stroke before it

has opened up volume

enough to enable the

steam to expand all the

way down to the lowest

pressure (terminal or back

pressure) . When the point

c is reached in the real

engine, the exhaust valve

is opened and enough

steam then blows out to

reduce the pressure to the back pressure Pa. The piston

then returns and drives out the remainder of the steam as

shown by the line de.

In the ideal method assumed in the preceding treat-
ment, the heat absorbing medium would be brought into
use at c, absorbing sufficient heat to reduce the pressure
from PC to Pd while the piston remained stationary at the
end of its stroke. The latent heat of vaporization remain-
ing in the steam at d would then be absorbed as the piston
was driven back from d to e.

Comparison with Other Cycles. The incomplete expan-
sion cycle is intermediate between the two previously dis-
cussed. This can be appreciated readily by an inspection
of Fig. 22. In this figure the area abd'e represents the
mechanical energy obtainable with the cycle without
expansion; the area abc'e represents the energy obtainable
from the same quantity of steam with complete expansion;
and the area abcde represents the energy obtainable from
the same amount of steam with incomplete expansion.

The later the point at which the exhaust valve is opened,
point c, the more nearly do efficiency and energy obtain-
able,approach the values for the complete expansion cycle „

60 STEAM POWER

The earlier the point at which the exhaust valve is opened,
the more nearly do efficiency and energy obtainable approach
the values for no expansion.

Despite the lower efficiency of the incomplete expan-
sion cycle as brought out in connection with Fig. 22 it is
universally used on all reciprocating engines excepting
those which make no pretense to economy and use no
expansion. The less efficient cycle is used for the simple
reason that complete expansion in a reciprocating engine
does not pay commercially. For complete expansion the
cylinder must be larger in the ratio of Vc to Vc> as shown
in Fig. 22 and the work obtained by completing the expan-
sion is a very small part of the total. In most cases it
would not be great enough to overcome the friction of the
engine, not to mention paying interest on the necessarily
higher cost of the larger cylinder and accompanying parts.

It will be shown in a later chapter that the steam tur-
bine can economically expand the steam completely and
the complete expansion cycle is therefore used with such
prime movers.

CHAPTER V
ENTROPY DIAGRAM

41. Definitions. In Chapter III temperature, pressure
and volume were discussed as criteria determining the con-
dition of water and steam. Other things may be used in
determining the condition of such materials. One which is
particularly useful from an engineering standpoint is known
as entropy and is designated by the Greek letter (/>.

For every condition of water and steam, there is a char-
acteristic value of entropy just as there is a characteristic
value of temperature, pressure, volume, heat above 32° F.,
etc. These values of entropy are given in the steam tables
in just the same way as the value of temperature, pressure,
volume, heat above 32° F., and such, are given.

The entropy of the liquid given for any particular pres-
sure is the change of entropy experienced by one pound
of the liquid when its temperature is raised from 32° F.
to the temperature of vaporization corresponding to that
particular pressure. It might be spoken of as the entropy
of the liquid above 32° F., just as q is spoken of as the heat
of the liquid above 32° F. It is represented by <fr.

The entropy of vaporization given for any particular
pressure is the change of entropy experienced by one pound
of the material while changing from water at the tempera-
ture of vaporization to dry saturated steam at constant
pressure. It corresponds to the latent heat of vaporiza-
tion and is designated by </>„.

The entropy of dry saturated steam at any pressure is
the sum of fa and 4>v and therefore is the total change of
entropy experienced by a pound of material in changing

STEAM POWER

from water at 32° F. to dry saturated steam at the particu-
lar pressure in question.

The entropy of superheat at any pressure and tempera-
ture is the change of entropy experienced by a pound of
dry, saturated steam at that pressure when superheated
to that particular temperature. It is designated by </>s.

The entropy of superheated stsam at any pressure and
temperature is the total change of entropy experienced by
one pound of material when changed from water at 32° F.

FIG. 23. — Temperature-Entropy Diagrams.

to superheated steam at the pressure and temperature
in question. It is equal to <&+<£„+</>«.

42. Temperature-Entropy Chart for Steam. Entropy
is particularly useful to the engineer because it enables him
to draw charts which lend themselves readily to an easy,
graphical solution of certain problems which would other-
wise involve complex calculations. . One of these charts
is known as the Temperature-Entropy Chart.

In making this chart, absolute temperature is generally
plotted on the vertical and entropy above some datum tem-
perature on the horizontal, as shown in Fig. 23 (a) and (6),
which represents the construction of a temperature entropy
diagram for water and steam. The entropy values on

ENTROPY DIAGRAM 63

this chart are plotted above 32° F. as datum tempera-
ture.

The water line or water curve is obtained by picking
out of the steam tables the values of fa, entropy of the liquid,
for different pressures and plotting them against the abso-
lute temperatures corresponding to those pressures. Ob-
viously, zero of entropy will occur at the absolute tempera-
ture corresponding to 32° F., i.e., about 492° F. abs.

The saturation curve or dry steam curve is obtained by
picking out of the steam tables the values of fa-}- <j>v for
different pressures and plotting against corresponding
absolute temperatures.

The entropy of vaporization is obviously shown for each
different temperature (or pressure) by the distance between
the water curve and the saturation curve, since the former
is distant from the vertical axis by an amount equal to fa,
while the latter is distant an amount equal to fa-}- fa.

Superheating lines are drawn by picking from the steam
tables the values of entropy above 32° F. for steam super-
heated to different temperatures at one particular pressure
and plotting against the proper temperatures. There will
be as many superheating lines on the diagram as one chooses
pressures for which to plot them. Only one is shown
in the figure.

One very useful property of this diagram follows from the
fact that points on its surface indicate the condition of the
material. For instance, if the temperature-entropy, or
T — <£, values of the material at a given condition should
plot to the left of the liquid line, the material must be in the
liquid condition; if they plot between the liquid line and the
saturation curve, the material must be a mixture of liquid
and saturated vapor; if they plot on the saturation curve,
the material must be dry, saturated steam; and if they
plot to the right of the saturation curve, the material must
be superheated steam. This all follows directly from the
definition of entropy above 32° F., as plotted in these dia-

64 STEAM POWER

grams. The various regions, or fields, into which the dia-
gram divides in this way are shown in Fig. 23 (a).

Another very useful property of this diagram follows
from the fact that area represents heat just as area on a
pressure-volume diagram was found to represent work.
Thus the area under the line ab, for instance, represents
the heat required to raise the temperature of one pound
of water from 32° F. to the temperature at b. Similarly
the area under the line be represents the heat required to
change a pound of water at the temperature at 6 to a pound
of dry, saturated steam at the same temperature. The
heat required to superheat this pound of saturated steam
at constant pressure up to the temperature shown at d
is similarly represented by the area under the line cd.

In this connection, it should be noted that this diagram
is plotted above absolute zero of temperature just as the
pressure-volume diagram is plotted above absolute zero of
pressure. The areas in question therefore extend down
to the absolute zero of temperature. In order to indicate
this in Fig. 23 (b), a large part of the chart is supposed to
have been broken out, so that the lower end of the diagram
could be moved up into view. In Fig. 23 (a), the bottom
of the diagram is drawn a few degrees below 32° F. and
this is indicated by putting T>0 opposite the horizontal
axis.

The various areas hatched in Fig. 23 (6) indicate the
various quantities of heat previously discussed. It should
be understood that the areas represent the heat quantities
only for the particular pressure which corresponds to the
temperature indicated by Tv. For a higher pressure, the
line be would be higher and the areas proportionately
larger; for a lower pressure the line be would be lower and
the areas smaller.

ENTROPY DIAGRAM

ILLUSTRATIVE PROBLEM

Starting with liquid at a temperature Tt corresponding to the
temperature of vaporization at a pressure of 50 Ibs. per square
inch absolute, assume the liquid raised to the temperature of
vaporization at a pressure of 100 Ibs. per square inch absolute
and then completely vaporized. Determine the various changes
of entropy and indicate them on a TV-chart.

The steam tables give entropy of the liquid, 4i, as equal to
0.4113 for water about to vaporize under 50 Ibs. per sq. in.
absolute, and 0.4743 for water about to vaporize under a pres-
sure of 100 Ibs. per sq. in. absolute. The difference, that is,
0.4743-0.4113=0.0630, must be
the entropy change experienced
by the liquid when its tempera-
ture is raised from the lower to
the higher value. These values
itre shown in Fig. 24.

The steam tables give entropy
of vaporization, <£», at 100 Ibs. per
square inch absolute as 1.1277.
Adding this to the entropy above
32° F. of the liquid at vaporiza-
tion temperature under 100 Ibs.
pressure gives 0.4743+1.1277 =
1.602 as the entropy above 32°
of dry, saturated steam at 100

Ibs. per square inch absolute. These values are all indicated in
their proper position in Fig. 24.

The total change of entropy experienced by the material in
changing from water at the temperature of vaporization under
50 Ibs. pressure to dry, saturated steam at 100 Ibs. pressure is
obviously equal to 0.0630+ 1.1277 =1.1907.

43. Quality from T^-chart. The entropy change ex-
perienced by steam in the process of vaporization is directly
proportional to the addition of heat. Thus, when half
the latent heat has been added to one pound of material,
the entropy change is J0». In general, if a fraction x of
the latent heat has been added, the entropy change has
been x<t>f during the process. Therefore, if the temperature
entropy condition of a pound of material should plot at a

FIG. 24.

STEAM POWER

point such as c in Fig. 25, it follows that the material is
a mixture of water and steam and that a fraction of the

pound equal to — is steam, the rest being water. But,
oct

be
by definition, the fraction — is x, the quality of the

material.

The temperature-entropy chart is very useful when used
in connection with this property of showing quality. Thus,
in Fig. 25, the area under be, down to absolute zero tem-
perature, represents the fraction of the latent heat of

Entropy

FIG. 25. — Quality irom Temperature- FIG. 26. — Constant Quality
Entropy Chart. Curves.

vaporization per pound which must be added to give a
pound the quality x.

For convenience in use, constant quality lines are
generally drawn on temperature-entropy charts. Such
lines are shown in Fig. 26. Each line is obtained by plot-
ting the temperature entropy conditions for a given quality
at different pressures. For this purpose, <f>v and <fr are
taken from the steam tables for a given pressure. The
numerical value of <j>v is then multiplied by the fraction re-
presenting the chosen quality, say 0.9, and the product
is added to <£/, giving the total entropy above 32° F. for
quality 0.9 at the particular pressure chosen. The same

ENTROPY DIAGRAM

68 STEAM POWER

process is repeated with the same value of the quality,
but with different pressures, until enough points have been
secured to make it possible to draw a smooth line through
them.

44. Volume from T0-chart. Since quality changes at
any given temperature, or pressure, are accompanied by
volume changes, it is possible to find a series of values for
the quality of a pound of wet steam which will make that
pound occupy the same volume at different temperatures.
Having found the quality which will be necessary at a num-
ber of different temperatures, the total entropy above 32°
F. can be found for each case and these values can then be
plotted on the T^-chart. Connecting the points so obtained
would give what is known as a Constant Volume Line.

Several of these constant volume lines are shown in their
correct positions in Fig. 27. It will be observed that,
for each volume, the quality must increase as temperature
(and pressure) increases in order to maintain a constant
value for the volume occupied by one pound of mixture.

45. Heat from T0-chart. Equations for obtaining the
total heat above 32° F. for wet and for superheated steam
were given in an earlier chapter. By means of these
equations, it is possible to find a succession of values for
quality and superheat which will give a pound of material
any chosen heat content at different pressures. If the
corresponding values of temperature and entropy are found
and plotted, what is known as a Constant Heat Line results.
Several of these lines are shown in Fig. 27.

46. The Complete T(£-chart for Steam. A very com-
plete, graphical representation of the properties of water
and steam can be procured by combining in one diagram
all of the lines discussed in preceding paragraphs. Such
a diagram is generally spoken of as the T ^-diagram or the
T<f>-chart for steam. An example of such a diagram is
given in Fig. 28.

This chart is very useful, as it enables one to solve by

•ui -bg 'sqi) oanssaaj a^nTOsqy

70 STEAM POWER

inspection many of the most difficult problems which arise
in the theory and practice of using steam. As an example,
assume that it is desirable to know what will happen if
water at the temperature of vaporization corresponding
to about 24 Ibs. per square inch absolute has its volume
increased indefinitely at constant temperature. The initial
condition of the water would be shown on the water curve
of Fig. 28 at the point at which the 700° absolute temperature
line crosses it. Increase of volume at constant temperature
would be indicated by a horizontal line running to the right
from this point. Obviously, vaporization will occur at
constant pressure (because the temperature is constant)
and the quality will change from zero to unity at which the
saturation curve will have been reached. Further increase
of volume can result only in the production of superheated
steam, since the line representing the process will rur\
out into the superheated steam field. It is also interesting
to note that the pressure on the material will have to be
decreased as the volume increases in the superheated steam
region, as is evidenced by the fact that the horizontal line
representing the assumed process cuts lower and lower
pressure lines as it is extended to the right in the super-
heated field.

Note also that the intersections of this horizontal line
with constant volume and constant heat lines afford the
means of determining volume and heat above 32° F. at
different stages of the assumed process.

PROBLEMS

1. Determine from the steam tables the change of entropy
experienced by one pound of water when its temperature is raised
from 32° F. to the temperature of vaporization under a pressure
of 100 Ibs. per square inch absolute.

2. Determine from the steam tables the entropy change experi-
enced by one pound of water when its temperature is raised from
32° F. to the temperature of vaporization under a pressure of
150 Ibs. per square inch absolute.

ENTROPY DIAGRAM 71

3. Determine the entropy change experienced by one pound of
water when its temperature is raised from the temperature of
vaporization corresponding to 100 Ibs. per square inch to that
corresponding to 150 Ibs. per square inch by subtracting the value
found in Prob. 1 from that found in Prob. 2.

4. Determine the change of entropy experienced by one pound
of material completely vaporizing at a temperature of 327.8° F.

5. Plot a TV-chart for one pound of water. Start by plotting
entropy of the liquid for various temperatures; then plot entropy
of saturated steam (above 32° F.); finally draw water line, satura-
tion line, and several lines showing change of entropy during vapor-
ization.

6. Determine from a TV-chart the quality which would be
attained by one pound of steam if it experienced a change which
carried it from the condition of dry saturated steam at 150 Ibs.
per square inch absolute to a pressure of 25 Ibs. per square inch
absolute by a process which would plot as a vertical line on the

TV-chart.

7. Assume a pound of mixed water and steam to have a qual-
ity of 80% at a pressure of 200 Ibs. per square inch absolute.
Determine from the TV-chart the heat above 32° per pound of
mixture and the volume occupied by the mixture. Determine
also the quality attained if the pressure of the material drops to
20 Ibs. per square inch absolute at constant entropy. How does
the heat above 32° F. change during such a process?

8. Assume a pound of mixture as in Prob. 7, but with a
quality of 30% at a pressure of 200 Ibs. Find all quantities called
for in that problem.

9. Assume a pound of material as in Probs. 7 and 8 above,
but superheated 200° at a pressure of 200 Ibs. per square inch
absolute. Determine all quantities called for in Prob. 7.

10. Choose a point on the TV-chart at which a constant volume
line intersects the saturation curve. Determine the change of
quality, entropy and heat above 32° F., if the material drops
to half pressure at constant volume.

CHAPTER VI

TEMPERATURE ENTROPY DIAGRAMS OF STEAM
CYCLES

47. Complete Expansion Cycle. This cycle was con-
sidered in Chapter IV and the PF-diagrarn was given there
as Fig. 21. The diagram of this cycle drawn to T^-co-
ordinates is shown in Fig. 29. The same letters are used

to represent corresponding points
in the two diagrams.

The entropy change during
the heating of the liquid is
shown by the part of the liquid
line between d and a, and the
heat supplied during that process
is represented by the area below
the line da, measuring clear
down to the absolute zero of
temperature.

The entropy change during

FIG. -9. — 7>-diagram, Com-
plete Expansion Cycle.

vaporization is represented by the line ab and the heat
supplied during the process is shown by the total area
under that line.

The adiabatic expansion of the steam is represented by
the line be, such an adiabatic change fortunately being a con-
stant entropy process and therefore easily drawn in this
diagram. Obviously no heat is received or removed dur-
ing this process, as there is no area under the line be.

The entropy change during condensation is represented
by the line cd and the heat rejected by the working sub-
stance during this process is represented by the area under
that line.

TEMPERATURE ENTROPY DIAGRAMS 73

48. Area of Cycle Representative of Work. It will
be remembered that area under a line in the PF-diagram
represents work in foot-pounds. That diagram, however,
gi^es no indication of heat received or rejected and it is
not possible to obtain any direct idea of efficiency from it.
In this respect, the TV-diagram is much better. Area under
the lines da and ab in Fig. 29 represents heat supplied
the working substance. Area under the line cd represents
heat rejected by the working substance. The difference
between these two, or the area enclosed within the lines
of the cycle, must therefore represent the heat converted
into mechanical energy per cycle.

This diagram therefore shows directly by areas the
heat supplied, the heat rejected, and the heat converted
into mechanical energy. Further, the ratio of the area
representing heat converted into work, and the area repre-
senting heat supplied must be the efficiency of the cycle.

Remembering also that if the lines of the cycle are drawn
upon a T^-chart such as that given in Fig. 28, all volume
changes, heat contents and qualities at different points
are shown without further work, it becomes evident that
this form of representation is decidedly convenient and far
superior to the pressure volume method.

49. Modifications for Wet and Superheated Steam.
The complete expansion cycle is supposed to represent an
idealization of what happens in a real prime mover. In real
cases, however, the steam may arrive at the prime mover wet
or superheated and it is desirable to investigate the method
of representing such conditions as well as their effects.

Wet steam corresponds to incomplete vaporization,
i.e., a quality less than unity at the upper right-hand corner
of the cycle. This might be shown for a given case by the
location of the point 6' in Fig. 29. The cycle would then
be ab'c'd and a smaller amount of work would be obtained
per pound of working substance as evidenced by the smaller
area enclosed within the lines of the cycle.

STEAM POWER

In the case of superheated steam, superheating occurs
at constant pressure .after vaporization is complete. This
would be shown by the location of the upper right-hand
corner of the cycle at some point 6" on the constant pressure
line which extends out from b. The cycle is now represented
by abb"c"d and evidently has a different shape than it
had in the preceding cases. Obviously the area enclosed
within the lines of the cycle is greater than it was before and
therefore more mechanical energy is obtained per pound
of steam.

50. Incomplete Expansion Cycle. The only difference
between the incomplete and complete expansion cycles is

FIG. 30. — TV-diagram, Incom-
plete Expansion Cycle.

FIG. 31. — 7'0-diagram, Cycle
Without Adiabatic Expansion.

the termination of the expansion in the former by means of
a constant volume line. This is shown to TV-coordinates
in Fig. 30 in which the incomplete expansion cycle is drawn
in heavy lines over the one in which expansion continues
to the back pressure.

The constant volume line is seen to cut off a corner,
thus reducing the area representing heat converted into
work. The heat supplied in each case is measured by the
area under the lines ea and ab. The efficiency of the cycle
with incomplete expansion can therefore be seen to be less
than that of the other cycle by simple inspection of the
diagram.

If the adiabatic expansion is terminated at a higher

TEMPERATURE ENTROPY DIAGRAMS 75

pressure, as by the constant volume line c"d" in Fig. 30,
still more of the work area is lost, but the same quantity
of heat is supplied, and therefore the efficiency is still lower
than when the expansion terminated at c. Obviously
as the point at which the adiabatic expansion is terminated
moves nearer and nearer to b as shown in Pig. 31, the cycle
becomes less and less efficient. If the constant volume
line starts at b, there is no adiabatic expansion and the
cycle becomes that previously considered as having a rec-
tangular shape in the PF-diagram. This cycle has the shape
indicated by abed in the TV-diagram of Fig. 31. Obviously
it is least efficient of all as was previously shown by other
means.

51. Effect of Temperature Range on Efficiency. It has
already been stated (see p. 26) that heat engines receive
heat at a high temperature, convert some of it into me-
chanical form and discharge the remainder at a lower tem-
perature. Inspection of the TV-diagram shows this very
clearly, and, remembering that the area of the cycle measures
the heat converted, these diagrams also show how raising
the upper temperature (or pressure) or lowering the lower
temperature (or pressure) will increase the efficiency. It
can be seen readily that lowering the lower temperature
will, however, be more effective in increasing the efficiency
than raising the upper temperature.

PROBLEMS

4- 1. Draw a complete expansion cycle to T ^-coordinates for the

following conditions (using TV-diagram for steam to get values);

weight of working substance, 1 lb.; initial pressure, 125 Ibs.

absolute; quality at beginning of adiabatic expansion, 100%

back pressure, 10 Ibs. absolute.

2. Determine the following values for cycle drawn in Prob. 1 :
(a) Entropy of liquid at beginning of vaporization;
(6) Entropy at beginning of adiabatic expansion;

(d) Volume at end of adiabatic expansion-

(e) Entropy at end of condensation.

76 STEAM POWEP

-T 3. Show by measuring the area on TV-diagrams, the increase of
efficiency resulting from the use of an initial pressure of 175 Ibs.
absolute and from the use of a terminal pressure of 2 Ibs. absolute
in place of the values given in Prob. 1.

•^ 4. Compare the efficiency of a complete expansion cycle with
conditions as in Prob. 1 with a complete expansion cycle with
same pressures but with a temperature of 500° F. at the beginning
of the adiabatic expansion.

5. Draw an incomplete expansion cycle to T ^-coordinates
for the same pressures as in Prob. 1, but with adiabatic expan-
sion ending at a pressure of 15 Ibs. absolute.

6. Compare work and efficiency of the two cycles of Probs.
1 and 5 above.

7. Draw a cycle without expansion for the conditions of
Prob. 1 to TV-coordinates and compare the work area with that
obtained in Probs. 1 and 5.

CHAPTER VII
THE REAL STEAM ENGINE

52. Operation of Real Engine. In previous chapters
the ideal steam engine was considered and several cycles
upon which it might be operated were discussed. Real
engines are built to operate on the same cycles, but because
of certain practical considerations they only imperfectly
approximate the ideal performance.

Real engines must be built of iron and steel for practical
reasons and these metals absorb, conduct and radiate heat
so that certain heat interchanges between the working
substance and engine and certain heat losses occur in
practical operation. These were eliminated in the ideal
case by simply assuming ideal materials not possessed
of the characteristics of real metals.

It is also practically impossible to generate steam
in the cylinder of a real engine as was assumed to be done
in the ideal case. Heat is practically obtained by the com-
bustion of fuels, and the higher the temperature attained
the better can the liberated heat be utilized in the genera-
tion of steam. To subject the cylinder to such high tem-
peratures and to control the heating and cooling as neces-
sary to produce a number of cycles in rapid succession would
lead to rapid wear and great practical difficulties. It has
been found best to generate the steam in a boiler which is
properly equipped for that purpose and then to transmit
it with its contained heat to the engine, which is constructed
in such a way as to utilize that heat to the best advantage.
If the steam is to be condensed, as assumed in the ideal
cases, it has also been found best to remove it from the

STEAM POWER

cylinder and to condense it in a separate piece of apparatus
properly constructed for that purpose.

The entire arrangement which results from these prac-
tical modifications in the case of a non-condensing engine

THE REAL STEAM ENGINE 79

is shown in Fig. 32. Steam is generated within the boiler
at some constant pressure PI and at the proper instant the
admission valve at one end of the cylinder is opened, allow-
ing steam to flow in and drive the piston outward. If
there were no losses, this would be represented by some
such line as ab at a height PI on the
PF-diagram of Fig. 33. Closing of pi
the valve after the piston had moved
part way out would cut off the
further flow of steam, and, with con-
tinued motion of the piston, the steam
within the cylinder would expand. FIQ 33

If no heat interchanges occurred,
this expansion be would be adiabatic as in the real case.

It will be observed that the two lines on the PF-diagram
thus far produced represent equally well the corresponding
two lines of the complete or incomplete-expansion cycles.
The heat supplied in the boiler is the same as that supplied
in the cylinder under the ideal conditions originally assumed,
and the work under the line a b is equal to the external
work done during vaporization just as in the ideal case.
If difficulty is experienced in connection with the statement
regarding external work, it is only necessary to picture the
process in this way: Assume that each pound of steam
formed in the boiler does the external work equivalent to
A Pu by pushing the pound previously generated ahead of
it as a piston, and that this motion communicated along
the pipe from layer to layer results in pushing an equivalent
weight (and volume) into the cylinder against the resist-
ance offered to the piston's motion.

When the piston arrives at the end of its stroke at the
point c, the opening of the " exhaust valve," connecting
the interior of the cylinder with the space in which the
pressure P^ lower than Pc is maintained, will permit some
of the steam to blow out of the cylinder with the piston
standing stationary at the end of its stroke. This would

gO STEAM POWER

give a constant volume change equivalent to the correspond-
ing line in the incomplete-expansion cycle.

The return of the piston from d to e, with the exhaust
valve still open, would force the remainder of the steam
out of the cylinder and into the space in which the pres-
sure P2 is maintained. The result, so far as the diagram
is concerned, is obviously the same as in the ideal case, and
if the steam were condensed within a proper vessel into
which it exhausted (instead of being exhausted to atmos-
phere), the result would also be the same so far as the shape
of the diagram is concerned. The pressure P^ might, how-
ever, be maintained at a lower value, thus giving a greater
temperature range.

The pressure rise ea within the cylinder would result
directly from the opening of the admission valve and the
admission of steam for the next cycle. But, if the working
substance is to be returned to starting conditions as was
dene in the ideal case, its pressure must also be raised to
Pi and its temperature to a corresponding value. The
pressure is raised in the case of condensing operation by
means of the boiler feed pump, which picks up the condensed
steam (condensate) and forces it into the boiler. The
temperature of the working substance is raised by passing
it through feed-water heaters or by heat absorbed directly
from the heated water in the boiler.

When operating non-condensing the working substance
exhausted during the last part of each cycle is really thrown
away by allowing it to mix with the atmosphere, but an
equivalent quantity of water is fed to the boiler by the
boiler feed pump and takes the place of the material lost
by exhaust to atmosphere. This method of operating
does not approximate the ideal as closely as does the con-
densing method, but the discrepancy is not very great.

53. Losses in Real Installations. The diagram given
in Fig. 33 was obtained by assuming the absence of certain
practical losses and is considerably modified when real

Adlatattc Ezpanilon
of Mixture at Point
of Cut-off.

Adiabatic Expansion of
all Material in Cylinder
if Initially Dry and

Back Pressure

^Saturated.
C

VoL

THE REAL STEAM ENGINE 81

apparatus is used. Thus the real engine, as shown in con-
nection with Fig. 9, has clearance and operates with com-
pression so that the clearance is filled with steam at a pres-
sure indicated by the point a' in Fig. 34 when the admission
valve opens.

There is also always some drop of pressure along the
steam pipe so that the pressure
at the engine is lower than at the
boiler. Further, the admission
valve can never be made to give
such a large opening into the
cylinder that there is not a
measurable drop of pressure in FlG. 34.-Theoretical and Real
flowing through it. As a result Indicator Diagrams,

of these actions the highest

pressure attained within the cylinder as indicated at point
a in Fig. 34 is always lower than the boiler pressure PI.

As the piston of a real engine moves out it acquires
a higher and higher velocity until it reaches a point near
mid-stroke. The entering steam therefore must flow
through the valve with increasing velocity if it is to follow
up the piston and fill the cylinder, but this usually neces-
sitates greater pressure drops as the piston moves out,
so that the admission line generally slopes downward instead
of being horizontal. There is also another phenomenon
which causes this line to slope. The metal of cylinder,
cylinder head and piston is in contact with comparatively
low-temperature steam during the latter part of each cycle,
and therefore acquires a lower temperature than that
of the steam about to enter. Therefore, when the high-
pressure steam enters the cylinder it gives up heat to
the walls at a comparatively rapid rate, and, if initially
dry saturated, this results in a great deal of condensation.
Such condensation is called initial condensation.

As the steam condenses after flowing into the cylinder
and forms water occupying a negligibly small volume,

82 STEAM POWER

it follows that steam must flow into the cylinder at a pro-
portionately greater rate in order to fill the space vacated
by the piston. But this results in an increased pressure
drop and therefore would give a sloping admission line.

When the piston has finally been driven out as far as
desirable by the action of high-pressure steam, the admis-
sion valve is closed, that is, cut-off occurs. This valve can
not be closed suddenly; the closure is more or less gradual
in all cases. As the opening becomes smaller it becomes
increasingly more difficult for the steam to flow through
and into the cylinder so that the pressure continues to drop
at an increasing rate until the valve is finally closed. This
gives the rounded cut-off shown at the point b.

The loss of pressure during admission is generally
said to be due to throttling or wire drawing, these terms
being intended to convey the idea that the steam has to
squeeze its way through the inlet openings with correspond-
ing loss of pressure.

When the cut-off has finally been completed, it leaves
the end of the cylinder filled with a mixture of steam and
water at steam temperature, and this mixture then expands
as shown by the line be. At the beginning of the expansion
the steam generally has a higher temperature than that of
the surrounding walls and it therefore continues to give
heat to those walls. Were the expansion adiabatic it
would follow the clot-dash line in the figure, but, as the
steam must not only convert heat into work, but must also
supply heat to the walls, it condenses more rapidly than
in the ideal case and its pressure and volume changes follow
some such law as that indicated by the upper part of the
curve be.

As expansion continues, the pressure and temperature
of the steam drop until some point is reached at which
the temperature has become equal to that of the walls.
Further expansion with drop of pressure and temperature
results in reducing the temperature of the steam below

THE REAL STEAM ENGINE 83

that of the walls, and then the direction of heat transfer
is reversed, the hot walls giving heat to the cooler steam
at an increasingly rapid rate. This heat causes re-evapora-
tion of some of the water formed before and thus tends to
increase the volume occupied by the material in the cylinder,
with the result that the lower part of the expansion curve
be approaches and generally crosses the curve which would
have been attained by adiabatic expansion in non-conduct-
ing apparatus.

In many real engines the re-evaporation is so great that
the steam is entirely dried and sometimes superheated
before the exhaust valve opens.

The exhaust valves of steam engines are always opened
before the piston reaches the end of the stroke, as it is found
necessary to do this if excessive losses are not to occur
due to the difficulty of forcing the large volume of low-
pressure steam through the exhaust passages. When opened
early enough, the steam flows out in such quantity before
the end of the stroke that the " back pressure " during the
return or exhaust stroke is only a pound or two above that
of the space into which the engine is exhausting.

During all of the exhaust period, the steam is probably
at a lower temperature than the walls to which it is exposed
and re-evaporation probably continues in most cases until
the closure of the exhaust valve. It seems probable that
the steam retained in the cylinder after the closure of the
exhaust valve is approximately dry, but little is really
known regarding the quality of the clearance steam.

The rise of pressure during compression has two bene-
ficial effects: It helps to bring the moving parts to rest grad-
ually, and it raises the temperature of the clearance steam
and of the walls of the clearance space to values nearer
that of the entering steam.

Remembering that area on a PF-diagram represents
work, it is easily seen that throttling losses and rounding
of corners due to slow valve action (which cause a loss of

84 STEAM POWER

diagram area) result in a loss of work. The fact that conden-
sation also causes a great loss is easily shown. A given
quantity of steam entering the engine with its supply of
heat can, in the ideal case, do a certain amount of work at
the expense of that heat. In the real case part of the heat
is stored in the walls during the early part of the cycle,
so that it is not available for the doing of work and is removed
from the walls and carried out into the exhaust as unutilized
heat during the later part of the cycle. The phenomenon
can be pictured by imagining the steam as dropping some
of its heat into a pocket in the walls of the cylinder when
entering the engine and then picking it up again and carrying
it out when leaving, so that the next charge of steam will
have to fill the pocket again.

The net result of condensation and re-evaporation
is the obtaining of less work from a given quantity of steam
than should be obtained, or the use of more steam than
theoretically necessary for a given quantity of work. This
effect is shown graphically by the two adiabatic expansion
lines of Fig. 34.

The initial condensation in real engines which are sup-
plied with saturated steam generally amounts to from 20 to
50 per cent of all the steam supplied, so that it is evident
that anything which will prevent part or all of this loss should
do much to improve the steam consumption of engines.
This subject will be discussed in more detail in later para-
graphs and various methods of decreasing losses from this
source will be considered.

54. Clearance. The term clearance is used in a two-
fold sense; (a) to refer to mechanical clearance or the
linear - distance between the two nearest points of cylinder
head and piston face when the piston is at the end of its
stroke, and (b) to refer to volumetric clearance or the
volume enclosed between the face of the valve, the cylinder
head and the face of the piston when the latter is at the
end of its stroke.

THE REAL STEAM ENGINE

yalve

The former is generally given in inches and varies
from a very small fraction of an inch in the best engines
to an inch or more in cheap and in poorly designed engines.
It is indicated by a in Fig. 35.

The volumetric clearance is expressed as a percentage
of the piston displacement or steamPort steam Chest
volume swept through by the
piston. It varies from 2 per
cent or less in the best engines
to as high as 15 per cent in
the cheaper and less economical
models. It is made up of the
parts designated by c in Fig. 35.

55. Cushion Steam and Cyl
inder Feed. It is customary
to imagine the steam operating
within an engine cylinder to FlG" 35. -Mechanical and Volu-

. , ,. , 7 . metric Clearances,

consist oi two parts, the cushion

steam and the cylinder feed. The former is that part of
the total which is contained in the clearance space before
the admission valve opens and serves to cushion the
reciprocating parts of the engine. The cylinder feed
is the steam which enters through the valve for each
cycle.

If the same cycle is produced time after time so that all
temperature effects are repeated at regular intervals and
so that all events occur at the same points in successive
cycles, the quantity of steam retained in the clearance
volume will be the same for successive cycles. It is
impossible to measure the quantity of this steam directly
and indirect methods are therefore adopted for that
purpose.

It is often assumed that the steam is dry and satu-
rated when compression begins, as at the point e in Fig.
34. With this assumption, the weight of cushion steam
can be determined by dividing the volume occupied, that is,

86 STEAM POWER

Ve, by the volume occupied by one pound of dry saturated
steam at the same pressure. Thus,

Cushion steam = 5- —^ Ibs. . (29)

Sp.vol. at pressure Pe

The weight of cylinder feed can be very accurately
determined by condensing and weighing the steam leaving
the engine in a given time and dividing by the number of
cycles performed during the same period. It can also be
determined by metering the steam entering the engine
or by measuring the water fed to a boiler supplying only the
engine in question. An approximate determination of the
quantity of the cylinder feed can also be made directly from
an indicator diagram by determining what is known as the
diagram water rate. This will be considered in detail at
a later point.

When cushion-steam and cylinder-feed have both been
determined, the weight of steam contained in the cylinder
between cut-off and release can be found by adding the two
quantities. Thus,

W = Wf+WK, (30)

in which

W = total weight of steam expanding in cylinder per cycle ;
Wf= weight of cylinder feed per cycle; and
WK = weight of cushion steam per cycle.

The volume which the mixture would occupy if dry
and saturated at any given pressure can be determined by
multiplying W the total weight by the specific volume for
that particular pressure.

56. Determination of Initial Condensation. The loss
due to initial condensation is so important that it is cus-
tomary to determine the amount of this loss when studying
engines. This can be done with fair accuracy by means
of the indicator diagram.

THE REAL STEAM ENGINE

To make such a study it is necessary to know the total
weight of material in the engine cylinder at the point of
cut-off. This weight may be determined by any of the
methods just given. With the weight known, the volumes
which this material should occupy at different pressures
if dry and saturated can be determined by jnultiplying by
the specific volumes at the various pressures. Plotting
these points on a PF-diagram and connecting them will
give a saturation curve for the material in the cylinder such
as the curve shown in Fig. 36.

By drawing this curve on the indicator diagram ob-
tained from the engine and then comparing distances
such as ab and ac as explained in section 26 of Chapter III

FIG. 36.

FIG. 37.

the quality of the steam within the cylinder at all pressures
between cut-off and release can be determined. The weight
of initial condensation (up to the point of cut-off) must
be the total weight of water shown as existing within the
cylinder at that point minus any water brought in by the
steam if it was not dry when entering the engine.

Should the saturation curve cross the real expansion
curve, as shown in Fig. 37, it indicates that the steam oc-
cupies volumes greater than the specific volumes toward
the end of the expansion; the steam within the cylinder
must therefore be superheated during this part of the
cycle.

Many formulas have been devised for giving the quan-
tity of initial condensation. They are all based upon the
results of experiment and generally only give reliable

88 STEAM POWER

values for cases similar to those used in developing them.
One formula of this sort which has been very widely tested
and been found to give reliable results within its field
of applicability is that devised by Robert C. H. Heck and
explained in his books on the steam engine. The formula is

-, (31)

VA v Pe

in which

m = the fraction representing initial condensation; for
ordinary cases it is the fraction of the cylinder
feed which is condensed during admission, but when
compression is very high and when great weights
of steam are retained in the clearance it is the frac-
tion of all the material within the cylinder which
exists in liquid form at the time of cut-off;
£ = a coefficient, which varies between 0.25 and 0.30 with
type of engine. May be assumed at 0.27 for average
work;

N = engine speed in revolutions per minute (R.P.M.);
s = a constant for any engine, equal to nominal surface in
square feet divided by nominal volume in cubic
feet. The nominal surface is the area of the inner
walls and the ends of a cylinder with diameter equal
to the internal diameter of the cylinder and with
a length equal to the stroke of the engine. The
nominal volume is the cubic contents of such a
cylinder;

s = — ( 2-^— }-4 ) in which D and S represent diameter and

stroke of engine in inches;
= a temperature function obtained from Table II as

there indicated;
= the absolute pressure in cylinder in pounds per square

inch just after completion of cut-off;

THE REAL STEAM ENGINE

e = cut-off ratio, that is, ratio of cylinder volume opened
up by time cut-off has just been completed to the
total piston displacement.

TABLE II
FOR FINDING VALUES OF 0 FOR USE IN HECK FORMULA

= k\— kz when ki and fa are chosen from table for highest and lowest pressures
existing in cylinder

175

210

269.5

321.5

160

389

230

441

179

220

277

100

332.5

170

397

240

447.5

183

229

284

110

343

180

405

250

454

186

238

291

120

353

190

413

260

460.5

191

246

297.5

130

362.5

200

420

270

467

196

254

304

140

371.5

210

427

280

473

200

262

310

150

380.5

220

434

290

479

57. Methods of Decreasing Cylinder Condensation.

Before discussing methods of decreasing the loss due to cylin-
der condensation it will be well to consider what things may
be expected to determine the extent of such loss. The
. condensation is due directly to the transfer of heat from
one body to another at lower temperature, and anything
which tends to increase the total amount of heat thus trans-
ferred will increase the total condensation.

It is therefore evident that the ratio of steam condensed
to steam supplied will be greatest when:

(a) The time of contact is greatest;

(6) The ratio of surface exposed to volume enclosed is
greatest, and

The time of contact can be controlled to a certain
extent by controlling the speed of the engine and, with
other things equal, the higher the speed the lower should
be the condensation.

The ratio of surface exposed to steam to the volume
occupied by steam has a great influence on the amount of

90 STEAM POWER

condensation which occurs. The surface of the clearance
space, including the interior surfaces of all ports or passages
leading to the valves, seems to have the greatest influence,
and the clearance space which is most nearly a short cylinder
without connected passages may be expected to give the
least h itial condensation.

The size of the engine is also important in this connec-
tion. The area exposed does not increase as rapidly as
does the volume inclosed when the diameter of a cylinder is
increased, and therefore large cylinders give smaller ratio
of surface to volume and therefore a smaller percentage
of steam condensed. Large engines thus have a decided
advantage over small engines.

The shape of the cylinder also has an effect. The
longer the cylinder with respect to its diameter the more
favorable its performance.

The point at which cut-off occurs is also intimately
connected with the condensation loss. In a given cylinder
with a given clearance the total condensation within the
clearance space may be assumed practically constant if
speed and temperature remain about the same. But if
the cut-off is made later larger quantities of steam are
admitted per stroke, and hence the fraction of the total
cylinder feed which is condensed decreases.

The temperature differences depend on upper and
lower pressures, that is, on the pressure range. The inner
surfaces of the walls follow as rapidly as possible the tem-
perature changes of the steam within them. Thus their
average temperature is somewhere between the upper and
lower temperatures of the steam. If now, with a given
upper steam pressure and therefore temperature, the lower
pressure be reduced, the average wall temperature also will
be reduced, and therefore the differences between the
temperature of the entering steam and the average tem-
perature of the walls will be increased with a resulting in-
crease in condensation loss.

THE REAL STEAM ENGINE 91

The methods of decreasing this loss can now be con-
sidered. They are given below under separate heads
with brief explanation when necessary.

(a) Clearance spaces should be properly designed so
that the minimum surface is exposed.

(6) The propoitions of cylinder (diameter and stroke)
and the speed of tlie engine should be so chosen that the
condensation loss is reduced to a minimum.

(c) The engine should be so proportioned that when
delivering its rated power the cut-off occurs at such a point
as to make the percentage of cylinder condensation the
minimum consistent with other requirements.

(d) The cylinder should be surrounded by spaces filled
with air or by materials which are poor conductors of heat
so as to decrease loss by radiation, because all heat lost in this
way must be supplied by the condensation of steam within
the cylinder. Such metallic parts as cannot be " lagged"
in this way should be polished because polished surfaces
radiate less heat than dull surfaces under like conditions.

(e) The cylinder may be surrounded by a steam jacket,
that is, a space filled with steam similar to that supplied
the cylinder. The use of such a jacket sometimes results
in a considerable saving and at other times in a great loss.
The cylinder proportions, speed and pressure range seem
to be the determining factors, and most long-stroke cylin-
ders operating at low rotative speed and with small pressure
ranges are jacketed.

(/) The engine may be compounded, that is, the expan-
sion of the steam may be made to occur in two or more
cylinders taking steam in series. This results in decreas-
ing the pressure range in each of the cylinders and effects
a decided saving under proper conditions. Compounding
will be considered in detail in a later chapter.

(g) The engine may be supplied with superheated steam.
If the steam is sufficiently superheated it can give up part
or all of its superheat to heat the cylinder walls, and thus no

92 STEAM POWER

condensation need occur. Heat interchanges between
metal and superheated steam also appear to be less rapid
than is the case when the steam contains water, so that a
saving results from this source also.

Tests made with saturated and with superheated steam
indicate that from 7° to 10° of superheat are generally
required to prevent 1 per cent, of initial condensation.
Results differ greatly with the character of the engine, with
its economy on saturated steam, with its valve gear, etc.
Superheats of from 25° to 50° can generally be used with
any well-designed engine, but higher temperatures usually
require specially constructed engines. Superheats as high
as 150° F. above saturation temperature are now quite
common, and there seems to be a tendency to consider a
value between 150° and 200° as the highest that is com-
mercially advisable under ordinary conditions.

58. Classification of Steam Engines. Steam engines,
are classified on many different systems, the one used in
any particular case being determined largely by circum-
stances. The principal methods of classification are indi-
cated in the following schedule :

CLASSIFICATION OF STEAM ENGINES

fLow speed

On the basis of rotative speed \ Medium speed

[High speed

On the basis of ratio of stroke to diameter \

Short stroke

On the basis of

D-slide valve

Balanced slide valve
Slide valve ,T ,,. ...

Multiported slide valve

Piston valve

valve gear

„ .. I Drop cut-off

Corliss valve i „ ... .

[Positively operated

Poppet valve

THE REAL STEAM ENGINE

[Vertical
On the basis of position of longitudinal axis I Inclined

[Horizontal
Single expansion or
Simple engine

f Compound expansion
| Triple expansion
I Quadruple expansion

On the basis of
number of cyl-
inders in which
steam expands

Multi-expansion
engine

On the basis of cylinder arrangement

On the basis of use

Single cylinder
Tandem compound
Cross compound
Duplex

Stationary engines

Portable engines

Locomotive engines

Marine engines

Hoisting engines

59. Rotative Speeds and Piston Speeds. High-speed
engines operate at a comparatively high rotative speed and
are characterized by a short stroke in comparison with the
diameter of the cylinder, the stroke generally being equal
to, or less than, the diameter. The piston speed, by which
is meant the feet travelled by the piston per minute, generally
falls between 500 and 700.

It is not considered advisible to allow piston speeds of
stationary steam engines to exceed about 750 feet per
minute for ordinary constructions and the great majority
of engines give much lower values. The piston speed will
obviously be given by the formula

in which

= 2LN,

(32)

piston speed in feet per minute;
stroke in feet ; and

N = revolutions per minute,

STEAM POWER

and it is evident from this formula that as the rotative
speed is increased the piston speed will increase unless
the length of stroke is proportionately decreased. As
a result, high-speed engines have short strokes in com-

o/O

350
325
300

275
250
225
800 o 200
700 | 175
600 rg 1-50

500^ 125

1
400 £ 100

IGH

Ac>

, t

|»C

*~~-

V0<

'**"

(

•"3*

"X1

8 10 12 14 16 18 20 22 24 26
Diameter of Cylinder (Inches)

FIG. 38. — Proportions of High Speed Engines.

parison with their cylinder diameters and slow-speed
engines have long strokes.

The characteristic relations between cylinder diameter
and stroke, rotative speed and piston speeds of high-speed
engines are given in Fig. 38.

High-speed engines are generally fitted with some

THE REAL STEAM ENGINE

R. P. M.

Stroke (L) in Inches ]

96 STEAM POWER

form of balanced slide valve, and are controlled by what
is called a shaft governor. They are very compact, having
small weight and occupying small space in comparison with
the power developed.

Slow-speed engines are, in general, the most economical
and are characterized by low rotative speed, long stroke,
and elaborate valve gear. The weight per horse-power
is high, and they generally occupy a great deal of space.
The Corliss engine is the best known and most widely built
engine of this type.

The characteristic relations of cylinder diameter to
stroke, rotative speed and piston speed for slow-speed
engines are given in Fig. 39.

Medium-speed engines generally operate at rotative
speeds between 150 and 250 R.P.M. They are generally
fitted with the better forms of multiported and balanced
slide valves, with poppet valves, or with a positively operated
Corliss type of valve.

60. The Simple D-slide Valve Engine. The simplest
and cheapest type of reciprocating steam engine manu-
factured is shown in part section in Fig. 40 with the prin-
cipal parts labelled. The cylinder, piston, steam chest
and valve are sectioned in order to show the internal con-
struction.

This engine, like most steam engines, is double acting,
that is, a cycle is produced on each side of the piston during
every revolution. Steam is admitted and expanded on one
side of the piston while steam is being exhausted on the other
side. The control of admission and exhaust is effected by
the slide valve and will be considered in detail in later sec-
tions.

The mechanical energy made available by the steam
operating in an engine cylinder is not developed at a uniform
or constant rate, but fluctuates, during each revolution,
above and below the amount required to overcome the
constant resistance at the shaft due to the work the engine

THE REAL STEAM ENGINE 97

is doing. If no provision were made to prevent it, this
would result in a very variable rate of rotation during
each revolution. When the energy made available was in
excess of the demand it would be used in accelerating
the moving parts of the engine and the speed of the latter
would increase. The reverse would occur when the supply
did not equal the demand.

The fly-wheel is used to prevent violent fluctuations

cabined Flywheel
and Beltwheel

FIG. 40. — Simple D-slide Valve Engine.

of this kind. It is made with a comparatively heavy rim
and a great deal of energy must be supplied to accelerate
it to any appreciable extent in a short time. Similarly
it can give out a great deal of energy when slowing down.
The fly-wheel therefore serves as a sort of reservoir in which
excess energy can be stored temporarily and from which
it can later be withdrawn when a deficiency exists. The
fly-wheel thus acts as a damper to variation of rotative

98 STEAM POWER

speed during each revolution, minimizing but not entirely
eliminating such variation. It may also serve as a belt
wheel, as shown in the illustration.

The governor controls the steam supply to the cylinder
in such a way that enough heat will be supplied to make
available the power demanded at the shaft. Were more
supplied the excess would be absorbed by the moving
parts and the engine speed would increase, were less supplied
the engine speed would decrease.

61. Engine Nomenclature. The meanings of several
terms used in describing engines are not self-evident, their

Belt Backward

^ Belt Forward

FIG. 41. — Engine Nomenclature.

definitions depending merely on accepted usage. Some
of these terms and their meanings are illustrated in Fig. 41.

The crank end of a horizontal engine is called the front
of the engine, so that the cylinder head nearest the crank
is called the front head and the stroke of the piston toward
the crank is known as the forward stroke. The forward
stroke of the piston is also spoken of as the outstroke, par-
ticularly in connection with single acting engines. The
stroke away from the crank is correspondingly designated
as the return or the instroke.

62. Principal Parts of Engines. The parts of engines
may be roughly divided into stationary and moving, such as
frame, cylinder, cylinder and valve chest covers, etc., which

THE REAL STEAM ENGINE

are stationary, and piston, piston rod, crosshead, fly-wheel,
etc., which are all moving parts when the engine is in opera-
tion. The moving parts are often divided into reciprocal

FIG. 42. — Frame for Small Vertical Engine.

ing and rotating parts. Thus the piston and all connected
parts through and including the crosshead, and the valve
and many connected parts in the case of slide valve engines,

Jaws for

main bearings

FIG. 43. — Frame for Medium Speed Center Crank Engine.

all reciprocate when the engine is in operation. The
shaft, fly-wheel, eccentric sheaves and governor constitute
the principal rotating parts.

100

STEAM POWER

Some engines also have oscillating parts, such as the
valves in Corliss types, which rock back and forth in the
arc of a circle, and the rocker arms in various forms of
valve gear, these arms rocking through a short arc about
a fixed pin near one end.

The principal parts and their functions are briefly con-
sidered in the following paragraphs:

(a) The Frame. The frame of the engine, sometimes
known as the bed, serves to support the other parts, to tie

Cylinder bolted
to finished
urface here.

Bored cross
head guides

FIG. 44. — Frame for Slow S^eed Engine of Corliss Type; Side or
Overhung Crank.

them together in their proper relations and to fasten the
whole structure to whatever foundation is used. The cross-
head guides and the seats for the main bearings are incor-
porated in the frame.

The frame is commonly made of cast iron in the form
of a hollow box which is properly ribbed to give the neces-
sary stiffness.

Examples of frames are shown in Figs. 42, 43 and 44.

THE REAL STEAM ENGINE

101

(6) The Cylinder and Steam Chest. The cylinder and
steam chest are generally incorporated in the same casting,
and surfaces of covered cavities in this casting are finished

Steam Pipe

Steam Chest'
Cover

m Chest

FlG. 45.

.Steam Chest Cover

to form the cylinder in which the piston operates and the
seat or seats upon which valves rest and move.

The cylinder may be single .walled with flanges on the
end to receive the cylinder head, as illustrated in Fig. 40

102

STEAM POWER

(plain D-slide valve), in which case a thin sheet-metal
jacket is fastened around it and the space between filled
with heat-insulating material. Or, the cylinder may be
cast with double walls, the. space between the two being
used as an air jacket or as a steam jacket.

Some cylinders are fitted with a liner, which is a plain
cylinder pressed into place within the cylinder casting
and forming the bore of the working cylinder. This prac-

FIG. 47. — Section of Atlas Medium Speed Engine, Showing Balanced

Slide Valve.

tice is common on the larger types, the liner being used so
that when wear has occurred it can be replaced cheaply,
instead of it being necessary to rebore or even replace
the main casting.

Examples of cylinder construction are shown in Figs.
45, 46, 47, 48, 49 and 50.

(c) The Piston. The function of the piston is two-
fold. It must prevent the leakage of steam by it from one
end of the cylinder to the other, and it must receive the

THE REAL STEAM ENGINE

103

104

STEAM POWER

pressures exerted by the steam and transmit them to the
other parts of the mechanism as it moves.

Chambers for

Stea.u Valves

(admission valves)

Chambers for
Exhaust Valves

FIG. 49. — Corliss Cylinder Casting.
- Steam Con nectlon

FIG. 50. — Corliss Cylinder with Lagging in Place.

Leakage of steam is prevented by the use of piston
rings, which are metal rings fitted into grooves in the circular
surface of the piston and pressed out against the cylinder

THE REAL STEAM ENGINE

Pistons

105

FIG. 51.

(a) (6)

FIG, 52.

'Bull Rings

FIG. 53. — Built-up Piston Used in Large Engines.

106

STEAM POWER

walls by spring action. They may be made of one piece
of metal turned into a ring of slightly larger diameter than
the cylinder, cut through and sprung into place, or they may
be made in pieces as shown in Fig. 51, and pressed out
against the wall by small helical or leaf springs.

. The piston itself may consist of a solid disk of metal
fitted with a hub and a short cylindrical part with grooves
for the rings, as shown in Fig. 52 (a) and (6) , or it may be
an elaborate built-up structure, as shown in Fig. 53.

FIG. 54.— Crosshead and Pin.

(d) The Piston Rod. The piston rod is a plain circular
steel rod fitted with such shoulders and threads at the ends
as are necessary for the fastening of the piston and the cross-
head. Examples of such fastenings are given in Figs.
51, 52, 53 and 55.

In large horizontal engines the piston rod sometimes
extends through the piston and rear cylinder head, and
the rear end is then supported by a small auxiliary cross-
head. The extension of the rod is known as the tail rod
and the auxiliary crosshead as the tail rod crosshead.

THE REAL STEAM ENGINE

107

Such constructions are used when the weight of the piston
is so great that it would cause serious cylinder wear if not
supported more perfectly than is possible with the ordinary
overhung arrangement.

FIG, 55. — Single Slipper Crosshead.

FIG. 56. — Crosshead with Adjustable Slippers.

(e) The Crosshead and Guides. The crosshead and
guides are used for the purpose of supporting the piston
and its rod and guiding them in a straight line. The
crosshead also serves to connect the piston rod and the
connecting rod through which the forces are transmitted
to the crank pin.

108

STEAM POWER

Crossheads are generally cast in the form of imperfectly
shaped boxes and carry slippers which are faced with anti-
friction metal where they come in contact with the guides.
The slippers may be flat and operate on planed guides,
as shown in Fig. 40, or they may be turned and operate in
bored guides, as shown in Figs. 48, 54, 55 and 56. Pro-
vision is generally, though not always, made for taking
up wear of guides and slippers by setting the slippers
further out from the body of the casting.

With the type shown in Figs. 40 and 54 this can be

FIG. 57. — Solid End Connecting Rod; for Overhung Cranks only.

done by the insertion of thin sheets of rnetal or paper
(known as shims) between the body of the crosshead and the
slippers. In the type shown in Fig. 56 the slippers are
finished with inclined surfaces where they come in contact
with the main casting, and the adjustment is made by wedg-
ing the slippers apart by the use of the adjusting bolts
shown.

The wrist-pin end of the connecting rod enters the
crosshead casting and is held in place by means of the
wrist pin, about which it oscillates when the engine is in
operation .

THE REAL STEAM ENGINE

109

(/) The Connecting Rod. This rod connects the re-
ciprocating crosshead with the rotating crank shaft and
transmits the forces from one to the other. It consists
of a body or shank and two ends or heads. The ends may

FIG. 58. — Connecting Rod with Bolted Strap Ends; May be Used
with Center or Side Crank Constructions.

be " closed " or " solid " as shown in Fig. 57; they may
be made with a strap bolted in place as shown in Fig. 58;
or the crank-pin end may be made of two half boxes bolted
together to form a " marine end " as shown in Fig. 59.
The ends are always made adjustable so that wear

, Crank End

Crosshead End.

FIG. 59. — Marine End Connecting Rod.

can be taken up, thus preventing noisy operation due to
hammering between the ends and the pins at times when
the direction in which forces act is reversed. With solid
and strap types this adjustment is generally made by
means of wedges similar to those shown in Figs. 57 and 58.

110

STEAM POWER

With the marine type shims are used between the two
halves, and the diameter of the hole formed by the latter
is decreased by the removal of shims of the required thick-
ness.

(g) The Shaft. The crank shaft itself is generally made

kPin

k.Arms

^-Journal*

FIG 60. — Crank Shaft, Center Crank.

Counterweights
or Balances

FIG. 61.— Center Crank.

Counterweight

FIG. 62.— Center Crank.

of steel, but the counterbalances are. often of cast iron. It
may be one forging throughout or may be built up by
shrinking the various parts together. Multicrank shafts
of large size are generally of built-up construction, the crank
pins being shrunk into the crank arms and the latter shrunk
on to the pieces of shaft.

The counterbalance weights are used to balance the

THE REAL STEAM ENGINE 111

centrifugal effect of the crank pin, part of the crank arms
and part of the connecting rod, all of which rotate off center.
In some engines part of the unbalanced effect of the recip-
rocating parts is also imperfectly balanced by these counter-
balances.

Various types of shafts are shown in Figs. 60, 61, 62,
63, and 64.

(h) Bearings. Bearings are distinguished as main and
as outboard bearings. Main bearings are those carried by

FIG. 63. — Crank Shaft and Disc, Overhung Crank.

FIG. 64. — Overhung Crank.

the frame of the engine and outboard bearings are carried
by separate pedestals or by pedestals fastened to a plate
which is in turn fastened to the frame. Center-crank
engines have two main bearings, and side-crank engines
only have one, the other end of the shaft being supported
by an outboard bearing.

The bearings of steam engines are generally formed
of babbitt-lined boxes carried within jaws machined in a
frame, or in a separate pedestal, and held in place by a
bearing cap. The boxes are made in two, three or four
parts to allow for adjustment to compensate for wear and

112

STEAM POWER

to give a certain degree of flexibility. Adjustment for
wear is either made by means of wedges or by means of
screws which force the various parts of the boxes toward
the shaft. An example of a three-part bearing with screw
adjustment as used with large side-crank engines is shown
in Fig. 65. The parts of a three-part bearing with wedge
adjustment are shown in Fig. 66.

Bearings are often lubricated by rings or chains, and
they are then known as ring- or as chain-oiling bearings.

FIG. 65. — Three Part Main Bearing with Screw Adjustment.

In the ring-oiling bearing one or more metal rings of large
diameter hang loosely on the shaft within the bearing and
dip into an oil reservoir below the shaft. Rotation of the
shaft causes the rings hanging on it to rotate and they
carry oil up from the reservoir and spill it out over the shaft
within the bearing. Chain bearings are similar except that
chains are substituted for rings.

(i) Fly-wheels. The function of the fly-wheel has al-
ready been considered and need not be discussed further.
The wheel is constructed with a heavy rim joined to a hub

THE REAL STEAM ENGINE

113

by six or eight arms. In the smaller sizes the wheel may be
cast in one piece, but best practice calls for a split hub in
that case to partly equalize certain casting strains which
result from unequal thicknesses of metal in different parts
of the wheel. Large wheels are -cast in two or more parts
both for the purpose of partly avoiding casting strains
and for the purpose of facilitating handling and shipping.

FIG. 66. — Three Part Bearing
Showing Wedge Adjustment.

FIG. 67.

A two-part wheel with the rim sections joined by
prisoner links shrunk in place and the hub fastened with
bolts is shown in Fig. 67.

PROBLEMS

1. A given engine has a piston displacement of 3 cu.ft. and a
clearance volume of 3%. Compression begins when 85% of the
exhaust stroke has been completed and the pressure within the
cylinder at that time is 16 Ibs. per square inch absolute. Deter-
mine the weight of the cushion steam on the assumption that this
steam is dry and saturated at the beginning of compression.

114 STEAM POWER

2. Assume the engine described in Prob. 1 to cut-off at | stroke
and with a pressure inside the cylinder equal to 115 Ibs. per square
inch absolute. Find the weight of cylinder feed if the quality
of the material in the cylinder at the time of cut-off is 75%.

3. Find the piston speed of an engine with a stroke of 2 ft.
and a rotative speed of 150 R.P.M.

4. Show by means of Heck's formula that initial condensation
increases with pressure range.

CHAPTER VIII
THE INDICATOR DIAGRAM AND DERIVED VALUES

63. The Indicator. The ideal steam engine cycle was
described in Chapter IV, and the sort of diagram which
would be obtained from a real engine was shown in Chapter

Drum

Point Holder

Cylinde

FIG. 68.

VII; but the means by which such diagrams are obtained
from operating engines was not given.

Indicator diagrams showing the pressure and volume
changes experienced by steam in the cylinders of real

115

116

STEAM POWER

engines are obtained by means of an instrument known as
an indicator. The operation of obtaining such diagrams is
known as indicating the engine.

An external view of one form of indicator is shown in
Fig. 68 and a section through the instrument is given in
Fig. 69. The method of connecting an indicator to the

-Drum
Spring

Piston Rod

Connected to Engine
Cylinder

FlG. 69.

cylinder of a steam engine and one method used for driving
it are illustrated in Fig. 70.

The indicator is intended to draw a diagram showing
corresponding pressures and volumes within the engine
cylinder and must, therefore, contain one part which will
move in proportion to pressure variations and another which
will move in proportion to volume changes. The one may

INDICATOR DIAGRAM AND DERIVED VALUES 117

be called the pressure-measuring and the other the volume-
measuring device.

The pressure-measuring device generally consists of
a piston, such as shown in the figure, working with minimum
friction in a small cylinder and fitted with a spring which
will resist what may be called outward motion (upward
in the figure). The cylinder containing this piston is
coupled to a short pipe connected with the clearance space
of the engine and, whenever the indicator cock in this
connection is open, the steam acting on the engine piston

FIG. 70. — Method of Attaching and Operating an Indicator.

will also act on the indicator piston. Steam of any given
pressure will drive the indicator piston out against the
action of the spring until the pressure exerted by the spring
is equal to that exerted on the face of the piston. The
indicator piston will thus move out different distances for
different pressures, and, through the piston rod and pencil
mechanism, will move the pencil point to various heights
corresponding to different steam pressures. The pencil
mechanism is so arranged that the point traces a straight
vertical line on the drum as the indicator piston moves in
and out.

Springs .are made to certain definite scales, thus there

118 STEAM POWER

are, for instance, 10-lb., 25-lb., 50-lb. and 100-lb. springs.
The number which is known as the scale of the spring
designates the steam pressure in pounds per square inch
which is required to move the pencil point 1 inch against
the action of such a spring. With a 100-lb. spring in the
indicator, a steam pressure of 50 pounds per square inch
acting on the indicator piston would drive the pencil up a
distance of half an inch, a pressure of 100 pounds per square
inch would give 1 inch of motion and so forth.

The volume-measuring device is of an inferential kind.
It simply indicates the position attained by the engine
piston at the time when a given steam pressure existed in
the cylinder and the volume occupied by the steam can be
calculated from piston position and cylinder dimensions.
The position of the piston is indicated by connecting the
cord wound around the drum to some part of the engine
which is rigidly connected to the piston. The crosshead
is commonly used for this purpose and, since the motion
of this member is generally much greater than the circum-
ference of the drum, it is necessary to use a reducing
mechanfsm of some sort. This mechanism must be very
accurate, so that it moves the drum as nearly as possible
in proportion to the motion of the engine piston.

The pencil point moves up and down as the pressure
within the cylinder varies, and the drum rotates under
the point in proportion to the motion of the engine pis-
ton, so that the combination of the two motions brings the
pencil point to successive positions on the drum which indi-
cate successive corresponding values of steam pressure and
piston position. By mounting a piece of paper, known
as a card, on the drum and pressing the pencil point upon
this paper, the successive positions occupied by the pencil
point will be recorded in the form of a series of curves and
straight lines.

If the drum is rotated with the lower side of the indica-
tor piston connected to atmosphere, the pencil will trace

INDICATOR DIAGRAM AND DERIVED VALUES 119

a horizontal line. This is known as the atmospheric line
and is used as a reference for locating the pressure scale.
If the indicator cylinder is then connected with the engine
cylinder and the drum is rotated by the reducing mechanism,
a diagram similar to that of Fig. 71 will be drawn upon the
card. The atmospheric line indicates the height assumed
by the pencil when atmospheric pressure acts on the
piston and, knowing the value of the existing atmospheric
pressure (barometer reading) and the scale of the spring,
a line at a height representing zero pressure can be drawn
on the card. This line is indicated in Fig. 72.

The length of the card between the lines a and b is
proportional to the length of the engine stroke and there-

£.

Volume Scale

FIG. 72.

fore to the piston displacement, that is, to the volum,e
swept through by the piston. Knowing the clearance
volume of the engine as a percentage or fraction of the
piston displacement, this fraction of the length of the
diagram can be laid off from the end of the diagram to
give a line of zero volume. This line is also indicated in
Fig. 72.

With the line of zero pressure and the line of zero
volume drawn in, all values of steam pressure and volume
occupied by steam can be read directly from the diagram,
and it thus forms a picture of what occurs within the real
engine cylinder.

The indicator diagram is used for a number of purposes,
the more important being:

120

STEAM POWER

Length of Diagram — >j

(1) The determination of the energy made available
within the cylinder, that is, the indicated horse-power, I.h.p.

(2) The determination of the amount of initial conden-
sation and of heat interchanges between walls and cylinder.

(3) The determination of what is known as the diagram
water rate.

(4) The study of the operation and timing of valves.
The second one of these uses has already been considered
in Chapter VII, the others are treated in succeeding sections.

64. Determination of I.h.p. The lines of the indicator
diagram show by their height the pressures or forces acting
on the engine piston as it moves. But the product of force
by distance is equal to work and these lines can be used there-
fore for determining the net
work done by the steam
upon the piston.

In Fig. 73 is shown the
upper part of the diagram,
the curved lines represent-
ing the successive pressures
in pounds per square inch
which acted on the left face
of the piston while it moved
outward. If the average
pressure could be deter-
mined and multiplied by
the area of the piston face, this product would be the
average total force acting on the piston. Multiplying this
by the distance traveled would give the work done by
the steam upon the piston. Expressed in the form of an
equation,

in which

#o = work done upon piston by steam during outstroke;
po = mean pressure (in pounds square inch) acting on
piston during outstroke;

FIG. 73. — Positive Work Area.

INDICATOR DIAGRAM AND DERIVED VALUES 121

a = area of piston face in square inches; and
L = stroke of piston in feet.

For the instroke shown in Fig. 74, the work done by
the piston on the steam is given by the similar expression

(34)

FIG. 74. — Negative Work Area.

in which E< and p« represent work done and mean pressure
respectively.

The net work done by the steam upon the piston per
cycle is then,

— Et = (p0 — pt)aL ft.-lbs. .

(35)

The values of po and pt can be found directly from the
diagram by dividing the areas AQ and At respectively by the
length I and then multiplying by the scale of the spring,
giving

po = -r-X scale of spring,
and

Pi = -r~x scale of spring,

122 STEAM POWEE

so that,

po — pt = p = — —j — - X scale of spring . . . (36)

area of diagram vy . . /0_,

— — X scale of spring. . (37)

The value of p evidently can be determined very simply
from the indicator diagram, and the work per cycle can be
found when p is known by substituting in the following
equation, obtained by putting p for po — pt in Eq. (35),

#cyoie = pXaXLft.-lbs ..... (38)

The pressure p is known as the mean effective pressure and
is often represented by M.E.P.

If n cycles are produced per minute, the net work done
by the steam upon the piston per minute will be

, ..... (39)
which is generally rearranged to read,

Emm^pLan, ....... (40)

in which form the group of letters forming the right-hand
member is easily remembered.

Since 33,000 foot-pounds per minute are equivalent
to one horse-power, it follows that the power made avail-
able as shown by the indicator diagram, that i«, the indicated
horse-power, must be,

in which

p = mean effective pressure in pounds per square inch ;

L = stroke of piston in feet;

a = area of piston in square inches;

= (diam. cyl. in inches)2 XTT/ 4 = .7854d2; and
n = number of cycles per minute.

INDICATOR DIAGRAM AND DERIVED VALUES 123

If an engine cylinder takes steam on one side of the
piston only, that is, if the cylinder is single acting, the num-
ber of cycles produced per minute is equal to the number
of revolutions per minute, but it should be noted that for
other arrangements this is not necessarily true. In the
case of double-acting engines which receive steam at both
ends of the cylinder, the number of cycles produced is equal
to twice the number of revolutions.

It should also be noted that the symbol a represents
the area of the piston face upon which the steam acts.
If a piston rod extend from the face of the piston to and
through the cylinder head (as is always the case at the
crank end of double-acting cylinders), the area of the
piston rod must be subtracted from that of the piston to
obtain the area on which the steam really acts. When a
tail rod is used, a correction must be made for each side
of the piston.

In the case of double-acting engines the indicated
horse-power may be determined in two ways: It may be
figured separately for the two ends of the cylinder, or
the values for the area and pressure may be averaged for
the two ends and the value of n chosen equal to twice
the revolutions per minute. The former is generally the
more accurate method.

It will have been observed that the area of the indi-
cator diagram must be determined before the mean
effective pressure can be found. This area is generally
measured by means of an instrument known as a planimeter,
and this is the most accurate method. It occasionally
happens, however, that a planimeter is not available when
the value of the indicated horse-power is desired. Under
such circumstances an approximate determination of the
area of the indicator diagram can be made by the method
of ordi nates.

For this purpose the length of the diagram is divided
into an equal number of parts, usually ten, as shown in

124

STEAM POWER

Fig. 75 and vertical lines 3/1, y2, ys, etc., are drawn at the center
of each of the parts into which the diagram has been divided.
The mean ordinate or height is then found from the equa-
tion,

=
m

number of vertical lines'

(42)

and the mean effective pressure is then determined by
multiplying ym by the scale of the spring.

An indicator diagram similar to that shown in Fig. 76
is occasionally obtained. The small loop on the end repre-
sents negative work, since the pressure of the steam which

FIG. 75.

FIG. 76.

does work upon the piston is lower than that which resists
the return of the piston. When using a planimeter, this area
is automatically subtracted from that of the rest of the
diagram, but care should be taken to see that this is also
done when the method of ordinates is used.

ILLUSTRATIVE PROBLEM

1. Determine the I.h.p. of a double-acting steam engine, having
a cylinder 8 ins. diameter, length of stroke, 12 ins., running at
100 R.P.M., the mean effective pressure (M.E.P.) on the piston
being 45 Ibs. Neglect the area of the piston rod.

I.h.p. =

pLan (pXa) Ibs.X(Lre) ft. per min.
33,000" 33,000 ft.-lbs. per min7~
(45X8X8X.7854) Ibs. XjfX 100X2) ft. per min.
33,000

INDICATOR DIAGRAM AND DERIVED VALUES 125

2260 lbs.X200 ft. per min.
33,000

= 14 nearly.

2. The I.h.p. of a double-acting engine is 14, the R.P.M. =100;
M.E.P. =45 Ibs.; length of stroke = 12 ins. Find the diameter
of the cylinder, neglecting area of piston rod.

First determine the area of the piston from the formula

plan 33,000 I.h.p.

pXLXn
33,000X24

. (approx.).

65. Conventional Diagram and Card Factors. It is

often necessary to approximate the mean effective pressure
obtained in the cylinder of an engine when no indicator
diagrams are available. The most common case is when
an engine is being designed to carry a certain load and it
is desired to determine the necessary cylinder dimensions
and speed. If the probable mean effective pressure can be
determined, the dimensions and speed can be found from the
equation,

,. , , vLan
I.h.p. per cylinder end = £

by rewriting it

pLnX0.7854d2
33,000 ~
from which

33,000 Lh.fr
" 0.7854 pLn '

Since n is equal to revolutions per minute for one cylinder
end, the product of L by n must be equal to half the piston

126 STEAM POWER

speed of the engine, and a proper value of this product can
be chosen for substitution in the equation. If a proper
value for p can then be predicted the only unknown remain-
ing will be the diameter d, and this can be found by solving
the equation.

The prediction of the mean effective pressure is made
either by drawing upon recorded experience in the form
of values obtained in similar engines previously constructed
or by means of what is known as a conventional indicator
diagram.

The conventional diagram is drawn with upper and
lower pressures equal to those expected in the case of the
real engine, and all expansions and compressions are drawn
as rectangular hyperbolas. The equation of the rectangu-
lar hyperbola is

PlVi-PnVn, (44)

in which subscript 1 indicates initial conditions and sub-
script n represents any later conditions with the same mate-
rial in the cylinder. This law is assumed because it is the
simplest and, as a rough average, gives values as close to
those actually attained as do any of the more complicated
laws.

The diagram may be drawn as nearly as possible like
the one which the engine may be expected to give or it
may be drawn with various simplifications which remove
it more and more from the approximation to an actual
indicator diagram. In any case, the mean effective pres-
sure is determined from this diagram and this value is then
multiplied by a corrective factor, the value of which has
been determined by experience. This corrective factor is
called the diagram factor or card factor and it is realty
the ratio of the area of the diagram the engine would
really give to the area of the conventional diagram
used.

The simplest form of conventional diagram is drawn

INDICATOR DIAGRAM AND DERIVED VALUES 127

by neglecting the clearance volume and has the shape shown

in Fig. 77. The upper line is drawn horizontal at a height

representing the highest pressure expected and of such a

length (compared with the length of the diagram) as will

approximately represent the fraction of the stroke at which

cut-off is to occur in the real

engine. The expansion curve is

then drawn in as a rectangular

hyperbola and extended until the

end of the diagram is reached.

The next line is drawn vertical

and the lower line of the diagram

is drawn horizontal at a height

representing the pressure expected

in the space into which the engine

is to exhaust.

This simple diagram can be
divided into the three areas shown
and the value of the work represented by these areas can be
determined from the equations given below, the first and
last of which should be self-evident from what has preceded.
The equation for the work represented by area A 2 can be
determined very easily by means of integral calculus. The
equations are,

AI represents P\V\ ft.-lbs.;

FIG. 77. — Conventional In-
dicator Diagram.

represents PiFi loge-^ = PiFi loger ft.-lbs.,

and

represents

ft.-lbs.,

in which P represents pressure in pounds per square foot and
V represents volume in cubic feet.

The total area is then equal to the sum ^1+^2 — ^3
and the net work is equal to a similar sum of the right-

128 STEAM POWEE

hand members given above. The net work must also
equal the mean-effective pressure Pm multiplied by the
total volume change, so that

\oger-P2V2, . . (45)
and

-P2 ... (46)

P2, . . . (47)
/

1 Vi

and substituting — for -==- this gives

-P2. (48)

V2
The ratio y" = r is called the ratio of expansion and its

FI 1
reciprocal. -=^- = — is known as the cut-off ratio. By the use

V 2 T

of this ratio the volume terms can be disposed of and the
equation above is obtained. This equation then gives the
mean effective pressure in terms of upper and lower pres-
sures and the fraction of the stroke at which cut-off is
desired in the real engine and no cylinder dimensions need
be known.

Since pressures in steam-engine practice are usually
given in pounds per square inch, the equation for mean
effective pressure is more useful in the form

(49)

in which p\ and p2 and pm are expressed in pounds per
square inch absolute. For convenience in the use of this

INDICATOE DIAGRAM AND DERIVED VALUES 129

equation the values assumed by the bracketed quantity
are given for various conditions in Table III.

TABLE III

1 +\oge r

1 +loge r

1.0

1.00

6.0

0.465

16.0

0.236

1.5

0.937

7.0

0.421

17.0

0.226

2.0

0.847

8.0

0.385

18.0

0.216

2.5

0.766

9.0

0.355

19.0

0.208

3.0

0.700

10.0

0.330

20.0

0.200

3.5

0.644

11.0

0.309

21.0

0.192

4.0

0.597

12.0

0.290

22.0

0.186

4.5

0.556

13.0

0.274

23.0

0.180

5.0

0.522

14.0

0.260

24.0

0.174

5.5

0.492

15.0

0.247

25.0

0.169

The values of the mean effective pressures obtained
from this form of diagram are very much higher than are
to be expected from real engines with the same initial and
terminal pressures and the same nominal ratio of expan-
sion. They are therefore corrected by multiplying by the
proper diagram factor as selected from Table IV. It is
obvious from the range of values given that the selection
of a proper value for the factor depends largely on expe-
rience, but such experience is quickly gained by contact
with real engines and a study of the practical diagrams.

TABLE IV
DIAGRAM FACTORS

Simple slide-valve engine 55 to 90%

Simple Corliss engine 85 to 90

Compound slide-valve engine 55 to 80

Compound Corliss engine 75 to 85

Triple-expansion engine 55 to 70

66. Ratio of Expansion. — The ratio of expansion used
above is sometimes called the apparent ratio. It is not the

130

STEAM POWER

real ratio of expansion for an engine with clearance.
such an engine the real ratio of expansion is

For

FIG. 78.

in which the symbols represent
the volumes indicated in Fig.
78.

The numerical values of r
and r' are often very different
and care should be used in dis-
tinguishing between them. The
diagram factors referred to in
Table IV are for idealized con-

ventional cards without clearance as shown in Fig. 77.

ILLUSTRATIVE PROBLEMS

1. Given an engine with a stroke of 24 ins. and cut-off occurring
at | stroke. Steam pressure of 160 Ibs. per square inch and
back pressure of 16 Ibs. Assume diagram factor =80%. Neglect-
ing clearance, find the probable M.E.P.

160X.7-16 = 112.0-16=96lbs.

Hence probable M.E.P. = .80X96 =76.8 Ibs.

2. A given double-acting engine indicates 75 I.h.p. under the
following conditions:

Cut-off at 20%; steam pressure, 140 Ibs. per square inch
absolute; piston speed, 600 ft. per minute; back pressure, 2 Ibs.
per square inch absolute.

Assume a diagram factor for this type of engine equal to 85% ;
and neglecting clearance, find a convenient size of the cylinder
(diameter and stroke).

INDICATOR DIAGRAM AND DERIVED VALUES 131

Solution.

_ /1+loge

_^ _

\ r I \ 5

= 73.1 -2 =71.1 Ibs. per sq.in.

Diagram factor =85%. Hence probable
M.E.P.=71.lX.85=60.41bs.

Therefore, since

75X33,000

d=9| ins. (approx.);

and since 2Ln = 600, assume L = 1 ft.
hence n=300R.P.M.

The engine is rated 9.5X12 ins., running at 300 R.P.M.

67. Determination of Clearance Volume from Diagram.

It was shown in a preceding paragraph that the clearance
volume of a cylinder must be known in order to draw the
line of zero volumes on the indicator diagram. This
volume can be determined accurately for any real engine
by weighing the quantity of water required to fill the clear-
ance space, but this procedure is often impossible and
an alternative, though approximate, method is often
resorted to.

This method is graphical and depends upon the assump-
tion of the law of expansion and compression. As in the
case of the conventional diagrams, expansion and compres-
sion are assumed to follow rectangular hyperbolas.

It is a property of this curve that diagonals such as
aa and bb drawn for rectangles with their corners on the

132

STEAM POWER

curve all pass through the origin of coordinates as shown
in Fig. 79.

If two points a and c are selected on the expansion
curve of a real diagram and a rectangle is drawn upon
them as shown in Fig. 80, the diagonal bd extended will
pass through the origin of coordinates, if the expansion
follows the assumed law. The point at which this diagonal
cuts the zero pressure line must therefore be the point
through which the vertical line of zero volume is to be drawn.

If the original assumption were correct, this construc-
tion would give the same point when different locations
of the points a and c were chosen and when used on the

a 0

FIG. 79. — -Rectangular Hyperbola.

FIG. 80.

compression as well as on the expansion line. In reality
it will generally give as many different locations for the
origin as are chosen for the rectangle abed. It is customary
to construct this rectangle of fair size and to locate it near
the center of the expansion curve.

68. Diagram Water Rate. As was shown in an earlier
chapter, part of the steam supplied an engine is generally
condensed upon the cold metal walls surrounding it The
indicator diagram therefore shows the volumes assumed by
the mixture of steam and liquid water in the cylinder,
but, since the volume occupied by the liquid is negligible,
it may be assumed to show the volumes occupied by the
part of the mixture which exists in vaporous form.

INDICATOR DIAGRAM AND DERIVED VALUES 133

Assuming that the vapor is saturated, the volume
occupied by one pound at various pressures can be found
from the steam tables and, therefore, the weight existing
in the cylinder can be calculated. The weight of steam
determined in this way is known as the indicated steam,
the diagram steam or the diagram water rate.

The diagram water rate is generally determined for a
point such as z in Fig. 81 just
after cut-off, though some
engineers prefer to use a point
nearer the lower end of the
expansion curve. The volume
occupied by the steam con-
tained in the cylinder at point
z is equal to Vz and its weight
can be determined by dividing
this volume by the specific volume V2 for the existing
pressure P2. Thus, calling the weight of eteam in the cyl-
inder Wz, /

«*-£. V ,,,v .. '.. . . (51)

FIG. 81.

This quantity of steam is a mixture of cylinder feed
and clearance or cushion steam and the weight of the latter
must therefore be subtracted from wz to obtain the weight
of cylinder feed wf. Assuming the cushion steam dry and
saturated at the point k, the weight of cushion steam is

(52)

so that the weight of cylinder feed per cycle as shown by
the diagram at the point z is

V T7

_ _ _.JLf__!_* /KO\

The formula is generally modified to give the steam
consumption per indicated horse-power hour, instead of

134 STEAM POWER

per cycle, and it is also expressed in different terms as a
matter of convenience.
For this purpose let

ija = clearance volume divided by piston displacement per
stroke

?/2 = piston displacement to point z divided by piston dis-

placement per stroke
= h
I,'
yk = piston displacement to point k divided by piston dis-

placement per stroke
_l*
I,'

a = area of piston in square inches ;
p = mean effective pressure in pounds per square inch;
L = stroke in feet ; and
n = number of cycles per minute.

The piston displacement is then TTT^ cubic feet and
the volumes at z and k are given by

and

Substituting these values in Eq. (53) gives the cylinder
feed per cycle as

aL(y+ya y*+yc\ ,

Wf m\ v, v, )-

Multiplying by the number of cycles per hour (60 Xn)
and dividing by the indicated horse-power, ~- jr— , gives

oo,UUU

INDICATOR DIAGRAM AND DERIVED VALUES 135

the diagram water rate, or steam shown by the diagram per
I.h.p. hour as

= 13,750 (ijz+yci yi+ya\

, . . . (55)

in which form the equation involves only values which can
be determined directly from the diagram without any
knowledge of the engine dimensions.

The value obtained for wd will vary as the location of
points z and k are varied because of the quality changes
occurring during expansion and compression, and it is
obvious that the diagram water rate is in no sense an accu-
rate measure of the real water rate of the engine. It is,
however, very useful for comparison with the real water
rate, the ratio giving an indication of the loss by con-
densation.

Average values for real water rates are given in Chapter
XL

ILLUSTRATIVE PROBLEM

Given the diagram shown in Fig. 82 and the following data
from an actual test, find the diagram water rate for point c, and
for point n. Double-acting steam
engine having:

Average piston area = 28.9 sq.-
in.;

Length of stroke =8 in. ;

R.P.M. =237; I.h.p. =8.75;
M.E.P. =31.6 Ibs.;

Clearance =13%; Beginning of
compression = 29% ;

Weight of condensate per hour
=371 Ibs.;

Quality at throttle =95%;

Sp. vol. at c=7.8;

Sp. vol. at n =12.57;

Sp. vol. at K =38.4 cu.ft. per lb.;

Assume % = 100%.

FIG. 82.

136 STEAM POWER

Solution. Substitution in Eq. (55) gives

, , _13,750/?/c+?/cz 2/*+2/cz

_13,750/0.39+0.13_0.29+0.13
= 31.6 \ 7.8 38.4

=24.2 Ibs. per I.h.p. per hour at point c.

Q1 _ T/

01. D \ Kn 1/fc

13,750/0.638+0.13 0.29+0.13

31.6 \ 12.57 38.4

= 21.83. Ibs. per I.h.p. per hour at pcint n.

371

Real water rate = X0.95 =40.2 Ibs.

&75

69. TV-diagram for a Real Engine. In Chapter VI
the T^-diagrams of the various ideal cycles were given and
attention was called to the fact that these diagrams were
. particularly useful, because they showed certain things
which were not apparent from the more common PV-
diagrams.

It has been customary for many years to draw TV-
diagrams for real engines by " transferring " the PV-
diagram to TV-coordinates, and various analytical and
graphical methods have been developed for this purpose.
There are certain unavoidable errors in all the methods
used for drawing these diagrams, and the expansion curve
is the only one of all the lines finally obtained which has
any claim to accuracy. Even this curve is generally incor-
rectly interpreted, because a knowledge of the exact weight
of clearance steam is necessary for an accurate interpreta-
tion and such knowledge is never available.

Under the circumstances it seems unnecessary to con-
sider in this book the rather complicated details involved
in the construction of T ^-diagrams purporting to show
the behavior of steam in real engines.

INDICATOR DIAGRAM AND DERIVED VALUES 137

70. Mechanical and Thermal Efficiencies. The method
of obtaining the indicated horse-power from the indicator
diagram has been given in preceding paragraphs. In the
real engine this power is not all made available at the shaft,
because some of It is used in driving the engine against
its own frictional resistance. Calling the power lost in
this way the friction horse-power, it follows that

I.h.p. = F.h.p.+D.h.p, .... (56)

in which

I. h. p. = indicated horse-power determined from the real

indicator diagram;
F.h. p. = friction horse-power, i.e., power required to

drive engine; and

D.h.p. = developed horse-power, i.e., power made avail-
able at shaft.

The developed horse-power is therefore always less
than the indicated horse-power. The better the construc-
tion of the engine the smaller the friction loss, and the
measure of this loss is usually given in the form of an ef-
ficiency. It is called the mechanical efficiency, and is defined
by the equation

(57)

Values of mechanical efficiency range from about 80 per
cent in the case of poorly designed and poorly adjusted
horizontal engines to about 95 per cent in the case of the
best vertical designs.

The efficiency determined by dividing energy made
available by heat supplied is known as the thermal efficiency.
There are two possible thermal efficiencies, one based on
the indicated power and the other on the developed power.
The former is called the thermal efficiency on the indicated
horse-power or the indicated thermal efficiency; the other

138 STEAM POWER

is known as the thermal efficiency on the developed horse-
power or the developed thermal efficiency. Obviously

Dev. ther. eff. = Mech. eff.X Indie, ther. eff. . . (58)

The heat supplied may be assumed in two different
ways; it may be taken as the total heat above 32° F. in the
steam supplied the engine, or it may be taken as this value
less the heat of the liquid corresponding to exhaust tem-
perature. The second method is preferable, since it is
reasonable to assume that the exhaust steam can be con-
densed to water at the same temperature and that this water
can be pumped to the boiler with the heat of the liquid
corresponding to this temperature. This is practically
parallel to the assumption made in treating the theoretical
cycles.

The thermal efficiencies are then

Indie, ther. eff.

I.h.p.X2545

Heat above exhaust temp, supplied per hour'
and

Dev. ther. eff.

= D.h.p.X2545

Heat above exhaust temp, supplied per hour*

(59)

(60)

Values of the indicated thermal efficiency range from
about 5 per cent in ordinary practice with small engines to
about 25 per cent in the best large engines. Values as low
as 1 per cent are not uncommon with small, poorly designed
and poorly operated engines.

The actual performance of the cylinder of an engine
is sometimes compared with the ideal possibilities as indi-
cated by the Clausius and the Rankine cycles. The ratio
of the work obtained in the real engine to that which could
be obtained from the same quantity of heat with a Rankine

INDICATOR DIAGRAM AND DERIVED VALUES 139

or Clausius cycle is a measure of the performance of the
real cylinder. This ratio is variously designated as cylinder
efficiency, indicated efficiency, relative efficiency, etc. Its
values range from less than 40 per cent to over 80 per cent,
the highest recorded value being just over 88 per cent.

PROBLEMS

1. Using Table I, Chapter I, plot the specific heat of water
between the range of temperatures of 20° F. and 300° F. for the
intermediate values given. By the ordinate method for finding
the mean height of an indicator diagram, determine the mean
or average specific heat over this range.

2. A double-acting engine is required to give 50 I.h.p. under
the following conditions:

Cut-off =25%;

Steam pressure = 150 Ibs. per square inch absolute;
Back pressure = 16 Ibs. per square inch absolute;
Piston speed =540 ft. per minute.

If the diagram factor for this type of engine is 75%, find the
diameter of the cylinder and select the stroke and R.P.M.

3. Assume a single-acting engine with 10-in. diameter and 12-in.
stroke, 10X12 ins., to have cut-off occur at various points between
10% and 50% of stroke. Assume also the pressures, speed, and
card factor as given in Prob. 2. Find the probable I.h.p. at
different cut-offs.

4. Given an 18X24-in. engine running at 120 R.P.M.
Back pressure =2 Ibs. per square inch absolute;
Clearance = 10%;

Cut-off =40%;

Diagram factor =85%.

Supposing cut-off to remain constant, find the I.h.p.'s cor-
responding to steam pressure of 50, 90, and 130 Ibs. per square
inch absolute.

5. Find the weight of dry steam which must be supplied per
I.h.p. hour for each case of the previous problem, assuming the
quality at cut-off to be 80%. Assume compression pressure to
be 30 Ibs. absolute and that steam is dry and saturated at end of
compression.

6. Find the quality of steam at cut-off in a cylinder, in which
the piston displacement is 0.1278 cu.ft.; clearance = 10%; cut-off
at 25% stroke; steam pressure at cut-off, 115 Ibs. per square

140 STEAM POWER

inch absolute, and weight of steam in the cylinder at cut-off =
0.012 Ib.

Actual vol.

JNote. Quality =— — — for the given pressure.

W eight XSp. vol.

7. The piston displacement of a certain engine is 0.2 cu.ft.
What weight of steam is in the cylinder at release where quality
is 90%, and pressure is 25 Ibs. per square inch absolute, if the
clearance is 10%, and release occurs at 95% of the stroke?

8. Find the weight of cushion steam in a 6X6 in. engine in
which clearance = 15%; compression begins at 85% of the return
stroke; back pressure is 14.7 Ibs. per square inch absolute, and
the quality of the cushion steam at the beginning of compression
is 95%.

9. Find the pressure and quality at the end of the compression
line of the previous problem, assuming it to be adiabatic.

10. An 8X 10 in. engine running at 300 R.P.M. is double-acting,
and cuts off at 15% of the stroke at a pressure of 120 Ibs. per
square inch absolute. It has a steam consumption of 35 Ibs. per
I. h. p. -hour. The compression begins at 60% of the return stroke
with a quality of unity and a back pressure of 5 Ibs. per square
inch absolute. Clearance = 10%.

If this engine delivers 27 H.P., and has a mechanical efficiency
of 90%, what is the quality at the point of cut-off?

11. In the previous problem, assume release to occur at 90%
of the stroke with an absolute pressure of 30 Ibs. per square inch.
What is the quality at this point?

12. A certain engine gives one horse-power hour at the shaft
for every 20 Ibs. of steam supplied. The steam has an initial
pressure of 150 Ibs. absolute and is dry and saturated when it
arrives at the engine. The back pressure against which steam is
exhausted is 4 Ibs. absolute.

(a) Find the thermal efficiency of this engine on the developed
or shaft horse-power.

(6) If the mechanical efficiency of the engine is 90%, what
is the value of the thermal efficiency on the indicated horse-power?

CHAPTER IX

COMPOUNDING

71. Gain by Expansion. The cycle which gives a
rectangular PF-diagram is the least economical of all the
ideal cycles described in Chapter IV. This comes from
the fact that none of the heat stored in the steam is con-
verted into work when this cycle is used. Thus, if the
cylinder shown by full lines in
Fig. 83 operate on this cycle
and be of such size that it will
receive just one pound of steam
per cycle, it makes available an
amount of work represented by
the area abed. The positive
work done by the steam upon
the piston is the equivalent of
the external latent heat of
vaporization while no use is
made of the heat stored in the
steam. This stored heat is re-
moved as heat during the con-
densation and exhaust, which give the lines be and cd.

If a piece be added to the cylinder as indicated by
the dotted lines, the same quantity of steam will make more
heat available by expanding after cut-off, as shown by the
curve be, the net work in this case being represented by
the area abefd instead of by the smaller area abed. But
the heat supplied is the same in both cases, namely that
required to form one pound of steam at the pressure PI,
so that the use of a large cylinder and the incomplete ex-

141

FIG. 83.

142

STEAM POWER

pansion cycle results in the development of more work
than can be obtained with the rectangular cycle from the
same amount of heat.

Obviously it would be theoretically advantageous to
add still more to the length of the cylinder and allow the
expansion to continue to back pressure, giving the com-
plete expansion cycle as shown in Fig. 84, thus obtaining
the maximum quantity of work at the expense of the heat
stored in the steam supplied the cylinder. Practically,
it is found inadvisable to continue the expansion to such a de-

FIG. 84.

gree in reciprocating steam engines, because at low pressures
the volume increases very rapidly for small pressure drops.
Thus a great increase is necessary in the size of the cylinder
if the last part of the expansion is to be completed, but
the amount of work obtained is comparatively small, as
shown by the small height of the long toe thus added to
the diagram. This may result in an actual loss, because the
increased friction losses of the very large cylinder may
more than balance the small increase of net work gained
by its use. It thus results that, in every real reciprocating
engine, there is some point beyond which it is not economi-
cal to carry the expansion, and the incomplete expansion

COMPOUNDING

143

FIG. 85.

cycle is therefore approximated in such engines rather
than the cycle with complete expansion.

Viewing the matter from another angle, a cylinder of
a certain size may be assumed as shown in Fig. 8-",. The
use of the rectangular cycle
abed in this cylinder will make
available the maximum quan-
tity of work possible with the
upper and lower pressures
chosen. If cut-off be made
to occur earlier as at &', the
expansion b'c' will result in a
loss of the quantity of work
obtained, as shown by the
area b'bc', but the steam used
per horse-power will be less,

so that there will be a gain in steam economy. Putting the
cut-off still earlier will cause a still greater loss of work
obtained from a cylinder of the chosen size, but theoreti-
cally will result in greater economy of steam.

Summing up, it may be said that the greater the ratio
of expansion the greater should be the economy in the use
of steam on a theoretical basis.

The lower pressure is set in real engines by the pressure
in the space into which the engine is to exhaust. If the
engine is to be operated non-condensing, the atmospheric
pressure determines the lowest possible exhaust pressure;
if the engine is to be operated condensing, the exhaust
pressure is set by the lowest pressure which can be eco-
nomically maintained in the condenser.

There is thus a real limit to the extent to which expan-
sion can be carried in any real engine with a given initial
pressure. A certain drop must exist at the end of the
diagram, for reasons already explained, and an expan-
sion line drawn backward from the top of the line repre-
senting this drop will give the earliest possible cut-off

144 STEAM POWEE

which can be used in the engine with a given initial
pressure.

The ratio of expansion can be further increased, how-
ever, by raising the initial pressure as indicated by the

dotted lines in Fig. 86,
pi n anci the limit in this direc-

tion would come with the
inability of- materials of
construction to withstand
the resulting strains.

These conclusions
drawn from the facts
V developed above must all

-P, be modified in the case of

*IG. 8b. .

real engines, because of

the effect of cylinder condensation. This has been shown
to increase as the cut-off is made earlier and as the
pressure (and therefore the temperature) range in a cyl-
inder is increased. There is, therefore, a limit beyond
which it is not advisable to carry the ratio of expansion in
a single cylinder.

Experience has shown that the best commercial results
are obtained from simple engines, that is, those expanding
the steam entirely in one cylinder, when (a) they are operated
non-condensing, (b) the initial pressure is between 80 and 100
pounds per square inch for the simpler forms of valves and up
to 125 Ibs. with the better forms of valves, and (c) the point
of cut-off is at about J stroke with the simpler valves and at
from I to \ stroke with the better forms of valves. These
values of cut-off correspond to nominal expansion ratios
of 5 and 4 respectively and to lower values when clearance
is taken into account.

72. Compounding. If the ratio of expansion is to be
increased above the values just given, some means must
be used for the reduction of loss by condensation. This
loss can be reduced by decreasing the surface exposed to

COMPOUNDING

145

high-temperature steam and by decreasing the temperature
range in a cylinder. Both of these results can be achieved
by what is known as compounding.

Assume that it is deemed advisable to produce a cycle
similar to that shown in Fig. 87 (clearance neglected) and
that in order to obtain
high steam economy (low
water rate) the ratio of
expansion chosen is very
much greater than four.
No gain in economy
would result from such
excessive expansion in a
single cylinder, in fact
there would be a well-
defined, unavoidable loss.
But suppose that the high-
pressure steam is admitted
to a small cylinder such
as that shown and is ex-
panded to the point /, is

High Pre&sure

Cylinder
H.P. Cylinder

Low Pressure Cylinder
L.P. Cylinder

FIG. 87.

then exhausted as shown
by fg into the larger cyl-
inder along gf and then expanded to the point c in the
larger cylinder. The cycle produced is the same as that
which would have been obtained by expanding entirely in
one cylinder, but the surface of the clearance space of
the high-pressure (H.P.) cylinder, which is exposed to high-
pressure steam is smaller than it would be in a cylinder of
the size required to hold the steam when fully expanded and,
moreover, the lowest temperature to which it is subjected is
that corresponding to the pressure at / instead of the
much lower temperature corresponding to the pressure
at d.

The condensation which would occur in the H.P. cylinder
would obviously be less than that which would result from

146 STEAM POWER

the use of one large cylinder and, remembering that the
greater part of the heat given up during condensation is
received again by the steam during exhaust, it is obvious
that approximately this same quantity of heat can again
be given to the low-pressure cylinder walls. Thus, although
there are two cylinders in which condensation and re-
evaporation occur, and although the sum of the heat given
to the walls of the high-pressure cylinder and that given
to the walls of the low-pressure cylinder might be greater
than that given to the walls of a single cylinder under similar
conditions, the use of two cylinders results in a consider-
able saving because loss in the high-pressure cylinder is
practically wiped out by the exhaust of the heat concerned
into the low-pressure cylinder.

If the loss by radiation and conduction from the high-
pressure cylinder be neglected, the result of the use of two
cylinders is practically to limit the loss by condensation
and re-evaporation to that occurring in the low-pressure
cylinder. As the ratio of expansion in this cylinder is in
the neighborhood of that common in simple engines, or

even less, and as the tem-
perature range is small,
the net loss is also small.
It is obvious that the
smaller the surface of the
high-pressure cylinder can
be made, and the smaller
the temperature range in
a single cylinder, the
smaller will be the net loss
FIG. 88. by cylinder condensation

and re-evaporation. A

saving should therefore be effected by using more than two
cylinders, and it is not inconceivable that five or more might
be used. The result of using five cylinders is shown in -Fig.
88, and it is evident that the clearance surfaces exposed to

COMPOUNDING 147

high temperatures, the temperature ranges per cylinder
and the ratios of expansion per cylinder are all small.
The gain in economy should therefore be correspondingly
great.

There are two limits to the possible multiplication of cylin-
ders in this way.

(1) As the number increases the radiating surface and
therefore the heat lost by radiation increases. The extent
of this effect can be appreciated by noting that every
cylinder with the exception of the low-pressure cylinder is
really an unnecessary addition, because the cycle could be
produced entirely in the low-pressure cylinder. On the
other hand, the surfaces of cylinders which operate at high
temperature are small as compared with that which would
be exposed to this temperature if the entire cycle were pro-
duced in the low-pressure cylinder.

(2) As the number of cylinders is increased, the first
cost, the complexity and the cost of lubrication and attend-
ance are all increased so that, for each installation, some
number will be found beyond which the interest on the
investment and the added cost of operation and mainte-
nance would more than balance the saving of fuel.

The second limit mentioned is the more important
commercially, as it is the first one reached. For ordinary
operating conditions in stationary power plants expansion
in two cylinders generally gives the most economical results.
The total ratio of expansion is generally between 7 and 16,
that is, the volume of steam at release in the L.P. cylinder
is from 7 to 16 times the volume at cut-off in the H.P.
cylinder. For large pumping stations and large marine
installations, expansion in three cylinders is generally
considered the most economical, and total ratios of expansion
of 20 or more are used. Four and five cylinders have
been used, but the resultant gains do not seem to warrant
any extensive installation of such units.

Engines using more than one cylinder for the expansion

148 STEAM POWEK

of steam in the way just described are called multi-expan-
sion engines, or compound engines, and the use of multi-
expansion is spoken of as compounding. Custom has
almost confined the use of the term compound engine
to those in which only two cylinders are used in series as
indicated in Fig. 89, and such engines are often spoken of
as 2x engines.

Engines in which three cylinders are used in series
are called triple-expansion or 3x engines. With four and
five cylinders in series the engines are known as quadruple
or 4z and quintuple or 5x, respectively.

In the case of triple-expansion engines of large size,

H.P. Exhaust and L.P. Admission

To
Condenser •

FIG. 89. FIG. 90.

the volume of the low-pressure cylinder required generally
becomes so great that it is found economical to use two
low-pressure cylinders instead of one. The flow of steam
in such an engine is represented diagrammatically in Fig.
90. This type is known as a four-cylinder, triple-expansion
engine.

All multi-expansion engines are generally operated
condensing, and the choice of type is determined partly by
the character of work to be done and partly by economical
considerations. In all cases the boiler pressure must be
chosen to suit the type of engine used. The pressures
ordinarily used with the different types are given in
Table V.

COMPOUNDING 149

TABLE V
BOILER PRESSURE COMMONLY USED

Type of Engine.

Boiler Pressure.
Pounds per Sq.in. Gauge.

Simple

80 to 125

High-speed compound.

100 to 170

Low-speed compound
Triple expansion and higher

125 to 200
125 to 225

73. The Compound Engine. The term compound
engine will be used hereafter in the commercial way as
referring to a 2x engine. Such engines may roughly be
divided into two types, receiver and non-receiver engines.
The latter are often called Woolf engines, after the man
who first used this construction.

A receiver engine has a vessel known as a receiver
located between the two cylinders and so connected with
them that the high-pressure cylinder exhausts into the
receiver and the low-pressure cylinder draws its steam
from the receiver. By using a receiver the cylinders are
made independent of each other so far as steam events
are concerned; the high-pressure cylinder can exhaust
at any time with reference to the events occurring in the
low-pressure cylinder.

A Woolf type has practically no receiver, the high-pres-
sure cylinder exhausting directly into the low-pressure
cylinder through the shortest convenient connecting pass-
age. As the high-pressure cylinder must exhaust directly
into the low-pressure cylinder it follows that cut-off must
not occur in the latter until compression starts in the
former; i.e., very near the end of the stroke.

An engine with a receiver of infinite size would give
a horizontal exhaust line for the high-pressure cylinder
and a horizontal admission line for the low-pressure cylinder,
since the small amount of steam given to or taken from the

150

STEAM POWER

receiver would have no appreciable effect upon the pressure
within that vessel. Neglecting throttling losses, the high-
pressure and low-pressure cards would therefore fit together
as originally indicated in Fig. 86.

With receivers of finite size there are pressure changes
during exhaust by the high- and admission to the low-pres-
sure cylinders, and real valves and connections also cause
certain throttling losses, so that the lines representing
these events are not horizontal nor do they exactly coincide.

A diagrammatic arrangement of the Woolf engine
is given in Fig. 91 with idealized diagrams obtained by

FIG. 91.

assuming hyperbolic expansions, no clearances, and no
throttling losses. The pistons must make their strokes
together in such engines, but they may move in the same
direction, as shown in the figure, or in opposite directions.

The ideal diagram would be that shown at (a) by the
lines AbcdCDA. The idealized high-pressure diagram is
abcda and the idealized low-pressure diagram is ABCDA.
The exhaust line da of the high-pressure diagram and the
admission line BC of the low-pressure diagram are pro-
duced at the same time. Corresponding points on these
two lines represent the common pressures assumed by the
steam not yet exhausted from the high-pressure cylinder,
the steam in the small connecting passage and the steam

COMPOUNDING

151

already admitted to the low-pressure cylinder. As the
movement of the low-pressure piston opens up volume
faster than the high-pressure piston closes up volume,
the volume occupied by the steam continues to increase as
the low-pressure piston moves out, and its pressure there-
fore decreases.

The two diagrams are shown back to back at (6) in the
figure and the horizontal line xX connects corresponding

FIG. 92.

points on the exhaust of the high pressure and the admission
of the low pressure.

Compound engines are also divided into two types
on the basis of cylinder arrangement. When the axes of
both cylinders coincide as shown in Fig. 92 they are called
tandem compounds. When the axes are parallel as shown
in Fig. 89, the engines are spoken of as cross-compound
engines.

74. Cylinder Ratios. The idealized diagrams of a com-
pound engine with infinite receiver
volume are shown in Fig. 93 by
abed and ABODE. The height of
the high-pressure exhaust line is the
same as that of the low-pressure
admission line and represents the
receiver pressure PR. The value of
the receiver pressure is determined
by the point chosen for cut-off in the low-pressure cylinder.
Thus if cut-off in the low-pressure cylinder is made to occur
earlier, as at some point c', the admission line for this

FIG. 93.

152

STEAM POWER

cylinder must move up to B'C' and the receiver pressure
must rise correspondingly. The exhaust pressure in the
high-pressure cylinder would also rise an equal amount.

Changing the point of cut-off in the low-pressure cylinder
also produces another result. As the receiver pressure rises
the work area of the high-pressure diagram is obviously de-
creased, while that of the low-pressure diagram is increased.
In a simple engine the area of the diagram becomes smaller
the earlier the cut-off, and it should be noted that just the
reverse of this occurs in the low-pressure cylinder of a com-
pound engine.

It is evident that the choice of the receiver pressure or
of the point of cut-off in the low-pressure cylinder determines
the relative areas of the high-pressure and low-pressure
diagrams and it also determines the relative size of the two
cylinders. The diagram of Fig. 93 shows that late cut-off
in the low-pressure cylinder calls for a larger high-pressure
cylinder than does early cut-off.

The ratio of the piston displacement of the low-pressure
cylinder to that of the high-pressure cylinder is called the
cylinder ratio. Designating this ratio by R, and using other
symbols as in Fig. 93,

(61)

The cylinder ratios chosen for real compound engines
vary greatly in different designs and no given ratio has been
proved the best for a given set of conditions. Normal
practice gives the average values listed in Table VI, but
cylinder ratios as high as 7 have been used with excellent
results.

TABLE VI

CYLINDER RATIOS FOR COMPOUND ENGINES

Cylinder ratio

Initial pressure (gauge) non-condensing

100

120

Initial pressure (gauge) condensing

100

120

150

COMPOUNDING 153

The cylinder ratio to be used in a given case may be
determined by any one of several considerations or by a
combination of them, the latter being more often the case.
Thus it may be deemed desirable to obtain the same amount
of work from both cylinders; or to obtain equal temperature
ranges; or to have cut-offs occur at the same fraction of
the strokes; or to have the same total load on the two
piston rods during admission; or to obtain the maximum
possible uniformity of turning effort at the crank. The con-
sideration of equal work is generally ,
regarded as the most important. p^

75. Indicator Diagrams and Mean ^
Pressures. The idealized diagrams p
for a compound engine with clearance,
with incomplete expansion in both p
cylinders, arid without compression
are given in Fig. 94. The nominal

total ratio of expansion would be LL+1H, but the total ratio
of expansion taking account of clearance is

T I r<]
Total ratio of expansion == L . ~7L, . . (62)

and the cylinder ratio is

R = ~- (63)

The mean effective pressures can be found from each
of the diagrams in the ordinary way and the indicated
horse-power of each cylinder determined therefrom. The
indicated horse-power of the engine is then equal to the sum of
the values obtained for the separate cylinders.

It is often convenient to refer the mean effective pres-
sure of all cylinders to the low-pressure cylinder as though
this were the only cylinder acting. In the simple form
of diagram, such as that shown in Fig. 93, it is obvious
that this could be obtained by measuring the area AbcDEA,

154 STEAM POWEE

dividing by the length AE and multiplying by the scale of
the spring, just as though the diagram were all produced
in one cylinder with the piston displacement equal to VL.
In the case of the diagrams given in Fig. 94 a similar method
could be adopted, or the mean effective pressure of each
cylinder could be determined separately and then the
equivalent pressure which would give the same result on
the low-pressure piston could be determined analytically.

Assume for this purpose that the mean effective pressure
of the high-pressure is equal to pH pounds per square inch,
that the mean effective pressure of the low-pressure cylinder
is equal to pL and that the cylinder ratio is R. The strokes
of all cylinders of a multi-expansion engine are generally
equal, so that the piston areas are in the same ratio as the
cylinder volumes (piston displacements). In the case of
a 2x engine, therefore, the area of the low-pressure piston
is R times as great as that of the high-pressure piston,
and the pressure required on the low-pressure piston to do
the same work as that done by pressure pn on the high-

pressure piston will be ^.
R

In the case of a 2x engine therefore the total M.E.P.
referred to the low-pressure cylinder is

(64)

This mean effective pressure acting on the low-pressure
piston only would give the same indicated horse-power as
is obtained with the two cylinders of the engine.

In designing compound engines it is customary to
determine the size of the low-pressure cylinder as though it
were to do all the work expected of the engine by receiving
steam at the highest pressure available and exhausting it
at the lowest. The mean effective pressure which would
thus be assumed to exist is the referred value PR just ex-
plained. Having found the size of the low-pressure cylinder.

COMPOUNDING 155

and the value of the referred M.E.P. the size of the high-
pressure cylinder can be determined so that the work done
by each cylinder will be just half of the total for which the
engine is being designed. This size will have to be such
that the high-pressure mean effective pressure referred to
the low-pressure cylinder (i.e., pH + R) is equal to half the
total mean effective pressure referred to that cylinder.
That is, the size will have to be so chosen that

ILLUSTRATIVE PROBLEM

A double-acting compound engine is capable of developing
500 I.h.p. The stroke is 18 ins.; revolutions per minute, 175;
mean effective pressure referred to L.P. piston, 45 Ibs. per square
inch; cylinder ratio, 3^. Find cylinder diameters.

From

"33,000'

500X33,000

«L.P. =

45X1.5X175X2

so that

dL.p. =\r^7 =30 ins. (approx,),

\ ./oO

with the cylinder ratio equal to 3j,
700

/200
dH.p. = *fogg = 16 ins. (approx.).

76. Combined Indicator Diagrams. When a compound
engine is indicated, the diagrams of the two cylinders as
drawn by the indicator are not directly comparable. The
scales of pressure and volume are different on the two dia-
grams, and correction must be made for this fact before the

156

STEAM POWER

diagrams can be compared. It is customary to do this and
to draw the average high-pressure and low-pressure diagrams
on the same set of coordinates in order to determine how
well they approximate the ideal diagram that would be
obtafned in one cylinder operating between the extreme
limits of pressure.

Diagrams approximating those that would be obtained
from high- and low-pressure cylinders are shown at (h)
and (7) respectively, in Fig. 95, and the result of drawing

b s I .... I ... .1 i .... i .... i

FIG. 95.

both to the same scales is shown at the left of this figure.
The curves Xh and Xi show the variations of quality along the
two expansion curves.

Drawing the two diagrams to the same scales in this
way is known as combining the diagrams and the result is
known as a combined diagram.

The curves SS and S'S' added to the combined diagram
are saturation curves. They do not, in general, form a
continuous curve, because of the different quantities of
steam contained in the two clearances and because any

COMPOUNDING

157

moisture in the high-pressure exhaust is generally removed
in the receiver. The volumes occupied by clearance steam
at initial pressures are indicated by the points 6' and B'
respectively. The lengths b'S and B'S' approximately
represent the volumes that would be occupied by cylinder
feed when in each cylinder if dry and saturated.

A combined diagram for a triple-expansion engine is
shown in Fig. 96. The heavy lines give diagrams con-
structed so as to represent
as nearly as possible what
may be expected to occur in
the cylinders of such an en-
gine, assuming perfect valve
action and hyperbolic expan-
sions and compressions. The
dotted diagrams indicate the
shapes that would be drawn
by indicators applied to the
real cylinders. The numer-
ous sharp angles are due to
overlapping of events, one
cylinder suddenly starting to draw from a receiver while
another is exhausting. It will be observed that the dotted
diagrams do not contain any of these sharp angles, but
that their general outline forms a fair average of them.

The curve cd is a rectangular hyperbola drawn as a
continuation of the assumed hyperbolic expansion line of
the high-pressure cylinder. The failure of the expansion
lines of the other cylinders to fall upon this curve is ex-
plained by quality changes, different quantities of clearance
steam in the different cylinders and withdrawal of moist-
ure from steam exhausted to receiver before admission to
the following cylinder,

-- --d

FIG. 96.

158 STEAM POWEE

PROBLEMS

1. Find the size of the cylinders of a double-acting compound
engine, which is to give 600 I.h.p., when using steam at a pressure
of 150 Ibs. per square inch absolute, and having a back pressure
of 2 Ibs. per square inch absolute. The cylinder ratio is to be 4,
and the total ratio of expansion 12, piston speed 750 ft. per minute,
and R.P.M. =150; diagram factor is 80%.

2. Given a 200 H.P. compound Corliss engine with cut-off
in the H.P. cylinder at 60% stroke. Ratio of expansion is 7;
clearance is 7%; card factor is 70%; pressure at the H.P. cyl-
inder is 165 Ibs. absolute. Find

(a) Cylinder ratio; V» 3

(6) Theoretical and actual M.E.P.;

3. Given a compound engine 18X40 ins., having a stroke of
28 ins. Steam pressure is 165 Ibs. per square inch absolute;
cut-off in H.P. cylinder occurs at 62% stroke; clearance equals
16%; back pressure equals 5 Ibs.; R.P.M. equal 150. Find

(a) Cylinder ratio; 4.4</ -. y

(6) Ratio of expansion; 7.V

(d) Lh.p. mo

CHAPTER X
THE D-SLIDE VALVE

77. Description and Method of Operation. The simple
D-slide valve, shown in place in Fig. 97, is so named because
of the similarity of its section to the letter D. It is located
in the steam chest, rides back and forth upon its seat and

FIG. 97.

serves to connect the two ports alternately with steam
and exhaust spaces respectively in order to give the neces-
sary distribution of steam.

The valve has to perform the following functions for
each end of the cylinder during each revolution of the
engine :

(1) It connects the proper port to the steam space or

159

160 STEAM POWER

steam chest at such a time that steam can enter the cylinder
as the piston moves away from the head.

(2) It shuts off this port and thus cuts off the supply
of steam when the piston has completed a certain definite
fraction of the stroke.

(3) It connects the port with the exhaust cavity shortly
before the piston reaches the end of the stroke, thus effecting
" exhaust " or " release "; and

(4) It shuts off the port again when the piston has com-
pleted the proper fraction of the next stroke, thus trapping
in the cylinder the steam which is compressed during
the remainder of the stroke.

Engine Crank
Main Connecting

It is obvious that the valve must be reciprocated upon
its seat and that its motion must be connected with that of
the piston in some way so that the proper phase relation
may be retained. This could be effected by the system
shown diagrammatically in Fig. 98, a small crank operating
on the end of a connecting rod giving the valve its short
stroke just as the main crank fixes the longer stroke of
the piston. Such an arrangement would, however, be
very inconvenient with many real engines, as the valve would
be located too far from the center line of the cylinder.

It is customary to use what is known as an eccentric
for the purpose of operating the slide valve. The parts
and arrangement of an eccentric, together with an illus-
tration of the wav in which it is mounted on the shaft of

FIG. 99. — Parts of Eccentric.

162

STEAM POWER

THE D-SLIDE VALVE

163

an engine are shown in Figs. 99, 100 and 101. The motion
it gives the valve is exactly the same as that imparted
by the crank first assumed, and it can
easily be shown that it is the exact
equivalent of such a crank.

Assume, for example, a crank such
as that shown in Fig. 98 with a length
of arm or throw equal to a. If the
crank pin is made larger while other
parts of the crank remain the same, as
shown in Fig. 102, the crank mech- FlG- 101. -Eccentric
anism is not essentially altered; the mo-
tion which it would impart to a connecting rod is not
changed. If this process of enlarging the pin be continued

1
FIG. 102. — Equivalence of Crank and Eccentric.

until the pin has become large enough to surround the
shaft and if the crank arm be then removed so that what
was the crank pin is fastened directly on the shaft, an

164

STEAM POWER

FIG. 103.

FIG. 104. — Slide Valve without Lap,

team Space

Exhaust Cavity Connected
to Exhaust Pipe

Valve and piston both start
to move toward right.

Piston at mid-stroke, valve
at end of stroke and about
to return.

illj \ Valve just closing left hand steam
port and about to open right hand
port to steam. Piston about to start
on return stroke.

Piston at mid-stroke, valve
wide open and about to return.

FIG. 105.

THE D-SLIDE VALVE

165

eccentric results. It is the exact equivalent of the original
crank; its center, which is the center of the crank pin,
revolves about the center line of the shaft in a circle with
a radius a just as in the original mechanism.

The eccentric makes it possible to place a short crank
(short arm) upon a large diameter shaft without having to
cut the shaft away as shown in Fig. 103, and it is therefore
very useful for driving valves.

78. Steam Lap. The simplest possible form of D-slide
valve would just reach the outer edges of the ports when
in its central position as shown in Fig. 104. The crank
driving it (that is the crank equivalent to the eccentric
which would probably be used in a real case) would have
to be located 90° ahead of the engine crank in the direction
of rotation, as can easily be seen by consulting Fig. 105,
which illustrates the mechanism in various critical positions.
The illustration shows that such a valve would give full
stroke admission, thus producing a rectangular cycle which
has already been shown to be very inefficient as a means of
obtaining work from the heat .used in
forming steam.

If cut-off is to occur before the
end of the stroke, the edge of the
valve must return and close the port
before the piston reaches the end of its
stroke. But since the crank mechan-
ism does not permit the valve to
remain stationary in any one posi-
tion, such early cut-off could only occur if the valve
over-traveled, as shown in Fig. 106, and this would un-
fortunately result in connecting the working end of the
cylinder to exhaust and in admitting steam to the other
side of the piston at such a time as to oppose the piston's
motion. The solution of the difficulty lies in making
the valve longer, so that when in its central position it
overlaps the outer edges of the ports as shown in Fig. 107.

Steam Space

-« — Valve Travel

Exhaust/
Cavity

Piston Travel

FIG. 106.

166

STEAM POWER

The amount of overlap of the outer edge is called the out-
side lap, and when steam is admitted by the outer edges of
the valve, as in the case under discussion, it is also called
the steam lap.

With such an arrangement the valve must be drawn
out of its central position by the amount of the lap when
the piston is at the end of its stroke as shown by a in Fig.

FIG. 107. — Steam and
Exhaust Lap.

FIG. 108. — Lap a and Lap
Angle a.

108 in order that steam may be admitted just as the
piston starts to move. It follows that the crank driving
the valve must be more than 90° ahead of the engine
crank and that it must be ahead by The angle required
to move the valve a distance equal to the outside lap.
This angle, represented in the figure by a, is called the lap
angle.

79. Lead. In real engines it is further desirable to start
the admission of steam just before the piston arrives at
the end of its stroke. This assists in bringing the moving
parts to rest, raises the pressure in the clearance to full
value before the piston starts, and gives a wider opening
through which the steam can flow during the early part of
the stroke, thus reducing wiredrawing and loss of area at
the top of the diagram. If the valve is to open before the
piston reaches the end of its stroke, the crank driving it
must be shifted still further ahead of the engine crank.
It must be shifted ahead by an angle which will draw
the valve through the distance which will give the desired
opening of valve with the piston at the end of its stroke
as shown by 6 in Fig. 109. The angle required, indicated

. THE D-SLIDE VALVE 167

by ft, is known as the angle of lead, and the width
of the steam opening with engine crank on dead center,
i.e., the distance b, is known as the lead. The lead
varies from less than ^ in. on small engines and with low
speeds up to over f in. on large engines and with very high
speeds.

80. Angle of Advance. The eccentric or valve-operating
crank must be ahead of the engine crank by an angle equal
to 90°+angle of lap a+angle of lead 0, as can be seen

FIG. 109.— Lead 6 and Lead Angle (3.

by an inspection of Fig. 109. The sum of a and ft is called
the angle of advance and will be represented by 5. This
is the number of degrees in excess of 90 by which the eccen-
tric leads the engine crank.

Fig. 109 shows that cut-off in an engine fitted with
a valve having lap and lead must occur when the engine
crank has turned through an angle equal to 180 — 2«, because
the valve will then have returned to the closed position.
Apparently, cut-off can be made to occur at any point
in the stroke by properly choosing the value of a, but it
will be discovered later that the exhaust events set a limit
to increase in the value of this angle and hence do not per-
mit of cut-off occurring earlier than a certain fraction of the
stroke.

81. Exhaust Lap. Inspection of Fig. 105 will show that
the simple valve without lap originally assumed will give
no compression, because the cylinder end is connected to
the exhaust cavity for the entire stroke. Inspection of all

168 STEAM POWER

the changes which have been suggested in the subsequent
paragraphs will show further that if the inner edges of the
valve are left in the original positions the exhaust events
will be considerably distorted in the case of a valve having
steam lap and lead.

This trouble may be remedied by moving the inner
edges of the valve closer together, making the exhaust
cavity in the valve shorter and giving inside lap as shown
in Fig. 107 by 6. When the inner edges of the valve control
exhaust, as in the case of the valve under discussion, this
inside lap is also called exhaust lap.

The length of the valve, the lap and the lead are gen-
erally chosen so as to give the desired arrangement of
admission and cut-off and then the exhaust edges are so
located as to give desirable release and
compression. In some forms this necessi-
tates the use of an exhaust cavity in the
FIG. 110. valve such as that shown in Fig. 110.

The amount by which the edges of the
valve fail to meet the inner edges of the port is spoken of as
negative inside lap. This dimension is indicated by c in the
figure.

It should be noted particularly that all measurements
of lap are made with the valve central on its seat and
that the measurement of lead is made with the piston at
the end of its stroke, i.e., with the engine crank on dead
center.

82. The Bilgram Diagram. The action of all slide
valves could be studied by means of drawings of the actual
mechanism, as has been done in preceding paragraphs, but
such a method is time and space consuming. Numerous
diagrams such as the Elliptical, the Sweet, the Zeuner and
the Bilgram have been developed for the purpose of simpli-
fying and expediting such studies and, when properly
understood, they are very convenient. The scope of this
book does not permit a discussion of all of these diagrams

THE D-SL1DE VALVE

169

and, since the Bilgram diagram is probably the most gener-
ally applicable, attention will be confined to it.

The construction of this diagram is illustrated in Fig.
111. The point 0 represents the center of the engine
crank shaft and the two circles drawn about this point
as a center represent respectively the paths traveled by the

FIG. 111.

pin of the valve crank and the pin of the engine crank.
These circles are drawn to any conven:ent scales.

The diagram is conventionally drawn in such a way
that the line OM represents the head end dead center
position of the crank and in all subsequent paragraphs the
relative positions shown by the small sketch in Fig. Ill
will be assumed. The cylinder will be assumed to the

170 STEAM POWER

left of the shaft and the engine will be assumed to run
" over."

With the crank in position OM, the eccentric (equivalent
crank) must be in the position OB, ahead of the crank by
an angle 90°-|-«+0 = 90+<5. The valve must then be
displaced to the right of its central position by an amount
represented by the distance DB, if a small correction for
" angularity " of the valve connecting rod be neglected.
As rotation continues, horizontal distances corresponding
to this line will always give the instantaneous valve dis-
placements. For position OB' , for instance, the valve dis-
placement will be D'B' '.

If the angle 5 is now laid off above OX, locating the point
Q as shown, a perpendicular QE dropped upon OX from this
point will equal in length the line DB, and will therefore
show the valve displacement when the crank is in head end
dead center position OM. This must be true, because the
triangles QOE and BOD are similar and have the sides
OQ and OB equal to the radius of the same circle.

The perpendicular QE is really a perpendicular dropped
upon the extension of the line representing the crank posi-
tion, and it is a general property of this diagram that a line
starting at Q and perpendicular to the line representing
any chosen crank position (or an extension of that line)
will show by its length the displacement of the valve when
the crank is in the chosen position. Thus assume the engine
crank to rotate through the angle 7 to the position OM' .
The eccentric will have rotated to B' and the valve dis-
placement will be represented by D'B' . A perpendicular
drawn from Q upon OX' , the extension of the crank posi-
tion, gives QE equal to B'D' and hence representing the
valve displacement to the same scale.

This construction drawn for different crank positions
OA, OM, OMi, OM2, etc., is shown in Fig. 112, the dash-
dot radial lines about Q representing the various values of
the valve displacement. The number of each of these

THE D-SLIDE VALVE

171

lines indicates the crank position to which it corresponds.
It will be seen that the displacement increases in value
until the crank position OM^ is reached, after which it
decreases again.

Steam Lap
Circle

Since the opening
ment minus the lap, as
by whicji the valve is
found by subtracting
placement the amount
head end dead-center
to lap plus lead, and is

FIG. 112.

to steam is equal to the displace-
shown in Fig. 109, the actual amount
open for any crank position can be
from the corresponding valve dis-
of lap possessed by the valve. For
position, the displacement is equal
shown by QE in Fig. 112. Subtract-

172 STEAM POWER

ing the lead EF, the remainder FQ gives the lap of the valve.
A circle drawn about Q with radius equal to QF (or a circle
drawn about Q and tangent to the line L) will cut off of
the lines representing valve displacement the amount
representing the part of each displacement used in over-
running the lap of the valve. The remainders, that is the
parts of the lines radiating from Q in Fig. 112 which are
outside of the lap circle, must then represent the amounts
by which the valve port is actually open.

It will be observed that the valve is open by the amount
of the lead when the crank is on dead center, position OM .
The crank position for which the valve displacement is just
equal to the lap, and hence at which the valve is just begin-
ning to open, can be found by drawing a tangent through
0 to the lower side of the lap circle and then extending
it to give the crank position OA in Fig. 112.

As the crank rotates clockwise from this position, the
valve opens wider until, when position OM^ is reached,
the greatest valve opening exists. Further rotation results
in partial closure of the valve and, when the crank has
finally rotated into position OC, the valve has just closed,
that is, cut-off has occurred, the displacement being just
equal to QG, the steam lap.

Thus this diagram, as so far developed, indicates crank
positions for admission and cut-off and the values of valve
displacement and valve openings for all intermediate
crank positions.

ILLUSTRATIVE PROBLEM

A certain valve has an external steam lap equal to 1| ins.
The lead is ^ in. and the throw of the eccentric is 1\ ins. (a) Con-
struct such parts of the Bilgram diagram as are necessary to
indicate "head end" crank positions for admission, maximum
valve opening and cut-off. (6) Indicate on this diagram the
amount of valve opening at various crank positions between
admission and cut-off, (c) Determine the value of the angle of
advance.

THE D-SLIDE VALVE

173

Draw a circle with radius equal to the eccentric throw, 2£
ins., using any convenient scale. This circle is designated by abed
in Fig. 113. Draw about the same center another circle of any
convenient size. Draw in. the horizontal diameter ac and extend
as shown. On the right-hand side of the circle draw the line ef,

Steam Lap
Circle

FIG. 113.

parallel to the horizontal axis and a distance above it equal to the
lead, iV in., to the same scale as that chosen for eccentric circle.
The steam lap circle must have its center Q on the upper right-
hand quadrant of the eccentric circle, and it must be tangent
to the line ef. Its radius must equal the steam lap, 1J in. to scale.
Therefore, with compass points set the proper distance apart, find
the center Q, about which a l£-in. radius circle will just be tangent
to the line rf, and draw the steam lap circle.

174

STEAM POWER

The crank position at admission is found by drawing the line
AO so that, if extended, it is tangent to the lower side of the
steam lap circle.

The crank position at cut-off is found by drawing the line

] .^Inside lap

FIG. 114.

M""0 in such position that it is tangent to the upper part of
the steam lap circle.

The crank position for maximum valve opening is found by
drawing the line M"0 in such position that a line through QO
be perpendicular to it. The amount of valve opening at this

THE D-SLIDE VALVE 175

crank position is shown by the length of the part of this per-
pendicular line outside of the steam lap circle, i.e., the distance
Og interpreted according to the scale chosen for eccentric and
steam lap circles.

When the crank is in position M' 0, the length of hi, interpreted
to scale, gives the amount by which the valve is open to steam.

When the crank is in position M'"0, the length of jk, inter-
preted to scale, gives the amount by which the valve is open to steam.

The angle indicated by 5 is equal to the angle of advance
because of the property upon which the construction of this
diagram is based.

83. Exhaust and Compression. The exhaust edge events
can be shown on the Bilgram diagram by a method similar
to that used for the steam edge events. The direction in
which valve displacements occur are indicated in the upper
part of Fig. 114 ih which the crank and eccentric circles
have been drawn to such scales that they coincide. In-
spection of the small sketch in the lower part of the figure
will show that head end release must occur when the valve
has traveled a distance equal to the inside lap to the left
of its central position. A crank position OR drawn tangent
to the lower part of a circle about Q with radius equal to
the inside lap will, therefore, be the crank position at re-
lease. Clockwise rotation from tbis position will result
in a wider opening to exhaust until position OM\ is reached,
after which the valve will begin to close. Final closure
will occur when the crank reaches position OK, the exten-
sion of which is tangent to the top of the exhaust lap circle.
At that time the valve will have returned (moving from left
to right) and will still have to move a distance equal to
the exhaust lap before attaining a central position.

ILLUSTRATIVE PROBLEM

Given the exhaust lap of a D-slide valve equal to f in.; the
steam lap \\ ins.; the throw of the eccentric, 2 ins.; and the
lead | in. Find the angle of advance, the maximum port opening
to steam and to exhaust, and the crank positions of cut-off, release,
compression and admission for the head-end of the cylinder.

176

STEAM POWER

Draw the eccentric (and crank) circle with a radius equal
to 2 ins., and draw the horizontal diameter as in Fig. 115.

Draw a horizontal line in the upper right-hand quadrant at a
distance of f+lj ins. above the horizontal diameter. Locate
the point Q at intersection.

FIG. 115.

Draw the steam lap circle with a radius lj in. and the exhaust
lap circle with a radius f in.

The angle of advance is the angle between OQ and the hori-
zontal.

The maximum opening to steam is given by the distance
Oa=f in. The maximum opening to exhaust is given by the
distance Ob = If in.

The crank positions shown are obtained by drawing lines

THE D-SLIDE VALVE

177

tangent to the lap circles. A represents admission; C, cut-off;
R, release, and K, beginning of compression.

The piston positions at the times of these events are given
to reduced scale by vertical projection.

84. Diagram for Both Cylinder Ends. The complete
diagram for the head end cylinder is shown in Fig. 114 with
all critical crank positions marked. The positions for the
crank end of the cylinder can be found in a similar way by
constructing a diagram in which the point Q and the lap
circles are located in the opposite quadrant. The resulting

E. Admission

FIG. 116.

diagram for both cylinder ends, with laps the same for both
ends of the valve, is given in Fig. 116.

85. Piston Positions. The valve events might be studied
entirely in conjunction with crank-pin positions, but it is
more convenient and customary to consider them in connec-
tion with piston positions. Piston positions corresponding
to different crank-pin positions could be found by drawing
the mechanism to scale for each different position as shown
in Fig. 117 for piston positions 1 and 2.

It is obvious that this would involve a great deal of work
and that, if drawn to large scale, it would consume a great

178

STEAM POWER

deal of space. Further, it is
convenient to be able to locate
relative piston positions on the
line which serves as the hori-
zontal diameter of the crank
circle of the Bilgram diagram.
The method used depends
upon the fact that the motion
of the crosshead is exactly the
same as that of the p'ston,
so that if the motion of the
crosshead end of the connect-
ing rod can be followed, it
will be equivalent to following
the motion of the piston itself.
It should also be noted that
the diameter of the crank cir-
cle must be equal to the
stroke of the engine.

Assume now, that the point
b in Fig. 117 be taken to rep-
resent the position of the pis-
ton when it is really in posi-
tion 1. When the piston has
moved to position 2, the cross-
head will have moved from a
to a! and the crank pin from
6 to 6'. If with a' as a center
the connecting rod be s wung
down to the horizontal its
right-hand end will arrive at
the point c. The distance be
must then represent the dis-
tance that crosshead (and pis-
ton) have moved from dead-
center position because ab and

THE D-SLIDE VALVE

179

a'c both represent the length of the connecting rod and c
must therefore be as far to the right of 6 as a' is to the right
of a. The point c may therefore be taken to represent
piston position when the connecting rod is in the position
a'b'.

' In general, if the horizontal diameter of the crank
shaft be taken to represent the stroke of the engine, the pis-
ton position corresponding to any crank position can be
found by taking a radius equal to the connecting-rod length
(to the same scale as the circle) and striking an arc from the

FIG. 118.

crank-pin position, using a center on the horizontal line on
the cylinder side of the crank circle.

An approximate method is also used for finding the piston
position. Instead of projecting down from the crank-pin
position with an arc, such as b'c in Fig. 117, a vertical line
through the crank-pin position is used. Such a line would
give c' as the piston position when c is really correct. This
method would give accurate results with a connecting rod
of infinite length. For ordinary lengths of rod, however,
the results are far from correct. The error is said to be due
to the angularity of the connecting rod.

The effect of the angularity of the connecting rod is
shown in Fig. 118 for different positions. On the outstroke
the piston is always farther ahead than the rectilinear pro-

180

STEAM POWER

jection would indicate. On the return stroke the piston is
always behind the position indicated by rectilinear projection.

FIG. 119.

86. Indicator Diagram from Bilgram Diagram. Since
the piston positions corresponding to different crank posi-

THE D-SLIDE VALVE 181

tions can be determined, it is a comparatively simple matter
to construct the indicator diagram which theoretically
would be given by an engine fitted with a valve of certain
dimensions. It is necessary to assume the upper and lower
pressure and also to assume the form of the expansion
and compression curves. These are generally taken as
rectangular hyperbolas.

The method of constructing an indicator diagram from
the Bilgram diagram is shown in Fig. 119. The crank-
pin positions for admission (A), cut-off (C), release (R)
and beginning of compression (K) are first found. These
pin positions are then projected to the horizontal diameter
by means of arcs with radius equal to the connecting-rod
length and with centers on the line MM produced to the
left. The intersections a, c, r and k indicate the piston
positions at which the corresponding events occur. These
are then projected vertically downward to intersect the
proper pressure lines and the card is drawn through the
intersections.

Diagrams constructed in the same way, but for both
head and crank ends, are given in Fig. 120. A symmetrical
valve was assumed, that is, one built exactly alike on head
and crank ends. The diagrams show that such a valve
cannot give the same results for both cylinder ends because
of the effect of the angularity of the connecting rod. It
is most evident in the case of cut-off. The cut-off in this
case occurs just before three-quarter stroke for the head end
and just after half stroke for the crank end of the cylinder.
All other events are distorted in the same way, but the actual
lengths of the variations are not as great as in the case of
the cut-offs and therefore the distortion is not as obvious.

The effect of the angularity of the connecting rod upon
the diagrams can be remembered easily if it is noted that all
valve events occur later with respect to piston position on
the outstroke and earlier on the instroke than they would
with a connecting rod of infinite length.

182

STEAM POWER

It is possible to " equalize " the cut-offs, that is, make
them occur at the same fraction of the stroke by using
unequal steam laps at opposite ends of the valve, but this
will result in still further distortion of admissions, as can be
seen by constructing a Bilgram diagram for this case.
Similarly, the compressions can be equalized by the use of

C.E. Admission

H.E. Admissi

JH.E} Card ir\ '$£•*• Card

FIG, 120.

unequal exhaust laps, but this results in distortion of the
release events.

Various linkages have been developed which are so
arranged that they distort the motion of the valve to just
the extent necessary to counterbalance the effects of the
angularity of the connecting rod. The scope of this book
does not, however, permit a discussion of such valve
gears.

THE D-SLIDE VALVE 183

87. Limitations of the D-slide Valve. The simple
valve discussed in the preceding paragraphs has numerous
limitations and is therefore only used on small and cheap
engines, or in cases where economy in the use of steam is
not essential. This valve, when used with steam entering
over the outside edges as previously considered, is pressed
to its seat by the live steam acting over its entire upper
surface. This pressure is practically unbalanced, as the
greater part of the lower surface of the valve is subjected
to the low pressure of the steam being exhausted. As a
result the friction to be overcome in moving the valve is
very great and there is an appreciable loss from this source.

Further, the shape of the valve makes necessary the use
of long ports which form part of the cylinder clearance
and which are alternately exposed to live and to exhaust
steam with results previously discussed. These ports can
be decreased in length by increasing the length of the valve,
but this in turn increases the area exposed to high pressure
and hence increases the friction loss.

It can be shown by means of the Bilgram diagram
that, if a cut-off earlier than about f stroke is desired,
the angle of advance, the amount of steam lap and the size
of the eccentric must all be made very great. This results
not only in large friction losses, but also in very early release
and compression, because of the great angle of advance.
As a result, slide valves of the simple D type are seldom used
when a cut-off earlier than | to ^ stroke is desired. It
should be remembered in this connection that the simple
engine generally gives its best economy with a cut-off of
about J stroke.

The drawing of lines representing the opening of the
valve to steam as in Fig. 112 will show that this simple
valve is further handicapped by the very slow opening
and closing of the steam ports, causing a great amount of
wire drawing with a corresponding loss of diagram area.
In order to get an adequate opening to steam the valve

184

STEAM POWER

must also be given a great displacement and, since this

occurs under great pressure, it results in great friction loss.

The unbalanced feature can practically be overcome

by rolling up the valve
and ports about an axis
parallel to the length of
the cylinder. This gives
what is known as a pis-
ton valve, shown dia-
grammatically in Fig.
121

Jt can also be

-Steam Ports

FIG. 121.— Piston Valve.

tially overcome by using

a balance plate or ring of some kind between the top of
the valve and the inside of the steam-chest cover, so
arranged that live steam is excluded from the greater part
of the upper surface of the valve. Valves of this type
are generally called balanced slide valves and are used on
many high- and medium-speed engines.

The valve travel required for obtaining a given opening

Steam

FIG. 122.— Allen Double Ported Valve.

can be decreased and the rate of opening and closing can be
increased by the use of multiported constructions. These
are so arranged that two or more ports open or close at the
same time, so that the total movement required for a given
opening is divided by the number of ports and the rate of
opening and closing is multiplied in the same proportion.
One simple type of double-portea valve is illustrated in
Fig. 122.

When several ports are used the valve often becomes

THE D-SLIDE VALVE 185

a rectangular frame crossed by a number of bars and is
known as a gridiron valve, because of its appearance. Such
valves are often combined with balance plates and give
very satisfactory results.

A number of designs of slide valves have been developed
for the purpose of making cut-off independent of the other
events. Many of these use a separate cut-off valve which
either controls the steam supply to the main valve or else
rides on the main valve and controls cut-off by covering
ports in that valve. Devices of the latter type are called
riding cut-off valves. They are either driven by separate
eccentrics, or by linkage from the eccentric controlling
the main valve, the linkage being so arranged as to give
the proper relative motion between main and auxiliary
valves. In such designs the main valve is proportioned
so as to give the desired admission, release and compres-
sion and the cut-off is then taken care of by proper adjust-
ment of the cut-off valve.

88. Reversing Engines. It was shown in one of the early
paragraphs of this chapter that the eccentric must be set
90°+ angle of advance ahead of the crank, ahead meaning
in the direction of rotation. To cause the engine to revolve
in the opposite direction, that is, to " reverse " the engine,
it is therefore only necessary to shift the relative positions
of eccentric and crank so that the eccentric leads the crank
by 90° -\-5 in the new direction of rotation. This corre-
sponds to shifting ahead (in first direction of rotation)
through an angle equal to 180 — 25 or shifting backward
through an angle equal to 180+26, as can be seen by inspec-
tion of Fig. 109.

In practice it is generally more convenient to use two
eccentrics, one set properly for rotation in one direction
and the other set properly for rotation in the opposite direc-
tion. This arrangement is shown diagrammatically in
Fig. 123. This figure is drawn for a vertical engine and in
such position that the engine is on crank-end dead center.

186

STEAM POWEE

FIG. 123.

Rever

ight Shaft

The point P represents the position of the center of the crank

pin; the point / represents the position of the equivalent
crank (center of eccentric) which
drives the valve for " forward,"
" ahead " or clockwise rotation; and
the point b represents the position of
the equivalent crank which drives the
valve for " backing/' " reverse," or
counter-clockwise rotation.

The real mechanism, in one of its
numerous forms known as the Stephen-
son Link Gear, is shown in perspective in Fig. 124. The

forward eccentric corresponds to / of Fig. 123 and the

backing eccentric corresponds

to 6 of that figure. The1

eccentric rods are fastened

to opposite ends of a curved

" link " and move the valve

through a " link block "

fastened to the end of the

valve stem. In the position

shown in the figure the link

is in such position that the

forward eccentric operates

practically directly on the

valve stem so that the valve

motion is practically entirely

governed by that eccentric.

If the reverse shaft were to

be rotated clockwise into the

backing position, the " sus-
pension rods " would pull

Clock
Rotat

FIG. 124.— Stephenson Link Gear.

the link over until the eccen-
tric rod of the backing eccentric was directly under the
valve stem. Under such conditions the valve motion
would be controlled almost entirely by the backing

THE D-SLIDE VALVE 187

eccentric and the engine shaft would rotate counter-clock-
wise.

If the mechanism were so set that the link block occupied
a position on the link between the ends of the two eccentric
rods, the valve motion would be controlled by both eccentrics
and would be a compromise between the motions given by
either eccentric separately. It is characteristic of this gear
that the cut-off is latest when either one or the other eccentric
is fully " in gear " and that it becomes earlier as the link
block approaches the center of the link. With the link
block in the center of the link the valve does not open at
all, i.e., the cut-off occurs at zero stroke.

There are numerous other forms of link gears, the best
known being the Gooch, the Allan and the Porter-Allen.
There are also numerous reversing mechanisms known as
radial gears in which the motion of the valve is controlled
by means of a " radius rod " which can be set to give the
desired valve motion. The valve motion is obtained in-
directly through the radius rod from an eccentric, from the
crank, or from the connecting rod. The limits of this
book do not permit a detailed discussion of these forms.

89. Valve Setting. From what has preceded it will
be evident that it is not only necessary that a valve and
its seat and driving mechanism be correctly designed, but
also that the various parts must be correctly connected up
in order that the valve may move in its proper phase rela-
tion with respect to the piston.

Adjusting the mechanism in such a way that the proper
phase relations are obtained is known as setting the valve.
This can be done with fair accuracy by a simple study of
the mechanism in various positions, as will be shown below,
but it is always advisable to check the setting by means of
indicator diagrams taken after the setting is completed.
Such diagrams will often show errors of such character or
size that they cannot be determined by measurement on
an engine which is not operating.

188

STEAM POWER

Before beginning operations it is always advisable to go
over the entire engine carefully and to eliminate excessive
lost motion at all pins and bearings in order that the relative
positions of parts obtained while setting the valve may
approximate those which will be obtained when the engine
is in operation. The effect of lost motion will be appreciated
after a study of Fig. 125. Assume that all parts of the
mechanism are tight except the crank-pin end of the con-
necting rod as shown. If the engine is rotated by hand,

FIG. 125.

for instance -, by turning the fly-wheel, the crank will pull
the piston mechanism and the piston will be drawn into the
position shown in the upper half of the figure when the crank
has turned through an angle a. On the other hand, when
the engine is operating under steam, the piston will push
the crank pin around and will occupy a position such as
that shown in the lower half of the figure when the crank
has been turned through the same angle a. Obviously,
the piston can occupy two very different positions for the
same crank position, and a valve setting based upon the
conditions shown in the upper part of the figure might be

THE D-SLIDE VALVE 189

very incorrect when used under the conditions shown in
the lower part of the figure.

Lost motion in any part of the mechanism can produce
analogous results and it is therefore necessary to remove as
much of it as possible before attempting to set the valve.
It is practically impossible to eliminate all lost motion, as
there must be sufficient clearance at all bearing surfaces
to accommodate a film of oil, and this alone would make
necessary the taking of indicator diagrams for the check-
ing of valve settings, even if it were possible to set perfectly
by measurement for stationary conditions.

In general, there are two adjustments which can be made
in setting a plain slide valve. The length of the valve
stem or eccentric rod can be changed and the eccentric
can be shifted around the shaft. It is necessary to under-
stand the effects of each of these adjustments.

Changing the length of the valve stem is equivalent to
shifting the valve upon its seat without
moving the engine as shown in Fig.
126. In this figure the valve is shown

in its central position by full lines. The
lap is the same at both ends. If, now, FIG. 126.

the valve is worked to the right upon
its stem by adjustment of the nuts shown, until it reaches
the dotted position, the head-end lap will have been de-
creased and the crank-end lap will. have been increased by
the same amount. This would make admission earlier and
cut-off later for the head end and admission later and cut-
off earlier for the crank end. Obviously, the effects of
changing the length of the valve stem are opposite for the
two ends of the cylinder.

Shifting the eccentric about the shaft simply changes
the time relation between valve motion and piston motion;
it does not alter the valve motion itself. If difficulty is
experienced in realizing the truth of this statement, it is
only necessary to draw several Bilgram diagrams for the

190 STEAM POWER

same valve, but with different angles of advance, and then
to construct indicator diagrams for both cylinder ends in
every case. It will be discovered that shifting the eccentric
ahead in the direction of rotation, for instance, will make
all events occur earlier with respect to piston position for
both ends of the cylinder.

In setting a plain slide valve which is built symmetrical
about a central axis, i.e., same inside and outside lap at
each end, it is first necessary to adjust the length of the valve
stem. This may be done by removing the steam-chest
cover so as to expose the valve and then rotating the engine
slowly by hand and observing the distance traveled by the
valve on each side of its central position. This is con-
veniently done by observing the distance between the outer
edge of the steam port and the outer edge of the valve when
the valve is fully open at each end. If the valve travels
further toward the head end than it does toward the crank
end, with reference to the port edges, the valve stem must
be shortened; if it travels further toward the crank end
the stem must be lengthened.

In making these adjustments it is advisable to turn the
engine only in the direction in which it is going to rotate, so
that any lost motion in the valve mechanism will have
approximately the same effect as when the engine is opera-
ting.

When the length of the valve stem is correctly adjusted,
the eccentric must be so set on the shaft as to give the proper
angle of advance. This is commonly done by shifting it
about the shaft until the proper value for the steam lead has
been obtained. In order to determine the value of the lead
it is necessary to be able to set the engine on each dead center.
This can be done approximately by turning the engine until
the crosshead has come to either end of its stroke, but it
will be found by trial that the fly-wheel and shaft can be
turned through a very large angle at each end of the stroke
without causing an appreciable motion of the crosshead,

THE D-SLIDE VALVE

191

so that this method is not very satisfactory for the purpose
of adjusting the eccentric. It is customary, therefore,
to work in such a way as to give a more accurate determina-
tion of shaft and crank positions for dead center.

The engine is rotated until the crosshead has been
brought near one end of its stroke, as shown in Fig. 127,
and a mark is then scribed across the crosshead and guide
as at ab. An arc xy is then marked on the fly-wheel by
means of a tram such as that shown, the end c being placed

FIG. 127.

at point P on some solid part of foundation or floor. The
engine is then rotated, clockwise in the figure, until the
crosshead has reached the end of its stroke and returned
to such a point that the marks on crosshead and guides
again coincide, as shown by dotted positions in the figure.
The arc x'y' is then scribed on the fly-wheel with the tram,
the end c again bearing on the point P. A point z is then
found by bisecting the arc ef and when this point is brought
under point d of the tram the crank will obviously be at
crank-end dead center and the piston at the crank end

192

STEAM POWER

(a) Perfect Cards for Slide Valve Type.

(6) Actual Card; Small Engine. Center Line of Valve on Center
Line of Seat; Eccentric Advanced to Give Normal Lead of
0.05 inch. Engine Running Over.

THE D-SLIDE VALVE

193

(d) Angular Advance of Eccentric Increased. Valve Stem Length
Same as in (6) and (c), Lead 0.375 Inch,

0) Angular Advance of Eccentric Decreased so as to Give Negative
Lead of 0.5 Inch. Length of Valve Stem Unchanged.

C. E.

H. E.

(/) Length of Valve Stem Changed; Angle of Advance as in (6).

FIG. 128.

!94 STEAM POWER

of its stroke. A point on the fly-wheel diametrically opposite
to z is next found, so that when it is brought under point
d of the tram the engine will be on head-end dead center.

It is probable that more accurate results are obtained
by rotating the engine in a direction opposite to that in
which it rotates under steam, because lost motion is then
taken up in the same direction as when working, but when
the whole process of valve-setting is considered it is ques-
tionable whether this is the correct direction of rotation.
Opinion and practice differ in this respect. In the end,
the setting should be checked by the taking of indicator
diagrams, so that effects of incorrectible lost motion may be
finally eliminated.

With the dead -center points found the engine is placed
on, say, head-end dead center, and the eccentric shifted until
the valve is open to steam by the desired lead. The eccen-
tric is then fastened in this position and the engine turned
to the opposite dead center. Because of angularity of con-
nections and of irregularities in valve and seat dimensions,
it generally will be discovered that the valve is not now
open to steam by the same amount as at the other end.
If it is desired that it should be, the valve can be shifted
on its stem about half of the distance by which it is out
and the eccentric can then be swung about the shaft to take
up the remaining distance. The effect should then be
checked by putting the engine on the opposite dead center.

Valves may be set for equal leads as above, or for equal
cut-offs or for any sort of a compromise desired. In any
case the procedure is about the same. The length of the
valve stem is adjusted, then the eccentric position is
adjusted, and then refinements are effected by small changes
of both adjustments. Remember always, that changing
the length of the valve stem changes events at opposite
cylinder ends in opposite directions, while shifting the
eccentric changes all events in the same direction.

The effects of various adjustments are shown by the

THE D-SLIDE VALVE 195

indicator diagrams given in Fig. 128. These diagrams were
taken from a small, slide-valve engine and serve very well
to show the way in which the indicator discloses poor
adjustments.

PROBLEMS

1. Given: angle of advance, 30°; throw of eccentric, 1£
ins.; lead, ^ in.; maximum exhaust-port opening, If in.; find
the steam lap, maximum opening to live steam, and the exhaust
lap.

2. Given: steam lap of f in.; lead of ^ in.; exhaust lap of
f in.; and the angle of advance equal to 30°. Find the valve
travel (=2X throw of eccentric) and maximum port opening to
steam and to exhaust.

3. An engine has an eccentric throw of If ins.; a steam lap
of f in.; and a lead of j^ in. Compression begins at J of the
return stroke.. Assume a connecting rod of infinite length and
find the angle of advance, the exhaust lap, and the maximum
port openings to steam and to exhaust.

4. Given: valve travel, 3 ins.; steam lap, f in.; exhaust lap,
1 in. ; and lead, | in. ; find maximum port opening, angle of advance,
and piston positions at cut-off, release, compression, and admission
for both ends of cylinder, with the length of the connecting rod
equal to 4£ times the length of the crank.

5. It is required to build an engine having a steam-port opening
of f in., a lead of ^ in., and a connecting rod four times the length
of the crank. Cut-off must occur at f stroke and release at 95%
of the stroke. Find the inside and outside lap, the throw of the
eccentric and the fraction of stroke completed by the beginning
of compression,

CHAPTER XI
CORLISS AND OTHER HIGH-EFFICIENCY ENGINES

90. The Trip-cut-off Corliss Engine. The slide valve
has certain limitations which can be partly, but never
wholly, overcome. In most slide-valve gears, for instance,
the various events occur more slowly than is desirable,

and this is particularly
true of cut-off. Ideal
valves would open sud-
denly to full opening
when necessary and
would close as suddenly
at the proper time, and
such action would give
minimum throttling loss
and rounding of corners
of the diagram. Engines

fitted with such ideal valves would therefore give indicator
diagrams with maximum work area as shown by the
dotted lines in Fig. 129, the full lines indicating the type
of diagram obtained with the ordinary slide valve.

Again, the simpler forms of slide valve involve the use
of long ports connecting with the clearance space within
the cylinder, thus adding greatly to the clearance surface
exposed and to the cylinder condensation. These ports
serve for both admission and exhaust, and their walls are
therefore periodically cooled by the exhaust steam with the
result that excessive condensation occurs during admission.
Many attempts have been made to devise valve gears
which should not be subject to the limitations of the

196

FIG. 129.

HIGH-EFFICIENCY ENGINES 197

simple slide valve. Some of these have resulted in the
development of the more complicated slide valves de-
scribed in the last chapter, but such designs generally leave
much to be desired. One of the earliest and most success-
ful solutions was made by Corliss, who developed what is
known as the trip-cut-off Corliss gear.

The long combined steam and exhaust ports are elimi-
nated by the use of four valves, two for steam and two for
exhaust. These are rocking valves and are located top and
bottom, at the extreme ends of the cylinder, with their
longitudinal axes perpendicular to those of the cylinder,
as shown in Figs. 50, 51 and 52. The exhaust valves are
located below so as to drain out water of condensation.
Details of valves of this type are shown in Fig. 130.

These valves may each be regarded as an elementary
slide valve which has a cylindrical instead of a flat face,
and which is oscillated about a center near the face instead
of being reciprocated, i.e., oscillated about a center at an
infinite distance.

The valves are operated as shown in Fig. 131 by short
links from a wrist-plate pivoted on the side of the cylinder
and rocked back and forth about its center by means of an
eccentric operating through the linkage indicated. The
locations of the various pins and the lengths of the various
links are so chosen that the valves travel at high velocity
when opening and closing, that they open very wide, and
that they close only far enough to prevent leakage and then
remain practically stationary until about to open again.
Throttling losses are thus decreased and wear caused by
useless motion after closure is minimized.

The opening of the admission valves in this gear is
effected positively by the linkage already explained, but
they are closed differently. For opening, the steam link
rotates the bell crank B in Fig. 132 and thus raises the
latch C. The hook on the end of one of the arms of this
latch engages the steam arm which is fastened on the end

198

STEAM POWER

HIGH EFFICIENCY ENGINES

199

200

STEAM POWER

of a rod which is slotted into the end of the valve. The
valve is thus drawn further open as the wrist plate revolves,
until the tripping end D of the latch strikes the cam indicated
by E. This throws the hook out of engagement and thus
disconnects the valve from the driving mechanism. The

Safety Cam
o Ste'amArm,

FIG. 132. — Brails of Corliss Trip-Cut-off Gear.

valve is closed by tlie'action of a dash pot, one form of which
is shown in Fig. 131. As the steam arm rises during the
opening of the valve it draws up the plunger or piston of the
dash pot, leaving a partial vacuum beneath it, and, when the

20"x 48"neavy Duty Corliss

110 Lb. Steam

6i)R.P.M.

FIG. 133.

valve is released by unhooking of the latch, atmospheric
pressure drives the rjlunger down and thus causes cut-off
to occur.. The action of a dash pot is found to be unsatis-
factory when the speed of the engine exceeds about 125
R.P.M. and most Corliss engines with trip-cut-off operate

HIGH-EFFICIENCY ENGINES 201

at still lower speeds. Under such circumstances the cut-
off is very rapid as compared with the piston speed, and the
diagram shows a comparatively sharp corner at this point.
A set of diagrams obtained from a large Corliss engine
operating at 80 R.P.M. is given in Fig. 133, and it is obvious
that little throttling occurs.

Because of the low speed at which these engines operate
the stroke can be made long with respect to the diameter
without attaining a prohibitive piston speed. The economy
mentioned in Chapter VII as resulting from the use of long
strokes can thus be obtained in these engines. An idea
of the saving in steam effected by the partial elimination
of throttling and condensation losses by means of the
Corliss gear can be obtained from the curves in Fig. 134
(a) and (6), which give average performances.

The position of the cam which determines the time
at which cut-off occurs is controlled by the governor of
the engine. When moved in the direction taken by the
steam arm it causes cut-off to occur later. Variation of
the point of cut-off is used in these and in most other engines
to control the amount of work done per cycle in order that
the engine may make available the quantity demanded at
the shaft, as will be explained in a later chapter. It is there-
fore desirable that the range of cut-off should be as great
as possible, but it has been found very difficult to design
trip-cut-off gears which will give a cut-off later than about
0.4 stroke if steam and exhaust valves are operated from the
same eccentric. Later cut-off causes poor timing of the
exhaust events.

This has led to the introduction of Corliss engines
with two eccentrics and two wrist plates per cylinder.
One set operates the steam valves and the other the exhaust
valveSi With this arrangement the range of cut-off is
unlimited.

91. Non-detaching Corliss Gears. Because of the low
speed at which trip-cut-off Corliss engines are Qperated,,

202

STEAM POWER

HIGH-EFFICIENCY ENGINES

203

sqi-uopdransuoo tuuajg

204

STEAM POWER

they are necessarily large, heavy and costly and efforts
have been made to design gears which shall possess the
advantages of the original Corliss mechanism without
the limitation as to speed.

In many models the Corliss valves are retained and are
located in the end^s of the cylinder as just described or in

Double-Ported

Exhaust Valve.

Corliss Type

FIG. 135. — Non-detaching Corliss Valves Located in Cylinder Head.

the cylinder heads as shown in Fig. 135. In some the wrist
plate and the connecting links are also retained, but in
others they are eliminated. In all engines of this type the
admission valves are closed positively, the closure being
effected by the same linkage that opens the valves to admit
steam. Quick action is obtained by the arrangement of
the operating mechanisms, the centers of rotation and the

HIGH-EFFICIENCY ENGINES 205

lengths of links being so chosen that the valve travel is
small when the valves are closed, that it is rapid when the
valves are opening and closing, and that the valves remain
practically wide open during most of the time that steam
is being admitted.

The advantages of small clearance and short and sepa-
rate ports are attained in these arrangements and the
operation of the valves is almost as perfect as that of the
trip-cut-off gear. Engines fitted with these modified
Corliss gears are operated at speeds considerably higher
than those permissible with the older arrangement, and they
may be classed with medium-speed engines.

Engines of this type are generally known commercially
as four- valve engines, but as this name applies equally well
to the ordinary trip-cut-off gear and to others which will
be described later, it is best to use some other designation.
The term non-detaching Corliss engines seems to best
describe them and is apparently gaining in favor.

Non-detaching Corliss engines generally give diagrams
intermediate between those obtained with the low-speed,
trip-cut-off mechanism and those obtained from slide-valve
engines with the simpler forms of valves, though the later
designs very closely approximate the performances of the
trip-cut-off Corliss engine.

92. Poppet Valves. Attention has already been called
to the fact that the use of highly superheated steam is
very effective in lessening or even eliminating initial con-
densation. Experience has shown that large valves and
valves with sliding surfaces such as slide valves and Corliss
valves do not work well with highly superheated steam.
The large castings warp so that contact surfaces do not
remain true and the lack of moisture which acts as a seal
with saturated steam leads to excessive leakage. Dif-
ficulty has also been experienced with the lubrication of
these sliding types of valves when using highly superheated
steam.

STEAM POWEK

An old form of
valve known as the
poppet valve has re-
cently been adopted
by some builders as
a solution of the
difficulties met in
the use of highly
superheated steam.
This form of valve in
four-valve arrange-
ment, combined with
designs in which short
ports and symmet-
rical cylinder cast-
ings are used, yields
very economical en-
gines which can be
safely used with a
degree of superheat
prohibitively high in
the case of the slid-
ing and oscillating
forms of valves.

Receiver

FIG, 1366. — Cross-section, Lentz
Engine.

HIGH-EFFICIENCY ENGINES

207

Sections of a modern type of poppet valve engine are
shown in Figs. 136 (a) and 136 (6), and details of the
admission valve and its operating mechanism are given in
Fig. 137 (a) and (6). The valves are all double-seated
(double-ported or double-beat), that is, they seat at both
ends and are made hollow so that the steam passes both
around the outside of the valve and through the valve
as shown by the arrows in Fig. 137 (6). This results in
large area for passage of steam and in quick opening and

To Cylindur

FIG. 137a.— Admission Valve and Operating
Mechanism, Lentz Engine.

FIG. 1376.

closing, as in the case of gridiron valves, with small actual
movement of the valve.

The valves are opened positively by eccentrics opera-
ting through cams and rollers as shown in Fig. 136 (b) and
they are closed by springs as rapidly as the return motion
of the cam permits. The eccentrics are mounted on a
horizontal lay shaft which is located to one side of the
engine, with its axis parallel to that of the latter, and which
is driven by bevel gears from the crank shaft of the engine.

Since this valve arrangement gives short steam and
exhaust poits, permits the use of small clearance, and

208 STEAM POWER

gives fairly rapid opening and closing of valves with little
throttling when open, it gives good economy when used
with saturated steam. By adding superheat the economy
is still further improved. The water rate of one of these
engines is shown for one load in Fig. 134 (a). A simple,
Lentz non-condensing engine is reported to have given
a consumption of 16.13 Ibs. of steam per horse-power hour
with 92.7° superheat, and a pressure of 133 Ibs., and this
figure is materially lowered by compounding, higher super-
heat, lower back pressure, etc.

93. The Una-flow Engine. A very interesting modifica-
tion of the steam engine, known as the Una-flow Engine,
has recently appeared. In this design an attempt is made
to decrease the loss due to condenastion in a very original
way and the results of tests seem to indicate that the design
makes possible very great economy.

In the ordinary forms of engine the entire wall of the
cylinder is subjected to the cooling action of the lowest
temperature steam during the entire exhaust stroke, and in
double-acting types these cooled walls are immediately
brought into contact with the higher pressure steam acting
on the other side of the piston, as well as coming into con-
tact later with the next charge of high-pressure steam.
The una-flow design minimizes this action by admitting
steam at the ends of the cylinder, exhausting it at the
center of length of the cylinder, and compressing the steam
caught in the clearance up to a value approximating initial
pressure, thus heating the clearance walls. The heating
of the clearance walls is further effected by partly jacket-
ing the head with live steam on its way to the admission
valve and the jacket is sometimes extended along the
cylinder to the point at which cut-off normally occurs.

One form of this engine is shown in Figs. 138 and 139.
The steam enters the cylinder head from below, passes
up to a double-seated poppet valve, flows into the cylinder
until cut-off occurs and then expands until the piston

HIGH-EFFICIENCY ENGINES

209

uncovers the exhaust ports. The steam is exhausted
until the returning piston again covers these ports, after
which the material trapped within the cylinder is compressed
as indicated in the diagrams. The ideal sought is to main-
tain each part of the wall approximately at the tempera-
ture which the expanding steam will have when reaching
it and thus to minimize thermal interchanges and loss.

FIG. 138. — Section of Una-flow Engine Cylinder.

Tests of these engines show that very great ratios
of expansion can be used in a single cylinder without the
excessive losses customary when such ratios are attempted
in the ordinary counter-flow type. It is thus possible to
obtain good economy with one una-flow cylinder expanding
from a high pressure to a vacuum ; conditions which would
involve the use of compounding with ordinary construc-
tion.

210

STEAM POWER

The results of tests on one of the first Una-flow engines
built in this country are shown in Fig. 134 (a) and (6). In
comparing with the curves it should be noted that two of
the tests were run with high superheat.

94. The Locomobile Type. In the effort to improve
the ecomony of small steam plants the Germans developed
a form of plant now known as the Locomobile Type. The

FIG. 139.

name came from the fact that these plants, as originally
made, were mounted on wheels and intended for portable
use by agriculturists and contractors. Their economy in
the use of fuel proved so great that they have since been
built for stationary use in sizes running well toward 1000
horse-power per unit.

A locomobile of American construction known as the
Buckeye-mobile is illustrated in Fig. 140, which shows a
longitudinal section of the plant. The tandem compound

HIGH-EFFICIENCY ENGINES

211

212 STEAM TOWER

engine is mounted on top of an internally fired boiler with
the engine cylinders located in the flues which lead the
products of combustion away from the boiler.

The steam generated in the boiler is passed through
a superheater suspended in the smoke box. The flow of
steam is from the rear toward the front of this superheater
(counter flow) so that the hottest steam comes in contact
with the hottest gas. The steam then passes through a
pipe contained within the flue to the high-pressure cylinder,
which is jacketed by the hot flue gases and in which the
loss of heat to metal is thus minimized. From the high-
pressure cylinder the steam passes to a receiver contained
in the smoke box, the receiver serving as a reheater to
evaporate any condensate exhausted from the first cylinder
and to superheat the steam admitted to the low-pressure
cylinder. From the low-pressure cylinder, the steam
passes through a feed-water heater in which it raises the tem-
perature of the boiler feed and then it passes to atmosphere
or to a condenser. Boiler-feed pump and condenser pump,
if used, are also integral parts of the plant, being driven
directly from the main engine.

It will be observed that every precaution is taken to
guard against initial condensation, and to minimize loss
of heat in flue gases and in exhaust steam leaving the
plant. The high economies achieved are due to such
facts alone.

Small plants of this type have given an indicated horse-
power hour on a little over one pound of coal when oper-
ated condensing, whereas the best large compound recip-
rocating engine plants seldom do better than about 1.75
Ibs. of coal per I.h.p. and often use 2 or more pounds when
operated condensing,

CHAPTER XII
REGULATION

95. Kinds of Regulation. There are two distinctively
different kinds of regulation referred to in connection
with reciprocating steam engines, one of which may be
called fly-wheel-regulation and the other governor-regula-
tion or governing.

The regulating effect of the fly-wheel has already been
referred to. The turning effort exerted at the crank pin
by the action of steam on the piston or pistons of an engine
is not constant, and the angular velocity of the engine shaft
is therefore constantly varying during each revolution.
It is the function of the fly-wheel to damp these variations
so that they do not exceed the allowable maximum for
any given set of operating conditions. The efficiency of
the fly-wheel in this respect is measured by the coefficient
of fly-wheel regulation dw which is defined by the equation

- ir* j -.1 • I . , T T T T

j V max ' min /«A\

$W= ~^T~ ~~> ...•;. (66)

in which

T7max = maximum velocity attained by a point on

fly-wheel rim or other revolving part;
V mm = minimum velocity of the same point, and
F = mean velocity of the same point

' max ~i ' rnln • , i

approximately.

Governor-regulation is absolutely different. Its function
is to proportion the power made available to the instan-
taneous demand. The fly-wheel takes care of variations

213

214

STEAM POWER

occurring during the progress of one cycle, while governor
regulation varies the work value of successive cycles.

96. Governor Regulation. If the effect of engine
friction be neglected, the power delivered at the shaft of
the engine will vary directly with the indicated horse-
power. Such an assumption is accurate enough for the
discussion which follows.

The indicated horse-power of a given engine is deter-
mined entirely by the value of the mean effective pressure
and the number of cycles produced in a given time, since
these are the only variables in the formula for indicated
horse-power. The power made available by an engine
might therefore be varied by varying the mean effective

FIG. 141. — Throttling
Governing.

FIG. 142.— Cut-off
Governing.

pressure, or by varying the number of cycles produced in
a given time, or by a combination of both processes.

All of these possibilities are used. In ordinary station-
ary power plants the mean effective pressure is generally
varied. In the case of pumping engines, working against
a constant head, but required to deliver different quantities
of water at different times, the number of cycles per minute
is generally altered by changing the speed at which the
engine operates. In locomotive and hoisting practice
both the number of cycles per minute (speed) and the
mean effective pressure are varied as required to meet
the instantaneous demands.

These variations may be effected manually as by the
driver of a locomotive, in which case the engine may be
said to be manually governed. Or, they may be brought

REGULATION 215

about mechanically, as in the case of most stationary power-
plant engines, in which case the engine may be said to be
mechanically governed. In some instances a combination
of manual and mechanical governing is used.

97. Methods of Varying Mean Effective Pressure. The
mean effective pressure increases and decreases with the
area of an indicator diagram of constant length, so that
the mean effective pressure can be changed by any method
which will change the area of the diagram. Two methods
are in use and they are illustrated in Figs. 141 and 142.
The first causes a variation in area by changing the value
of the initial pressure. This is generally done by chang-
ing the opening of a valve in the steam line just outside of
the steam chest. It is called throttling governing, and the
valve is called a throttling or throttle valve. The latter
name is also commonly used for the valve located near the
engine, which is used to shut off the supply of steam entirely
when the engine is not in operation.

The second method, illustrated in Fig. 142, is known
as cut-off governing. The variation of cut-off determines
the amount of steam admitted to the cylinder per cycle
and is used to measure out the quantity required for the load
which happens to exist at any instant. Cut-off governing
is used on most modern stationary engines and is exclusively
used in large reciprocating engine power plants.

98. Constant Speed Governing. Most engines used for
such purposes as the operation of mills and the driving of
electrical and centrifugal machinery are required to run
at practically constant speed irrespective of the load. They
are furnished with mechanical governors which so regulate
the power made available that there shall never be any
appreciable excess or deficiency which would respectively
cause an increase or a decrease in speed.

These mechanical devices always contain some sort
of tachometer which moves whenever the speed of the engine
exceeds or falls below the proper value. The tachometer

216 STEAM POWER

is so connected to the valve gear that it decreases the
power-making ability of the engine whenever the speed
starts to increase and it increases the power-making ability
if the speed drops.

Since the valve gear must have a different position
for each load in order that it may throttle or cut off as
necessary to suit that load, it follows that the tachometer
which controls the position of the valve gear must also
have different positions for different loads. But tachom-
eters assume positions dependent on speed, and therefore
different loads can only be obtained if the tachometer and
the engine to which it is connected operate at different
speeds for different loads.

Constant-speed governing is therefore an anomaly.
The device which is supposed to maintain constant speed
irrespective of load must be operated at different speeds,
as the load varies, in order that it may maintain the valve
gear in the different positions required to handle the differ-
ent loads. All so-called constant-speed engines have their
highest speed when carrying no load, and the speed gradually
decreases to a minimum as the load increases to a maxi-
mum. The total variation is generally between 2 and 4%.

The efficiency of a governor in this respect is measured
by means of the coefficient of governor regulation, do,
which is defined by the equation

in which

n2 = highest rotative speed attained by the engine;
m = lowest rotative speed attained by the engine, and
n = mean speed

ri2+n\
= — ;r — approximately.

99. Governors. The mechanical devices which are
used for controlling the power-making ability of an engine

REGULATION

217

as described above are known as governors. There are
many varieties and only a few of the more prominent can
be described.

(a) The Pendulum Governor. One of the earliest forms
of governor used on steam engines is illustrated in Fig.
143. It is often called a
fly-ball governor. This
governor is driven by
gearing, chain or belt
from the engine, and the
weights assume some
definite position for
each different speed, thus
drawing the collar to
different positions. The
valve gear is connected
to this collar and is
moved correspondingly.

A similar governor
is shown in Fig. 131,
which also indicates the

way in which the collar is connected to the valve gear in
the Corliss type of engine. The governor rods are moved
as the collar moves and they in turn alter the position
of the knock-off cam, and thus vary the time at which
cut-off occurs. As the speed increases due to a decrease
of load, the governor weights and collar move up, and this
shifts the cams so as to produce earlier cut-off and decrease
power-making ability.

(6) Shaft Governors. On medium- and high-speed
engines fitted with some form of slide valve it is found best
to use what are known as shaft governors. They are gen-
erally carried within the fly-wheel of the engine, operate in
a plane passing through the rim of the wheel at right angles
to the shaft, and operate upon the eccentric in such a way
as to vary the cut-off with speed (and load) changes.

FIG. 143.

218

STEAM POWER

FIG. 144.

One simple form of such a governor is shown in Fig.
144. The eccentric is not mounted directly upon the

engine shaft, but is carried
by a pin P in the fly-wheel
and is slotted so that it can
swing back and forth across
the shaft, about P as a center.
Its position at any time is
determined by the position
of the governor weights W,
which draw the eccentric
down (in the figure) as they
move out.

The center of the eccentric
is indicated by a heavy dot
in the figure, and it will be seen that this center would
travel in the arc of a circle about P, as the weights moved.
If the path of the eccentric center is drawn on a Bilgram
diagram, it will be found that this motion is equivalent to
decreasing the length of the eccentric crank and increasing
the angle of advance, resulting in earlier cut-off as the
weights move out with increasing speed and decreasing
load. Other events will also be changed as the eccentric
swings, and some of these changes are occasionally unde-
sirable.

Numerous designs have been developed in which the
eccentric is so guided as to produce various sorts of rela-
tions between the different steam and exhaust events.
All can be divided into two classes, those in which the
eccentric swings about a fixed center variously located, and
those in which the center of the eccentric is guided to
move in a straight line. All can be studied by plotting
the path of the eccentric center (path of Q) on the Bilgram
diagram.

The Rites Inertia Governor is a form of shaft governor
so designed as to act very quickly with change of speed,

REGULATION

219

and to be very powerful, so that it can shift heavy parts.
It is shown in place in the wheel in Fig. 145. With changes in
speed it acts like a governor of the type just described,
swinging (with increasing speed) about a fixed point P in
the wheel as its center of gravity G moves outward under
the action of the centrifugal effect C and against the
action of the spring. This motion shifts the center of the

FIG. 145.

eccentric from E toward c, giving the desired variation in
cut-off.

Superposed upon this action is that of inertia. Assume
the wheel and governor to be rotating clockwise at a given
constant speed. If the engine speed is suddenly increased,
the wheel will move faster, but the governor bar will tend
to continue rotating at the same speed because of its inertia.
It will thus lag behind the wheel, rotating about P and bring-
ing about an earlier cut-off. The position thus assumed
will later be maintained by centrifugal effect if the new speed

220 STEAM POWER

is maintained. The particular advantage resulting from
using inertia in this way is speed of action. In many forms
of governor the inertia of the moving parts actually resists
the efforts of the governor to assume the new position
required by changed load and speed.

CHAPTER XIII
THE STEAM TURBINE

100. The Impulse Turbine. One of the oldest of modern
water wheels is the tangential or impulse wheel shown
diagrammatically in Fig. 146. Water flowing from a
reservoir above the wheel passes through a nozzle and the

FIG. 146. — Tangential or Impulse Wheel.

jet, moving at high velocity, strikes buckets on the rim
of the wheel and causes the latter to revolve. Theoretically
the velocity of the water in the jet would be

v = V2gh feet per second, . . . « (68)
in which

g = gravitational constant, 32.2, and
h = head in feet as shown in the figure.

The kinetic energy possessed by the moving water would
be

* = f 4 111 . . . (69)
221

222 STEAM POWER

in which w represents pounds of water discharged per second
and g and v have the same meanings as above.

If the buckets of the wheel could reduce the velocity
of the water to zero they would absorb all of this kinetic
energy and (assuming no losses within the buckets and the
bearings of the wheel) would make all of it available at
the shaft for the doing of useful work.

Any fluid moving at velocity v and striking buckets in
the form of a jet would possess kinetic energy in quantity
given by Eq. (69) and would drive the wheel in the same
way. Steam might therefore be used instead of water
with exactly the same results, and steam is so used in what
are known as impulse steam turbines.

Experience shows that steam will flow at high velocity

from any opening made in the
steam space of a boiler or
from any open-ended pipe con-
nected to such a boiler. This
is commonly said to be due
to the high pressure within
the boiler, the spectator pic-
turing the process as the
driving out of part of the

steam by the high-pressure steam within the boiler, just as
though the part leaving were a solid piston and were driven
out as is the piston of an engine during
admission, as shown in Fig. 147, ^

An hydraulic analogy is given in Fig.
148. The vessel shown is supposed to be
fitted with a piston, and it is assumed to be possible to exert
any desired pressure upon the piston. Any such pressure
exerted is the exact equivalent of some given head of water
and the resultant jet velocity would be given by Eq. (68)
by substituting for h the head in feet equivalent to the
pressure exerted upon the piston.

When an " elastic " fluid such as steam is being con-

THE STEAM TURBINE 223

sidered it is, however, necessary to take account of other
factors. The steam within the boiler exists at a high
pressure ; after issuing it exists in the atmosphere at a lower
pressure. But low-pressure steam contains less heat than
does steam at high pressure, and this difference must exist
in some form, as it is energy and could not possibly have been
destroyed during the flow.

Experiment shows that steam after flowing into the
atmosphere from a boiler in this way has exactly the same
characteristics as though it had expanded adiabatically
behind a piston through the same temperature range, ex-
cepting for the fact that it has a very high velocity, which
it would not possess if expanded behind a piston. Experi-
ment further shows that, if small losses be neglected, the
kinetic energy possessed by a jet of steam is exactly equal
to the energy which would be turned into work if that steam
acted on a piston as in an ordinary engine.

A complete picture of the process of flow can then
be made by assuming the steam flowing out in the form of
a piston driven by high-pressure steam, as before, and adding
to this the idea that this piston expands adiabatically as
it travels from the region of high to that of low pressure.
This expansion liberates heat contained within the piston
or plug of steam and this heat is used in imparting addi-
tional velocity to the moving steam which is giving up this
heat.

The result of using such a jet upon a theoretically
perfect tangential or impulse wheel would be to rob the
jet of all this energy. But the energy possessed per pound
of steam in the jet i^-just the same as that shown under the
upper lines of a complete expansion cycle using one pound
of steam. The area under the upper horizontal line of the
PF-diagram of the cycle as shown in Fig. 21 may be assumed
to represent the work done upon one pound of steam (flow-
ing out) by another pound which is being evaporated and
pushing out the first in order to make room for itself. The

224

STEAM POWER

area under the expansion curve in the PF-diagram repre-
sents the energy converted into velocity energy by the adia-
batic expansion of the flowing steam. The lower hori-
zontal line represents the negative work during condensa-
tion to water at the lowest pressure and temperature, and
the left-hand line represents the pumping of this water back
into the boiler and the raising of its temperature to the value

i>ianbragm

FIG. 149. — Early Form of Impulse Turbine.

maintained within the boiler. The complete expansion cycle
is therefore the cycle . upon which the impulse steam turbine
operates and, as a matter of fact, it is the theoretical cycle
of all steam turbines.

The ideal impulse turbine would therefore be acted
upon by a jet which possessed available kinetic energy
represented by the area of the complete expansion cycle.
If the buckets could entirely remove this energy, that is,
could reduce the velocity of the jet to zero, the same amount

THE STEAM TURBINE

225

of energy could theoretically be made available at the shaft
of the turbine.

An example of a simple form of impulse steam turbine
is given in Fig. 149, in which the essential parts of an early
form of Kerr turbine are shown. The wheel, the diaphragm
and nozzles are all inclosed within a casing. The space
on one side of the diaphragm is connected to the steam pipe
and that on the other is in communication with the space
into which the exhaust steam is to be exhausted.

Another form of impulse turbine is shown in Fig. 157.
It will be described later.

101. Theoretical Cycle of Steam Turbine. It was shown
in the preceding section that the steam turbine operates
on the complete expansion cycle. If a turbine could remove
from the steam passing through it and convert into mechan-
ical form all of the energy which is theoretically possible,
it would therefore make available mechanical energy
represented by the area of the PF-diagram of the complete
expansion cycle. The area of the corresponding T<f>-
hagram would show the
Same quantity measured in
thermal units. The theory
of the steam turbine can
therefore be studied by
means of these two dia-
grams.

In Fig. 150 is shown the
T^-diagram of the complete
expansion cycle for several
different conditions. The
figure abed represents con-
ditions when the steam is
dry and saturated at the beginning of the adiabatic ex-
pansion cd. Constant quality lines are designated by x
and x'. It is obvious that by the time the steam has
expanded down to the pressure at d it will have a quality

FIG. 150.

226 STEAM POWER

less than unity. If, therefore, it be in the form of a jet
issuing from a nozzle and having a high velocity by virtue
of its adiabatic expansion, the jet will really be a mixture
of steam and water.

If the steam be superheated at constant pressure as shown
by ce before passing through the nozzle, it is evident from the
figure that the jet issuing from the nozzle will contain
less water than in the preceding case, because the condition
of the material in the jet after adiabatic expansion will be
as shown at / instead of as shown at d. The cycle in such
a case would also be larger by an amount indicated by the
area cefd, representing just that much more heat converted
into mechanical energy per pound of steam or other unit
for which the diagram happened to be drawn.

If superheating had been carried to the point indicated
by g before expansion, the jet would obviously issue from
a nozzle in the form of superheated steam as shown by the
point h in the figure. In that case the cycle would be
abcgha, and superheat would have to be removed from the
low-pressure steam to bring it to the conditions indicated
at i before condensation could begin.

If desired, the PF-diagrams for such cycles can be drawn
very easily. The line be, or be or by is a horizontal line in
the PF-diagram. The line ha is similarly horizontal and the
line ab is vertical. The adiabatic expansion is represented
by a curved line in the PF-diagrams, but can be drawn
easily because the necessary data are obtainable from the
T<£-diagram, in which this expansion is represented by a
straight line.

ILLUSTRATIVE PROBLEM

Draw the PF-diagram for a steam turbine receiving one pound
of steam at a pressure of 200 Ibs. absolute, with a tempera-
ture of 500° F. and exhausting against a pressure of 0.5 Ibs.
absolute.

First, locate on a TV-chart for steam the point representing
the condition of steam at 200 Ibs. pressure with a temperature

THE STEAM TURBINE

227

of 500° F., and draw a vertical line extending downward
until it cuts the horizontal temperature line corresponding to

3 O

•sqyui'bg aad 'sqi-

0.5 Ib. pressure. This is practically at 540° F. absolute, or about
80° F.

228 STEAM POWER

Second, take from the steam table the volumes of one pound
of steam at, say, 200 Ibs., 140 Ibs., and 100 Ibs. absolute pressure
when superheated to the values shown by this vertical line. These
will be about 2.75 cu.ft., 3.58 cu.ft., and 4.67 cu.ft., respectively.
Plot these volumes with corresponding pressures on a PF-chart
as shown in Fig. 151.

Third, take from the !F</>-chart the pressures at which the
vertical line intersects different volume lines in the wet steam
region and plot volumes against pressures on the TV-chart.

Fourth, draw a smooth curve, as shown, through all points
so determined.

Fifth, draw horizontal top and bottom lines and a vertical
line at the left of the diagram. This vertical line should be to
the right of the pressure axis by an amount representing the
volume of one pound of water, but the volume is so small that
it cannot be plotted to any ordinary scale.

102. Nozzle Design. It was stated in preceding sec-
tions that the energy which would be converted into work
by the introduction and adiabatic expansion of steam
behind a piston is converted into kinetic energy when steam
flows out of an orifice or nozzle and that an ideal impulse
turbine could absorb all this kinetic energy from the jet,
bringing it to rest and making the energy available in the
form of useful power at its shaft. It is, therefore, of interest
to determine the velocity which a jet will acquire under
different conditions.

This could be done by evaluating the area of a diagram,
such as that of Fig. 151, and then putting this value in
place of K in Eq. (69) and solving for v, but it can be done
much more accurately and expeditiously in other ways.
The heat energy which can be converted into kinetic
energy of the moving jet and which can later be con-
verted into useful work by the turbine wheel is represented
by the area enclosed within the lines of the complete
expansion cycle when drawn on the T^-diagrams. That
is the area abed in Fig. 152, for instance, for the case of
wet steam at the beginning of expansion. But this area
is equal to that representing the heat supplied minus

THE STEAM TURBINE

229

that representing the heat rejected, that is, Qi — (J2, so
that

X(in B.t.u.)=Qi-Q2. . . . > (70)

The values of Qi and Q2 can be found very readily by
plotting the points c and d upon a jT</>-chart for steam and
observing the constant heat lines upon which they fall,

FIG. 152.

or they can be obtained even more conveniently from what
is known as a Mollier Chart for steam. In this chart,
entropy above 32° F. is plotted against heat above 32° F.
as shown in Fig. 153. An adiabatic expansion on this chart
is shown by a horizontal line, since this shows a constant
entropy change just as a vertical line on the T<j> chart shows
a constant entropy change.

If a point is found in this chart giving conditions corre»
spending to those at point c in Fig. 152, the value of Qi

230

STEAM POWER

THE STEAM TURBINE 231

can be read directly under that point on the horizontal
axis. A horizontal line drawn from that point to the
terminal-pressure line will give the point corresponding to
d of Fig. 152 and the value of $2 can be read on the hori-
zontal axis immediately below that point. The difference
between the two readings gives the value of the kinetic
energy K or of the mechanical energy which an ideal tur-
bine could make available, but the expression will be in
British thermal units and not in foot-pounds.

This value of the kinetic energy, i.e., K = Q\ — Qz, may
then be placed in Eq. (69), giving,

3)=-ft.-lbs., . . . (71)
since Qi and $2 refer to one pound of steam, or

Q2)=y-ftAbs., . . . (72)

when w represents the number of pounds of steam flowing
per second.

Solving*cither Eq. (71) or Eq. (72) for v gives,

v = V778X2g(Qi-Q2)

Q1-Q2 feet per second. . . (73)

The design of a nozzle consists simply in choosing such
sections that the desired amount of steam may flow through
it with the desired pressure drop, as the velocity obviously
is determined by that pressure drop. This is very con-
veniently done by working in terms of one pound of steam,
since all formulas and charts are generally given on that
basis, and then multiplying the cross-sectional areas found
by the number of pounds of steam required.

Assume for instance that it is desired to design a nozzle

232 STEAM POWER

to pass one pound of steam per second with an initial pres-
sure of 100 Ibs. per square inch abs., and a terminal
pressure of 60 Ibs., the steam being initially dry and satu-
rated.

The Mollier chart shows that Qi is equal to about 1187
B.t.u. per pound of steam, while Q% is equal to about 1147
B.t.u. The velocity with which a jet would issue from a
theoretically perfect nozzle under these conditions may
then be found by using Eq. (73). This gives

v = 224 Vl 187 -1147
= 1416 feet per second.

The shape of the entrance end of the nozzle is generally
made such that the steam will enter it without great dis-
turbance and the shape beyond that point is determined
by methods which will be explained below. The cross-
section of the discharge end must be such as to pass the
required quantity at the velocity found above to be equal
to 1416 feet per second. This is easily done by deter-
mining the volume of steam discharged.

Drawing the adiabatic expansion on the Tc/>-chart will
give the quality at the end of the expansion; or, the quality
can be determined by finding what quality a pound of
steam at 60 Ibs. pressure must have to give it a heat content
of 1147 as found above. With the quality known the ter-
minal volume per pound can be found by multiplying the
quality by the specific volume at terminal conditions.
Thus for the case under discussion the quality will be
about 96.7% and as the specific volume at 60 Ibs. is 7.17
cubic feet, the volume to be passed per second, per pound
of steam is 0.967X7.17 = 6.94 cu. ft approximately. If the
velocity is 1416 feet per second the area per pound of steam
must be 6.94 -^ 1416 = 0.0049 sq.ft.

The exact shape of the nozzle is determined by deciding
upon the way in which pressure, or velocity, or volume

THE STEAM TURBINE

233

shall change as the steam passes through it. Suppose, for
instance, that a nozzle is to be constructed of the length
shown by ab in Fig. 154, and that the pressure is to vary
along its length as shown. Assume also that the nozzle
is to pass 10 Ibs. of steam per second. Taking initial
pressure as 100 Ibs. and terminal as1 60 Ibs., the conditions

Pressure, Lbs. per Sq. In.

§ 2 g 88

1400
1200
1000
800

400
200

.__ —

— •

/•

^^-"

„ '

^--^

riati

an o:

Vel(

>city

T.f»ii«rt.h r»f "NToy/lfi

FIG. 154. — Nozzle Design.

will be the same as in the problem above. The discharge
area will have to be 10X0.0049 sq.ft. or 0.049 sq.ft.

The area at the plane x% must be that required to pass
the steam when it has the velocity resulting from expansion
from 100 down to 64 Ibs., just as though the nozzle ended at
that point. This can be found just as the terminal area
was found above. Similarly the sections at x\, and x can
be found by figuring velocity and area for expansions to

234

STEAM POWER

74 and 90 Ibs., respectively. If the various areas required
are determined in this way, the nozzle will have a longi-
tudinal section about as shown by the dotted lines in the
figure and the variation of velocity will be about as shown
by the curve.

If the shape of a nozzle is determined in the same way
for a case in which the terminal pressure is less than about
0.58 of the initial pressure, the nozzle will be found to have

a very different shape. This is
shown in Fig. 155. The nozzle
is known as an expanding nozzle
and the smallest section is known
as the neck. The pressure Pn
in the neck is always equal to
about 0.58 PI and the velocity
in the neck is always equal to
just over 1400 feet per second.
It is therefore the section at
the neck which determines the
quantity of steam which a nozzle
will discharge if expanding to a
pressure equal to or lower than
0.58 Pi.

103. Action of Steam on
Expanding Nozzle. ^^ ^^ R hag ^

stated that the steam acting in an impulse type of turbine
delivers energy to the wheel of the turbine by giving up
its kinetic energy. In an ideal turbine the steam jet would
be brought to rest and would thus give up all of its kinetic
energy.

In real turbines it is impossible to bring the jet to rest,
as practical design problems prevent it. There is there-
fore always a loss in real machines because of the residual
or terminal velocity of the steam as it leaves the wheel.
Thus let the black section in Fig. 156 represent the section
of a bucket or blade sticking out radially from the rim of a

THE STEAM TURBINE

235

wheel, the wheel revolving about the axis indicated by the

dot dash line but located behind the plane of the paper.

If minimum loss by eddying is to be experienced at the

point at which the steam jet

enters the blade, the jet must

enter the blade along a tangent

to the curve of the inside of the

blade at the entrance edge. This

direction is shown by the line

marked vr in the figure.

Were the bucket stationary,
the steam jet would move as
shown by iv, but as the bucket
moves ahead, and, so to speak,
runs away from the jet, the
steam must really travel in a
direction such as that indicated
by va in order to strike the
bucket in the direction indicated
by vr. The conditions governing
the flow of steam into a bucket
are the same as those governing the speed with which and
direction in which an individual runs toward and jumps
upon a moving vehicle. He will experience least shock
when he is moving ahead at the same rate as the vehicle
at the instant when he gets on board. His motion must
therefore be made up of two, one toward the vehicle and
the other in the direction of the vehicle's travel.

In the case of steam flowing onto a blade as shown in
Fig. 156, the various velocities are so related that when
drawn to scale the real or absolute velocity of the steam,
va, and the real or absolute velocity of the blade, vb, form
two sides of a triangle of which the closing side represents
vr) the velocity of the steam relative to the bucket. The
value and direction of vr is obviously obtained from va by
geometrically subtracting the velocity of the bucket.

FIG. 156

236 STEAM POWER

After entrance, the steam flows around the inner curve
of the blade and is finally discharged with the same rela-
tive velocity as that with which it entered, and at an angle
set by the tangent to the inner curvature of the discharge
edge of the blade as shown by VR. But, since the steam
has been moving ahead with the same velocity as the
bucket during the entire time that it was in contact with
the bucket, it is also moving ahead with a velocity Vb when
it leaves the wheel. Its real or absolute velocity is then
VA, which is found by combining VR and VD as shown in the
figure.

The kinetic energy possessed by the jet when entering

*/'/' 2>

the blade is equal to -~- ft.-lbs., and that which it possesses

when leaving is * . Obviously, the energy removed

U'l • J.U Ul J • < Tf 4-U Ul J

while passing over the blade is -^ --- ~ — '• I* the blade

were theoretically perfect, it would be so constructed that
rA2 would be zero and all of the kinetic energy would then
l)e removed. This is practically impossible in a real mechan-
ism, and there is always a loss due to the residual velocity
VA. The best that can be done is to so choose the angle
of jet and blade, and the velocity of blade with respect
to the steam that the actual numerical value of VA is made
as small as possible.

Designs usually work out in such a way that this occurs
when the blade velocity is equal to about 0.47 of the abso-
lute velocity of the steam jet.

104. De Laval Impulse Turbine. The expanding nozzle
already described was first used by De Laval in an impulse
type of turbine. The essential elements of this device are
shown in Fig. 157. The nozzles are arranged at such an
angle to the plane of the wheel that the steam jets strike
radially arranged blades at the proper angle to enter without
much loss. The blades deflect the jets as shown and

THE STEAM TURBINE

237

absorb the greater part of their kinetic energy, so that
the steam flows away from the wheel with low absolute
velocity.

As many nozzles are used as are required to make avail-

Nozzle

Turbine Shaft

Steam In

Nozzles -
FIG. 157. — Single Stage, De Laval Impulse Turbine.

able the amount of energy desired at full load, and pro-
vision is made for shutting off one or more nozzles by hand
when conditions do not warrant the use of all. Governing
for ordinary variations of load is effected by throttling
the steam flowing to the nozzles in use, thus altering the
initial pressure as necessary.

238 STEAM POWER

A section through the wheel and casing of such a tur-
bine directly connected to a centrifugal pump is given
in Fig. 158. The steam flows into the live steam space
through a throttle valve controlled by the governor; the
valve and connections are not shown in the illustration.
From the live steam space the steam flows through nozzles
not shown, and into the exhaust steam space, thus acquir-
ing a high velocity. The buckets of the wheel are located
just in front of the discharge ends of the nozzles and the
steam moving at high velocity must pass through them
before moving on toward the exhaust outlet.

105. Gearing and Staging. It has been stated that the
most efficient operation with ordinary designs is obtained
when the blade speed is equal to about 0.47 of the absolute
steam velocity or, roughly, half the velocity of the imping-
ing jet. To get high economy in the use of steam, large
pressure drops are used and very high jet velocities result.
When the buckets of a turbine are operated at peripheral
speeds equal to half these jet velocities one of two diffi-
culties is often met. The stresses induced in the wheel
structure by centrifugal effects become so high as to offer
serious difficulties in design, or the rotative speed of the
unit becomes too high for direct connection to the machine
which is to be driven.

One method of partly overcoming the latter difficulty
is to operate the turbine at or near the theoretically desir-
able speed and transmit the power to the driven machine
through gears which decrease the rotative speed to the
necessary extent. This method was used with all of the
early De Laval turbines which were of comparatively small
capacity. It is now being successfully applied to marine
propulsion and other purposes for which large units are
used. It is only a partial remedy in the case of large units,
however, as the gears necessary for the desired reduction
and the size of the turbine wheels would both become
excessive,

THE STEAM TURBINE

239

240 STEAM POWER

Another and very common method is known as com-
pounding or staging. This may be of two varieties. The
pressure drop in each stage may be limited to that
which will give a reasonable velocity and a number of
such stages may be put together in series on one shaft.
This would give one set of nozzles and a wheel for each
stage, the steam discharged from one wheel with very
low velocity expanding to a lower pressure through the
nozzles of the next stage and impinging upon the
wheel of that stage with the resultant high velocity.
Such an arrangement is known as pressure staging or
pressure compounding, and is extensively used in large
turbines.

The pressure staging method is illustrated in Fig. 159
as applied to the De Laval type of impulse turbine. The
combined increase in diameter of wheels and increase
in length of blades gives the necessary increase in area to
pass the larger volumes of steam as the drop of pressure
continues from stage to stage.

Instead of staging on a pressure basis, staging on a veloc-
ity basis may be used. In such a case the drop in pressure
through one set of nozzles is great and the resultant veloc-
ity high. The steam moving at this high velocity is then
directed upon the buckets moving at such peripheral velocity
that they absorb only part of the kinetic energy of the steam,
discharging it with a lower absolute velocity than that
with which it entered, but one which is too high to be
thrown away. The steam then passes through a set of
stationary vanes which direct it upon the blades of a second
wheel, in passing through which it gives up still more of
its kinetic energy with a corresponding further decrease
of velocity. If the velocity still possessed by the steam
warrants it, a second set of stationary guide vanes and a
third set of moving buckets can be supplied for further
reducing it and by carrying this velocity staging through
a sufficiently great number of stages any initial velocity

THE STEAM TURBINE

241

could be absorbed theoretically without the use of wheels
with high peripheral speeds. Practically, losses due to

friction, eddying and other sources limit the number of
velocity stages to two or three.

242

STEAM POWER

FIG. 160.— Early Form of Curtis Turbine.

THE STEAM TURBINE 243

Velocity staging is combined with pressure staging
in the Curtis type of turbine. A section through part
of an early design of vertical turbine of this type is shown
in Fig. 160. The turbine illustrated had four pressure stages
and each pressure stage had two velocity stages.

Many varieties of impulse turbines have been developed
and all of the larger ones employ several wheels and sets
of nozzles and diaphragms to obtain the necessary staging.
The same result has been obtained in some of the smaller
models by discharging the steam from nozzles on to a set
of buckets which are able to absorb only a fraction of the
kinetic energy, catching it at discharge and returning it
for another passage through the buckets, and so on until
the greatest practical fraction of the kinetic energy has been
absorbed.

106o The Reaction Type. If high-pressure ' steam or
other fluid be forced into a de-
vice arranged as shown in Fig.
161 and free to revolve about
a vertical axis, the jets blowing out
of the nozzles will cause the mecha-
nism to revolve in the direction
indicated by the arrow. This rota- FIG. 161.

tion is said to be due to the reaction Elementary Reaction
of the jets, and the mechanism there- Turbine,

fore constitutes a simple form of reaction turbine. By
increasing the number of nozzles
any amount of steam could be dis-
charged and therefore any amount
of work could be obtained.

This multiplication of nozzles
can, however, be more conveniently
accomplished by fastening radial
JTIG 162. vanes to the periphery of a wheel

as shown in Fig. 162, the space
between any two vanes constituting a nozzle through which

244

STEAM POWER

FIG. 163.

the steam can discharge. By mounting such a wheel
within a casing as shown in Fig. 163 it forms a simple
reaction turbine. One of the characteristic differences
between the impulse and the reaction
types lies in the distribution of pressures.
In the impulse type the nozzles are
fastened into a stationary part of the
turbine and the drop of pressure occurs
entirely within the nozzles. The wheels
are therefore immersed in a space in
which a uniform, low pressure exists.
In the reaction type, on the other hand,
the nozzles are carried on the wheel and
there must be a higher pressure on one side of the wheel
than there is on the other. Since there must also be me-
chanical clearance between the blade tips and the interior
of the casing, it follows that the reaction type will be
handicapped by considerable leakage which does not exist
in the impulse type, excepting as some of the jet may
" spill " over the ends of the blades.

The difference of pressure on the two sides of the wheel
also causes a tendency toward motion of the wheel along
the shaft, or of the wheel and shaft, in a direction away from
the higher pressure.

Many unsuccessful efforts have been made to design
efficient reaction turbines, but no pure reaction type has
yet been commercialized. The turbines commonly called
reaction turbines are really combinations of reaction and
impulse types.

One example of what is commercially called a reaction
turbine is shown in Fig. 164. Alternate rings (or rows)
of stationary and movable blades guide the steam as it
expands from the high pressure at one end to the low pres-
sure at the other. The stationary blades project inward
from the interior surface of the stationary casing and the
movable blades project outward from the external surface

THE STEAM TURBINE

245

246

STEAM POWER

of the cylindrical rotor. The rotor blades act like those
of an impulse turbine in partly reversing the direction of jets
of steam which reach them with comparatively high veloci-
ties, but they also act like the movable nozzles of a reac-
tion turbine since the steam in passing through them expands
and acquires kinetic energy, giving a reaction on discharge.
The stationary blades serve to redirect the steam so that
it strikes the next set of moving blades at the proper angle

and they also serve as
nozzles in which velocity
energy is acquired. This
is shown diagrammatical ly
in Fig. 165, in which S
denotes stationary, and M
movable blades.

The Parsons type, il-
lustrated in Fig. 164, may
be described as a multistage
type in which impulse and
reaction are utilized in con-
junction.

The balance pistons
shown in the figure are
used to balance the end
thrust caused by the differ-
ence in pressure existing on
opposite sides of the wheels
in the case of reaction turbines. Each piston is of such
a diameter that it presents a surface equal to the blade
surface acted upon by one of the unbalanced pressures,
and by connecting across as shown in the figure a high
degree of balance is secured.

The overload valve is used to admit high-pressure
steam to the low-pressure blades for carrying excessive
overloads. The larger area of the passages through these
blades permits an abnormal amount of high-pressure steam

THE STEAM TURBINE 247

to pass, thus giving a high load-carrying capacity with
decreased economy.

107. Combined Types. The clearance at the ends of
the stationary and moving blades in the Parsons type of
turbine permits considerable steam to leak by, as previously
explained. This clearance must have almost the same
length (measured from blade tip to opposing metal) in all
stages in order to insure freedom from rubbing, but it is
more detrimental in the high-pressure stages than in the
low. The high-pressure blades are much shorter than the
low-pressure blades and a leakage length of a certain amount
is therefore equal to a greater fraction of the total blade
length. The density of the high-pressure steam is also so
much greater than that of the low-pressure steam that many
more pounds can leak through an opening of a given size
in a given time. In discussions of this character, it should
not be forgotten, however, that leakage area is determined
by the dimension already referred to multiplied into a
circumference and that the circumference is much greater
at the lower end.

Because of these and other reasons many manufacturers
have come to the conclusion that the impulse type is best
for the high-pressure end of the turbine and the reaction
type for the low-pressure end. Many such combinations
have been produced and they are giving very good results.

108. Economy of Steam Turbines. In general, the
economies of steam turbines and reciprocating engines are
about the same when each type is operated at normal load
and under the best conditions. It is probable that very
large turbines have a slight advantage over reciprocating
engines (as generally built) in the matter of economy and
the reverse of this statement appears to be true for most
small units, although very economical turbines have been
produced in small sizes in the past few years.

The turbine, however, generally gives a flatter water-
rate curve than does a reciprocating engine; that is, for

248 STEAM POWER

loads each side of the most economical the steam per horse-
power hour does not increase above the value attained
at most economical load as rapidly in the case of turbines
as it does in the case of most reciprocating engines. With
a very variable load, therefore, or with a load which is far
removed from the rated value, the turbine probably gives
a better average performance than does the reciprocating
engine. This is particularly true in large sizes.

It has been shown that the turbine operates on the complete-
expansion cycle and it will be remembered that the recipro-
cating engine operates on a cycle with incomplete expansion.
The turbine is therefore able to make better use of very
Ipw-pressure steam than can the piston type.

Trial on a T</> or Mollier chart will show that a turbine
receiving steam at about atmospheric pressure and expand-
ing it down to a vacuum of from 28 to 29 ins. should make
available as much work as one receiving steam at a high
boiler pressure and expanding down to atmospheric. In
other words a drop of 100 Ibs. or more above atmospheric
pressure makes no more energy available than does a drop
of about 13 Ibs. below atmospheric, or the lower the initial
pressure the more heat is converted into work by a given
pressure drop. A small decrease in back pressure (terminal
or condenser pressure) is therefore very effective in the case
of turbines. Tests show that an increase of one inch of
vacuum will cause an increase of economy of from 3 to 10
per cent, depending upon the type of turbine and upon
other factors.

Experience has shown that reciprocating engines are
fully the equal of turbines in the high-pressure ranges,
in many cases they are even superior, but the turbine is far
superior in the low-pressure region and in cases where very
great ratios of expansion are to be used. Advantage has
been taken of the superior ability of the turbine to handle
low-pressure steam by constructing mixed plants, recipro-
cating engines being used for expanding down to the neigh-

THE STEAM TURBINE 249

borhood of atmospheric pressure and turbines expanding
the steam exhausted by these engines to the lowest vacuum
which can be maintained economically. This system has
been found particularly useful for increasing the capacity
of a reciprocating-engine plant. The capacity of such a
plant can often be almost doubled without any increase
in boiler capacity by simply inserting turbines into the
exhaust lines between the engines and the condensers, and
then arranging the pressures so that the turbines carry
the expansion from about 16 Ibs. absolute down to a vacuum
of from 28 to 29 ins. Turbines used in this way are called
low-pressure or exhaust-steam turbines.

Superheat is also very effective in bettering turbine
economy, every ten degrees of superheat generally causing
a saving of about 1 per cent in the weight of steam required
per horse-power.

The steam turbine is generally cheaper than the recipro-
cating engine of like capacity if the conditions of operation"
permit the use of the high rotative speed characteristic
of the turbine. It is therefore extensively used for direct
connection to blowers, centrifugal pumps and electrical
machinery. Most of the larger electric power stations
which have been installed within the past few years have
used turbines to drive the generators, and single units
direct connected to 20,000 K.W. generators are now numer-
ous. Units rated at 45,000 K.W. are also being installed
and units of still larger capacity have been designed.

PROBLEMS

1. A steam turbine produces one horse-power hour at its shaft
for every 30 Ibs. of steam supplied. The initial pressure is 200
Ibs. absolute and the steam is superheated 200° F. The turbine
exhausts against a back pressure of 14 Ibs. absolute.

Find the thermal efficiency on the assumption that heat
of liquid at exhaust temperature is not chargeable to the turbine.

2. Develop a complete expansion cycle for one pound of
material used under the conditions of Prob. 1 and find the energy

250 STEAM POWER

made available per cycle. From this value determine the number
of pounds of material theoretically required per horse-power hour
and compare with the value given in Prob. 1.

3. Find the additional quantity of energy which would theoret-
ically be made available per pound of steam in above problems if
the back pressure could be lowered to \ Ib. absolute.

4. Develop a complete expansion cycle from an initial pressure
of 225 Ibs. absolute with a superheat of 200° F. to a back pressure
of | Ib. absolute. Assume that this is to be divided up into six
parts, each making available the same quantity of energy. Find
the pressure drop for each part. Note that this is most easily
done with the help of the Mollier chart.

5. A steam turbine receives steam at a pressure of 225 Ibs.
per square inch absolute and with a superheat of 190° F. and
exhausts into a condenser in which a pressure of f Ib. per square
inch absolute is maintained. The turbine is direct connected to
an electric generator and produces a K.W.-hour on 12 Ibs. of steam.
If a K.W.-hour is equivalent to 3411 B.t.u., what is the thermal
efficiency of the combination?

6. Develop a complete expansion cycle for the conditions of
Prob. 5 and determine the pounds of steam which would be re-
quired theoretically to develop energy equivalent to 1 K.W.-hour.
Compare with the value given in Prob. 5.

7. Determine the velocity theoretically attainable by expanding
steam in one step from the initial to the final conditions of Prob.
5 above. What would be the value of the kinetic energy of such
a jet per pound of steam flowing?

8. Determine the shape of a nozzle required to discharge
1000 Ibs. of steam per hour, initial conditions being 100 Ibs. per
square inch absolute, and dry saturated steam; final pressure
being 2 Ibs. absolute.

9. Determine velocity and kinetic energy of jet in Prob. 8.

CHAPTER XIV
CONDENSERS AND RELATED APPARATUS

109. The Advantage of Condensing. The lowest pres-
sure which can be used in a steam-engine cylinder, that is
the exhaust pressure, is determined by the pressure prevail-
ing in the space into which the steam is exhausted. With a
given initial pressure the amount of work which can be ob-

FIG. 166.

tained theoretically from a given weight of steam increases
as the exhaust or back pressure decreases, as shown by the
areas of the two diagrams in Fig. 166, and experience has
shown that at least a part of this theoretical increase can be
obtained in real engines. It is therefore desirable to ex-
haust into a space in which the lowest possible pressure
exists when the work obtained per pound of steam is the
only consideration.

The most available space into which an engine can

251

252 STEAM POWER

exhaust is that surrounding the earth and already occupied
by the earth's atmosphere. The pressure in this space
is approximately equal to 14.7 Ibs. per square inch at sea
level and is due to the weight of the atmosphere. Since
the superincumbent column of atmosphere decreases in
depth as one moves upward, its weight also decreases and
atmospheric pressure therefore averages less than 14.7
Ibs. per square inch at high altitudes and has a greater
average value at points below sea level.

If it is desired to exhaust into a pressure lower than
atmospheric a means of maintaining such an abnormal
pressure within some sort of vessel must be devised. It
is the purpose of a condenser and its associated apparatus
to make available a space in which such a low pressure can
be maintained. Its method of operation will be considered
in later sections.

There is also another advantage which may be ob-
tained by the use of a condenser. It often happens that
the water available is not well adapted to use in boilers.
It may be salt water as in marine practice, or it may contain
a number of undesirable gases and solids in solution as often
occurs in stationary practice. Some types of condensing
apparatus are so arranged that the steam exhausted from the
engine is converted into liquid water without admixture
and can therefore be returned to the boiler as practically
pure water, thus largely eliminating the troubles that
would ensue from the use of poor feed water.

110. Measurement of Vacuums. Assume that some
non-volatile liquid, that is, a liquid that did not vaporize,
could be found and also that it contained no gases in solu-
tion. If a long tube were inserted in a vessel filled with
such a liquid and had its upper end connected with some
form of vacuum pump which could remove air from its
interior, as shown in Fig. 167, liquid would rise in the tube
as the air was removed. Removal of air would result in
lowering the pressure within the tube, but the constant

CONDENSERS AND RELATED APPARATUS 253

atmospheric pressure on the liquid surface outside the tube

would then force liquid up the tube to such a height that

the pressure pa of air in the tube

plus the pressure due to the column

of liquid of height h within the

tube just equaled the pressure due

to the atmosphere on the surface

of the liquid in the vessel. If the

pump could remove all of the air

from the tube, liquid would rise to

such a height that the pressure

exerted by it on a plane passing

through the lower surface just

equa-led that of the external atmos-
phere.

The same result could be at-
tained by using a tube closed at

one end, filling it with the liquid,

and then inverting so that the end

rested in the liquid as shown in

Fig. 168. If the tube were long

enough, the liquid would drop to some such
point as shown, under which conditions the
height of liquid would just balance atmos-
pheric pressure. This would only be true if
the liquid did not volatilize and did not contain
gases in solution; with these assumptions the
space above the liquid in the tube would con-
tain absolutely nothing. This space would be
said to be perfectly vacuous, or a perfect vacuum
would be said to exist in that part of the tube.
A device of this character is used to
measure the pressure of the atmosphere and
is known as a barometer. Mercury is used

as the liquid because its high density makes it possible

to use a short tube and because it may be considered

FIG. 167.

FIG. 168.

254

STEAM POWER

as non-volatile at ordinary temperatures. The average
atmospheric pressure at sea level, equal to 14.7 Ibs. per
square inch approximately, can support about 30 ins.
of mercury, so that this figure is generally taken as the
standard sea level barometer reading. An atmospheric
pressure of less than 14.7 Ibs. would be shown by a barom-
eter reading of less than 30 ins.; a greater atmospheric
pressure by more than 30 ins. Corresponding values of
atmospheric pressure and barometer reading are given
in Table VII. To this have also been added the altitudes
to which the different values would correspond if a pressure
of 14.7 Ibs. existed at sea level and there were no variations
of atmospheric pressure excepting those due to change of
elevation. Values of this type can only be roughly approx-
imate, because local barometric variations are constantly
occurring and the sea-level atmospheric pressure varies
both sides of 14.7 Ibs.

TABLE VII
ATMOSPHERIC PRESSURE, BAROMETER READING AND ALTITUDE

(Negative signs mean distance below sea level.)

Barometer,
Inches of Mercury.

Atmospheric Pressure,
Pounds per Square Inch.

Altitude,
Feet (Approximate).

25.00

12.27

4750

26.00

12.76

26.50

13.01

3250

27.00

13.25

27.50

13.49

2250

28.00

13.74

28.50

13.98

1300

29.00

14.23

29.25

14.35

29.50

14.47

450

29.75

14.60

30.00

14.72

Sea level

30.25

14.84

30.50

14.96

-450

30.75

15.09

31.00

15.21

-900

CONDENSERS AND BELATED APPARATUS 255

The exact value of standard atmospheric pressure
at sea level is taken at 29.921 ins. of mercury, which is
equal to 14.696 Ibs. per square inch and corresponds to
the 760 mm. of mercury, used by scientists as standard.

A tube with both ends open and arranged as shown
in Fig. 167 can be used to measure the degree of vacuum
existing in the space to which its upper end is connected,
and many vacuum gauges are constructed on this principle,
using mercury as the liquid. The extent to which the
pressure is lowered in the top of the tube is indicated by
the height to which the mercury column rises and this
height in inches is used as a measure of the vacuum. Thus
if a perfect vacuum were created and if the atmospheric
pressure were equal to T4.7 Ibs. the gauge would show
about 30 ins. of mercury above the level in the reservoir.
If the vacuum were less perfect the gauge would show a
shorter column.

It should be noted that the reading of the vacuum
gauge does not give the pressure existing in the vacuous
space, but gives the amount by which the pressure has been
reduced below that of the atmosphere, the difference
being expressed in inches of mercury. By subtracting this
reading from the existing atmospheric pressure expressed
in the same units, the absolute pressure in the partially
vacuous space (expressed in inches of mercury) is obtained.

It is obvious, therefore, that a vacuum-gauge reading
of say 28 ins. of mercury does not always mean the same
absolute pressure. With a barometer reading of 28 ins.
it would represent a perfect vacuum; with a barometer
reading of 30 ins. it would represent a partial vacuum, the
absolute pressure in the partially vacuous space being
equal to 2 ins. of mercury.

111. Conversion of Readings from Inches of Mercury
to Pounds per Square Inch. It is often necessary to con-
vert readings of pressure in inches of mercury into pounds
per square inch. This can be done with sufficient accuracy

256 STEAM POWER

under ordinary circumstances by multiplying the inches
of mercury by the constant 0.4908. Thus,

Barometer in inches X 0.4908 = atmospheric

pressure in pounds per square inch . . . (74)
and

(Barometer in inches — vacuum gauge in inches)
X 0.4908 = absolute pressure in partially
vacuous space in pounds per square inch. . (75)

ILLUSTRATIVE PROBLEM

A vacuum gauge constructed like that shown in Fig. 167
reads 27 ins. when the barometer reads 29.5 ins. What is the
absolute pressure in the partial vacuum above the mercury?

The absolute pressure is equal to 29.5-27=2.5 ins. of mer-
cury, which is equal to

2.5X0.4908 = 1.227 Ibs. per square inch.

112. Principle of the Condenser. A perfect vacuum
could be created in any closed vessel with impenetrable
walls if a pump could be devised which could remove all
material contained within that vessel. Or, any degree of
vacuum can be maintained in any partially closed vessel
by fitting it to a pump which can remove all material
flowing into the vessel as fast or faster than it enters, raise
the pressure of this material to atmospheric or. higher and
discharge it.

The latter principle is made use of in real condensers,
a pump of some form, or an equivalent, removing from the
condenser the material exhausted by the engine and in-
leakage from the atmosphere, and discharging it at atmos-
pheric pressure at a sufficiently rapid rate to maintain
the desired vacuum. If the condenser and connections
can be made leak proof, the pump or equivalent has to handle
only the material exhausted from the engine.

'A steam engine exhausts a mixture of steam, water
and gases, the gases being a mixture of those originally

CONDENSERS AND RELATED APPARATUS 257

dissolved in the boiler-feed water and air which leaks into
those parts of the system in which a partial vacuum is main-
tained. If the pump had to handle the same volume of
material as is exhausted by the engine, no gain of work
would result from condensing, because the pump would
have to do at least as much work in raising the pressure
of this material to atmospheric and discharging it as could
be obtained by allowing it to expand in the engine.

Steam, however, occupies a much larger volume than
water at the same temperature and pressure. Thus steam
at 212° F. occupies a volume of about 26.79 cu.ft. per
pound, but water at the same temperature and pressure
occupies a volume of only about 0.0167 cu.ft. per pound;
at a temperature of 120° F. which is often used in condensers,
the specific volume of steam is about 203 and that of water
only 0.0162. Therefore, if the steam is condensed after
exhaust from the engine and before entering the pump to
be discharged to atmosphere, the pump work is greatly
reduced. The volume of the condensate is almost negli-
gible in comparison with the volume of steam exhausted,
and the work of pumping it is almost negligible in compari-
son with the work it made available in the engine.

Gases contained in the exhaust steam cannot be lique-
fied and must be pumped as gases. The work required
to pump them can, however, be reduced by lowering their
temperature as far as possible.

The condenser equipment may be regarded as con-
sisting of a combination of a partially closed vessel and
some form of pump. The vessel is so constructed that a
low temperature can be maintained within it and that
large quantities of heat can be removed from it for the
purpose of condensing the exhaust steam and of cooling the
contained gases. This is generally done by using large
quantities of cool water.

The absolute pressure within the condenser is made
up of two parts. The two parts are, (a) that due to the

258 STEAM POWER

water vapor, since the space over the condensed water will
always be filled with saturated steam at the same tempera-
ture (approximately) as that of the water, and (6) that due
to any gases present.

The pressure of the saturated steam (water vapor)
can be found from the steam tables opposite the temperature
existing in the condenser and it is the pressure which would
exist in the condenser of an ideal system in which no gases
were mixed with the working substance. The pressure
of the gases can be found by subtracting from the total
measured pressure in the condenser the pressure exerted
by the water vapor as shown in the steam tables. The
pressures exerted by the water vapor and gases are spoken
of as partial pressures, since their sum makes up the total
pressure within the condenser.

The presence of gases causes a two-fold loss. First,
it increases the pressure against which the engine has to
exhaust, thus raising the back-pressure line on the diagram
and decreasing the work area. Second, it increases the
work which must be done by the pump which otherwise
would only pump the condensate and such saturated water
vapor as accompanied it.

113. Types of Condensers. The condensers actually
used in steam plants can be roughly divided into two types,
as

(a) Contact condensers and

(b) Non-contact condensers.

In the first type the water which is used for condensing
and cooling it intimately mixed with the exhaust from the
engine within the condensing vessel, and the resultant
mixture of condensing water, condensate and gases is drawn
out of this vessel and discharged to atmosphere by the
pump.

In the second type condensing water flows on one side
of metal surfaces of some sort and the exhaust is led over
the other side, the heat being transmitted through the

CONDENSERS AND RELATED APPARATUS 259

Condons

metal. In condensers of this type the condensate and
gases are not mixed with the condensing water and the
condensate can therefore be returned to the boiler as feed
water with the advantages already mentioned.

114. The Jet Condenser. One of the earliest forms of
contact condensers which is still very widely used for
moderate vacuums is commonly
known as the jet condenser. The
principle of operation of the jet
condenser is shown in Fig. 169.
Water, under pressure, entering
as indicated, is broken up into water in
fine streams or jets and sprayed
into the exhaust coming from
the engine. The resultant mix-
ture flows downward into the
neck of the condensing vessel or
" condenser head " and is re-
moved by some form of pump.
This pump handles gases, vapors
and water and is known as
a vacuum pump, a wet-vacuum
pump, or a wet-air pump, the
term wet signifying that it han-
dles the water as well as the

^Mixture to Puinp

FIG. 169. — Jet Condenser.

gases.

The pressure within such a
condenser head would be theo-
retically equal to that corresponding to the temperature
of the resultant mixture if no gases were present. In
practice the pressure of the water vapor would roughly
correspond to the average temperature near the top of the
vessel and there would be a partial pressure due to gas
as well. This gas would consist of that brought over by
the engine exhaust plus that released from the condensing
water under the low pressure within the condenser.

260

STEAM POWER

Details of a complete jet condenser and. of the method
of connecting it to an engine are given in Fig. 170. The
atmospheric relief valve is installed in all condensing sys-
tems and is arranged to open automatically to atmosphere
if the pressure within the system rises to atmospheric, that
is, if the " vacuum is lost."

FIG. 170. — Jet Condenser and Method of Connecting to Engine.

With the jet condenser the pressure might start to
rise because of slow action or even stoppage of the pump.
As the condenser head filled up the rising water would
ultimately entirely cover the jet and condensation would
then practically cease. In the arrangement shown in
Fig. 170 there is an additional safety device which breaks
the vacuum in the exhaust system if the water in the head

CONDENSEES A^'D RELATED APPARATUS 261

rises above a certain height, thus preventing the external
atmospheric pressure from forcing this water back along
the exhaust pipe and into the cylinder, an event which
would probably result in a wrecked engine.

The jet condenser here described is known as a parallel-
flow type, because everything within the condensing vessel
flows in the same direction. The gases and vapors handled
by the pump theoretically have the same temperature
as that of the mixture with which they flow out at the
bottom of the condenser head. The temperature of this
mixture therefore determines the temperature of the gases
and vapors pumped.

There are numerous forms of contact condensers which
more or less closely resemble the types of jet condenser
just described. They are occasionally all classed as jet
condensers, but more often are given distinguishing names.

One very common form of contact condenser is generally
known as a barometric condenser. It consists essentially
of a condenser head, similar to that used with the jet con-
denser already described, and a tail pipe or barometric
pipe which partly or wholly takes the place of the wet-
vacuum pump by removing part or all of the mixture formed
within the condenser. One model of such a condenser is
shown in Figs. 171 and 172.

The exhaust from the engine enters the head through
the large pipe shown and divides into two parts, one part
passing down through the center of the head and the re-
mainder flowing downward in the annular space A. The
condensing or injection water enters as shown and is divided
by the spraying cone and injected into the engine exhaust,
which enters the central tube of the condenser. The
mixture thus formed flows downward and finally meets the
discharge from the lower end of the annular space A, which
is then condensed. The mixture of injection and con-
densing water together with such gases as have been en-
trapped, then flows downward into the tail pipe, which is

262

STEAM POWER

over 34 ft. in length and which dips into the " hot well
at its lower end. As atmospheric pressure can only sup

/To Vacuum or
Dry Air Pump

Air Cooler

Exhaust

Water for Cooling
/ "Air"

FIG. 171. — Barometric Condenser.

port a column of water about 34 ft. high, the tail pipe forms
an automatic wet-vacuum pump, water flowing from it as
rapidly as it accumulates within it.

CONDENSERS AND RELATED APPARATUS 263

264

STEAM POWER

Atmospheric Relief Valve

Experience has shown that the maintenance of a high
vacuum with this type of condenser depends upon the ex-
tent to which gases are removed from the condenser head.
These gases are generally called air, as the greater part of
them is air. In the type illustrated such " air " as is
not trapped by the descending mixture rises through the

hollow spraying cone, then
through the air cooler and
flows out through the pipe
indicated to the vacuum or
dry-air pump. The air in
rising through the center
of the spraying cone is
cooled by the water flowing
around it, and it is further
cooled by coming into con-
tact with water as it works
its way through the air
cooler. This results not
only in lowering its tem-
perature, but also in caus-
ing the condensation of a
great deal of the water
vapor accompanying it.
This condensed vapor col-

FIG. 173.— Baragwanath Barometric |ects in the SPace surround-

Condenser. mg the air cooler and

flows down into the head

through the drain shown. The vacuum pump, therefore,
handles cool gases containing little water vapor and prac-
tically no liquid water. It is sometimes called a dry-air
pump or dry- vacuum pump for this reason.

The entrainer shown in the exhaust system in Fig. 172
is so shaped that water collecting in the exhaust piping
and flowing into the entrainer is picked up by the exhaust
steam and carried into the condenser.

CONDENSERS AND RELATED APPARATUS 265

The flow of steam and injection water in this condenser
is parallel, but the material on its way to the dry-vacuum
pump flows upward and the cooling water flows downward
so that counter-current flow is used in this part of the appa-

BARAGWANATH

CONDENSER

ORDINARY

SETTING

£.

Engine

I1
R

SYPHONING ITS

WATER FROM
TANK OR FLUME

FIG. 174.

ratus. This has the advantage of bringing the air leaving
the condenser into contact with the cooling water just as
it enters and therefore when it has its lowest temperature.

A somewhat similar condenser, arranged so that it
requires no pump, is shown in Figs. 173 and 174 (a) and (6).

266

STEAM POWER

Exhaust and injection water mix as shown, the quantity
of injection water being regulated by the hand wheel on
top of the condenser. The mixture flows downward through
the narrow neck and the velocity attained in this part of the
tail pipe is so high that all air and similar gases are swept
along with the current.

For starting, the
valve V in Fig. 174 (b)
is opened, allowing water
to flow into the lower
part of the tail pipe.
This creates a partial
vacuum, and atmos-
pheric pressure then
forces water up the in-
jection pipe and into
the condenser head. The
valve V is then closed
and the condenser con-
tinues to siphon its own
water. Because of this
action this type is often
called a siphon conden-
ser. By supplying a
circulating pump as in-
dicated in Fig. 174 (a)
it can be converted into
a barometric condenser

similar to the type already discussed except for the fact
that it requires no air pump.

The barometric or tail pipe of any barometric condenser
can be replaced by any kind of a pump, and centrifugal
pumps are often used for this purpose. When large quanti-
ties of gas are to be handled, as when a dry-air pump
is not used, the centrifugal pump must be specially
designed.

FIG. 175. — Westinghouse-Leblanc Air
Pump.

CONDENSERS AND RELATED APPARATUS 267

A recently developed type of condenser in which the
barometric tube is replaced by a centrifugal pump and in
which a separate air pump of a rotary type is used is illus-
trated in Figs. 175, 176 and 177. It consists essentially of
the condensing head and well, combined with a centrifugal
tail pump and a rotary air or vacuum pump as indicated

Submerged not
less thau 3 ft.

FIG. 176. — Westinghouse-Leblanc Condenser.

in Fig. 177. Injection water entering through nozzles in
the head meets the exhaust, and the resultant mixture
flows down into the well through the large nozzle shown.
The liquid is continuously removed from the bottom of this
well by the centrifugal tail pump and discharged to the hot
well. The air and associated vapors are drawn down the
air pipe and discharged by means of the device shown in

268

STEAM POWER

Fig. 175. Water enters the central part of this pump as
indicated in Fig. 176 and is discharged through the station-
ary nozzles N and the moving vanes V shown in Fig. 175.
The water is thus caused to form a series of " pistons "
which move rapidly downward in the discharge nozzle N'
and which trap small plugs or lamina of " air " between

them and thus discharge
the " air " to the atmos-
phere. The connection
marked P is used for prim-
ing at starting when neces-
sary.

In small units the cen-
trifugal tail pump may be
omitted and the design so
remodeled that all the injec-
tion water passes through
the rotary air pump which
discharges the entire mix-
ture from the condenser
just as it discharges the air
and associated vapors in
the larger sizes.

115. Non-contact Con-
densers. The type called
the surface condenser is
the best-known example of
non-contact condenser. It

consists essentially of a large cylindrical or rectangular
vessel into which the exhaust is discharged and through
which pass numerous bronze or alloy tubes which carry the
condensing water, and the surfaces of which act as the
condensing and cooling surface.

One form of surface condenser mounted above the
pumps which serve it is shown in Fig. 178. The exhaust
enters at the top of the rectangular shell and works its

.Water to
Air Pump

FIG. 177. — Westinghouse-Leblanc
Condenser.

CONDENSERS AND RELATED APPARATUS 269

270 STEAM POWER

way down over the water-cooled tubes. The condensate,
mixed with gases and vapors, is drawn from the bottom of
the shell by the wet-vacuum pump and discharged to the
hot well.

The condensing water is forced through the tubes
of the condenser by means of the reciprocating circulating
pump, entering the lower tubes at the right-hand end in
the figure, making two " passes " through the condenser and
leaving at the top. Because of the path of the water a
condenser of this type is sometimes called a two-pass or
double-flow condenser.

With the arrangement illustrated, the steam which
condenses upon the upper tubes falls as a rain from tube
to tube until it finally settles at the bottom and is drawn
off. The outer surfaces of the lower tubes are therefore
practically covered with water and this has two disad-
vantages. First, these tubes carry the coolest circulating
water and they therefore cool the condensate coming in
contact with them while the water flowing through them
is unnecessarily heated. Cooling of the condensate means
a lower hot-well temperature than would otherwise be
obtained, but if the condensate is to be used for boiler
feed, the temperature of water in the hot well should be
maintained as high as possible, since this water will eventually
have to be heated to boiler temperature with a correspond-
ing expenditure of heat. Second, tubes which are being
used to heat water covering them are of little use as condens-
ing surface, and hence a large part of the surface in such a
condenser is comparatively inactive.

The ideal arrangement would carry away the liquid
condensate as fast as formed, leaving the tubes first entered
by the condensing water to act as the final condensing
and cooling surfaces, thus bringing gases and non-condens-
ible vapors into contact with the coolest surfaces j ust before
entering the vacuum pump. Numerous designs which
approximate this ideal have been developed recently and

CONDENSERS AND RELATED APPARATUS 271

they give better results than do the earlier and simpler
forms. The improvement is shown by the values of con-
densing surface per developed horse-power of engine. In
early designs it was customary to supply 2| sq.ft. of tube
surface or more per horse-power. Some of the most recent
installations are giving better vacuums with only 1 sq.ft.
per horse-power.

One of the newer models passes the condensate through
a set of tubes so located that the engine exhaust strikes
them before impinging on any tubes carrying condensing
water. This results in a partial condensation of the exhaust
and raises the temperature of the condensate within the
tubes to very near that of the exhaust, thus heating the
boiler feed to a temperature practically corresponding
to the exhaust temperature of the engine.

Surface condensers are commonly operated with a
vacuum of from 24 to 26 ins. of mercury when used with
reciprocating engines and with a vacuum of 28 to 29 ins.
when receiving the exhaust of steam turbines. When
operated at the lower vacuums wet-vacuum pumps are gen-
erally used, but the best types of dry-air pumps must be
installed in combination with well-designed condensers
when the higher vacuums are sought.

116. Water Required by Contact Condensers. The
weight of circulating water required varies with the type
of condenser and with the conditions of operation, such as
initial temperature of water, vacuum desired, etc. It can be
determined approximately by calculation and the values
thus found must then be increased by such factors as
experience has shown to be necessary.

In contact condensers the water and the condensate
are discharged as a mixture and therefore have the same
average discharge temperature.

Let ti = initial temperature of injection water in F.°;
t2 = temperature at which mixture is discharged in

272 STEAM POWER

X = total heat above 32° F. of steam as exhausted;
W = pounds of injection water per pound of exhaust
steam.

Assuming the exhaust steam to be dry saturated, each
pound of steam in condensing to water at a temperature
of fa degrees must give up an amount of heat equal to X
minus the heat of the liquid at t°2 or roughly X— (fa — 32)
B.t.u. This same quantity must be absorbed by the in-
jection water, while its temperature rises from t\ to fe
degrees. Each pound of water can then absorb approxi-
mately (fa — ti) B.t.u. and the pounds of injection water
per pound of steam will be

(78)

The value of fe would be that corresponding to the
absolute pressure in the condenser if it were not for the
air and similar gases which exert some pressure. It is
generally 10 or more degrees F. below the temperature
corresponding to the vacuum. Values of t% in the neigh-
borhood of 110° to 125° F. are customary with recipro-
cating engines and values as low as 80° are used with high
vacuums in connection with steam turbines.

The weight of water used per pound of steam as given
by Eq. (78) will vary between about 15 for very low initial
and moderate discharge temperature to about 50 with
average initial and moderate discharge temperature. Ex-
perience shows that it is necessary to add 10 per cent or
more to the values of W obtained from equation (78) to
obtain the weight of water which will probably be used.

ILLUSTRATIVE PROBLEM

Find the quantity of water theoretically required per pound
of steam condensed in a contact condenser in which a vacuum
of 25.5 ins. of mercurv is maintained when the barometer reads

CONDENSERS AND RELATED APPARATUS 273 .

29.5 ins. of mercury. The initial temperature of the water is
60° F.

The absolute pressure in the condenser is 29.5-25.5=4.0 ins.
of mercury and the steam tables give for this pressure, X = 11 15.0
and ti = 126. Substituting in Eq. (78) gives

w 1115.0-126+32

126 -60~ = approximately.

117. Weight of Water Required by Non-contact Con-
densers. In the case of non-contact condensers there is
no definite relation between the discharge temperature
of the cooling water and that of the condensate. Experi-
ence shows that the discharge temperature of the circulating
water is generally from 10 to 20 degrees lower than the
temperature corresponding to the vacuum.

The^temperature of the condensate (hot-well tempera-
ture) is generally 15 or more degrees below that correspond-
ing to the vacuum, but good design makes the hot-well
temperature very closely approximate that corresponding
to the vacuum.

Assuming

ti = initial temperature of injection water in F.°;

£2 = final temperature of injection water in F.°;

tc = temperature at which condensate is discharged, i.e.,

hot- well temperature, in F.°;
X = total heat above 32° F. of steam as exhausted,

and

W = pounds of injection water per pound of exhaust
steam.

The weight of water which must be circulated per pound
of steam can be found as in the case of the contact con-
denser. It is given by

(79)

12 — h

274 STEAM POWER

Values in the neighborhood of 25 Ibs. of water per pound
of steam are common with low vacuums and 50 or more
pounds are often used with vacuums over- 28 ins. of
mercury.

ILLUSTRATIVE PROBLEM

A surface condenser receives circulating water at a temper-
ature of 65° F. and discharges it at a temperature of 80° F. It
maintains a vacuum of 28.0 ins. with the barometer at 29.5, and
the temperature of the condensate discharged to the hot well is
equal to 85° F. Find the quantity of circulating water theoretically
required.

This vacuum corresponds to an absolute pressure of
29.5—28.0 = 1.5 ins. of mercmy. Assuming this all due to steam
(neglecting presence of air) the value of X may be taken from the
steam table as 1100.1 B.t.u. Substitution' in Eq. (79) then gives

1100.1-85+32

" = ~~ ~^ — 7^- =69.9 approximately.

80— bo

118. Relative Advantages of Contact and Surface Con-
densers. The contact types are as a rule much cheaper
than the surface condensers, and they are less subject to
serious depreciation, the tubes of surface condensers often
corroding seriously in very short intervals of time. On
the other hand, the injection of the cooling water into
the condensing space in contact types results in the intro-
duction of large quantities of dissolved gases, and much
of this material is liberated under the reduced pressure,
thus tending to increase the condenser pressure, that is,
decrease the vacuum. Where pumps are used to carry
away the mixture with contact condensers, these pumps
have to handle a much larger quantity of water than the
corresponding pump in a surface condenser, and the work
of pumping this water out of the vacuum into the atmos-
phere combined with the additional work required of
the pump which handles the " air " may partly balance
the advantage of lower first cost of the contact type.

CONDENSERS AND RELATED APPARATUS 275

A surface condenser must always be installed where
it is desirable to use the condensate as boiler feed, and
it is generally used when very high vacuums (low absolute
pressures) are to be maintained. The surface condenser
is at a serious disadvantage, however, when required to
handle the exhaust of reciprocating engines. The exhaust
from such engines always contains large quantities of
lubricating oil carried out of the cylinder, and unless this
material is separated before the exhaust enters the con-
denser it is deposited on the outer surfaces of the tubes
and decreases the conductivity of those surfaces. Such
oil can be eliminated to a great extent before the exhaust
enters the condenser by means of oil separators, which are
generally made up of a series of baffles upon which the
steam impinges and upon which the oil is caught.

119. Cooling Towers. The large quantity of circula-
ting water required by condensing plants is often an item
of great economic importance. When such plants are
located near a river or near tide water, the circulating
water can generally be procured for the cost of pumping.
When they are located in the middle of cities or in regions
where water is scarce, the cost of water may be excessive
or it may even be impossible to obtain a continuous supply
equal to the demand of the condensers.

In such cases the condensing water is often circulated
continuously, being cooled after each passage through the
condensers. This cooling is generally done by exposure
of a large surface to the air. The resultant evaporation
of some of the water with the absorption of its latent heat
of vaporization cools the remainder so that it can be used
again. This sort of cooling may be effected by running
the water into a shallow pond of large area, or by spraying
it into the air over a small pond or reservoir or by passing
it through a cooling tower.

Cooling towers are large wood or metal towers generally
filled with some form of baffling devices. The hot water

276 STEAM POWER

is introduced at the top and spread into thin sheets or
divided up into drops as it descends. Air enters at the
bottom and flows upward, cooling the water by contact
and by the partial evaporation which results. The cir-
culation of air may be natural, i.e., due to the difference
of temperature between the air inside and out, in which
case a stack is fitted to the top of the tower; or the cir-
culation may be forced by fans located in the base of the
tower. In the latter case the apparatus is called a forced
draught cooling tower.

CHAPTER XV
COMBUSTION

120. Definitions. Certain substances are known to
chemists as compounds, because they can be separated by
chemical processes into simpler substances. Thus water
and many of the most familiar materials known to man
are compounds which can be separated into two or more
simpler materials.

Those substances which cannot be further broken up by
the processes used in separating compounds are called
elements; they are regarded as elemental, as the stones
of which the compounds of nature are built up. About
seventy-five of these elements are now known, but many
of them are comparatively rare. Pure metals are all
elements; the oxygen and nitrogen which are mixed to form
the greater part of the atmosphere are elements; carbon,
which forms the greater part of most fuels, is an element.

In many cases the combination of elements to form
compounds is accompanied by the liberation of heat, and
some of these combinations are used by the engineer for the
purpose of obtaining heat in large quantities. When the
elements which occur in fuels, such as coal, wood and
petroleum, combine with oxygen, the process is spoken of
as combustion. The quantity of heat liberated when a
pound of any material combines with oxygen (burns) is
called the heat value or calorific value of that material.

Fuels contain a great number of elements, but only
three of these ordinarily take part in combustion and are
therefore spoken of as combustibles. They are carbon,
hydrogen and sulphur. The sulphur content is generally

277

278 STEAM POWER

very small, and the carbon and hydrogen are therefore the
most important constituents.

The combustion of each of these elements will be con-
sidered in detail in the following sections, but before this
can be done two other idea's must be developed.

The smallest particle of an element which can be
conceived of as entering into combination to form a com-
pound is known as an atom of that element. It has been
found that the atoms of each element have an invariable
and characteristic mass. The lightest atom is that of
hydrogen, and its weight is considered unity. The atom
of carbon is twelve times as heavy as that of hydrogen and
carbon is therefore said to have an atomic weight equal to
twelve. Similarly the atomic weight of nitrogen is four-
teen and that of oxygen is sixteen.

The smallest particle which can be formed by the com-
bination of atoms is known as a molecule. Like or unlike
atoms may combine to form molecules. Thus two hydro-
gen atoms combine to form a molecule of hydrogen, and
hydrogen gas as it ordinarily exists may be pictured as
made up of a collection of such molecules. Similarly,
gaseous oxygen and gaseous nitrogen may be pictured as
collections of molecules which are made up of two like
atoms.

When unlike atoms combine to form a molecule, they
form a molecule of a compound. Obviously a molecule of
any compound is the smallest particle of that compound
which can exist.

For convenience, the different elements are represented
by abbreviations; thus oxygen is represented by O, nitro-
gen by N, hydrogen by H, carbon by C and sulphur by S.
When these abbreviations are written in chemical equa-
tions, such as will be given later, they stand for an atom
of the substance. Hence O in a chemical equation would
mean one atom of oxygen. The symbol O2 is used to mean
two atoms of oxygen in combination, hence, one molecule

COMBUSTION 279

of oxygen. The symbol 202 means two groups of two
oxygen atoms in combination, hence two molecules of oxygen.
The simplicity and elegance of this system will become
apparent as the chemical equations which follow are de-
veloped and explained.

121. Combustion of Carbon. Carbon can unite with
oxygen or burn to form two different compounds — carbon
monoxide (CO) and carbon dioxide (C02). The monoxide
is formed by the combination of one atom of oxygen with one
atom of carbon; the dioxide, by the combination of two
atoms of oxygen with one of carbon. The dioxide, therefore,
contains twice as much oxygen as does the monoxide.

Carbon burned to carbon monoxide has not combined
with the largest possible quantity of oxygen, and combus-
tion is therefore said to be incomplete in such cases. When,
however, carbon dioxide is formed, the carbon has combined
with as much oxygen as possible and combustion is said to
be complete.

It will be shown later that much more heat is liberated
when the dioxide is formed than when carbon burns to the
monoxide. Hence, when liberation of heat is the object of
combustion, the process should be so conducted as to result
in the formation of the maximum quantity of dioxide and
the minimum amount of monoxide.

122. Combustion to CO. The combustion of carbon and
oxygen to form the monoxide can be represented by the
equation

C+O = CO, ...... (80)

or by the equation

2C+O2 = 2CO ....... (81)

The former is the simpler and will be considered first, but
the latter is the more perfect and indicates more to the
trained eye than does the simpler form.

The simple equation states that one atom of carbon
combined with one atom of oxygen to form one molecule

280 STEAM POWER

of carbon monoxide. It can, however, be so interpreted
as to show much more than this. The carbon atom is twelve
times as heavy as the hydrogen atom, while the oxygen atom
is sixteen times as heavy as that of hydrogen. The equation

C + O=.CO,

therefore, shows that an atom, which is twelve times heavier
than the hydrogen atom, unites with one which is sixteen
times heavier than the hydrogen atom to form a molecule
which is 28( = 12 + 16) times heavier than the hydrogen
atom.

In other words, the weights of carbon and oxygen

12 3 1

combining are in the ratio of -— = T = TT. If a sufficient

16 4 1J

number of carbon atoms to weigh one pound be used, a
quantity of oxygen weighing 1| Ibs. will be required to
combine with them to form carbon monoxide. The re-
sultant carbon monoxide will contain the pound of carbon
and the 1J Ibs. of oxygen and will therefore weigh 2J Ibs.

The same weight relations would hold irrespective of
the weight of carbon used, and the simpler equation may
therefore be put

1 weight of C+1J weights of O = 2J weights of CO. (82)

ILLUSTRATIVE PROBLEM

To illustrate the use of this equation, assume that 9 Ibs. of
carbon are burned to carbon, monoxide and that it is desired to
find the weight of oxygen used, and the weight of the product.
The weight of oxygen used must be 1^ times the weight of carbon,
that is, 1^X9 = 12 Ibs. The weight of the product must be 2£
times the weight of the carbon, that is 2|X9 =21 Ibs.; or, it must
be the weight of the carbon burned plus the weight of the ox}^gen
used, that is, 9+12 =21 Ibs.

In general, the oxygen used for combustion is obtained
from the atmosphere, which may be considered as a median-

COMBUSTION 281

ical mixture of oxygen and nitrogen in unvarying porportions.
These proportions are roughly, 0.23 of oxygen to 0.77 of
nitrogen by weightier 0.21 of oxygen to 0.79 of nitrogen
by volume, as shown in Table VIII. The weight of
air which contains one pound of oxygen is therefore

0.23+0.77 , .

— = 4.35 Ibs., and this weight of air contains
U..£o

4.35-1=3.35 Ibs. of nitrogen.

In the problem previously considered it was found
that 12 Ibs. of oxygen would be required to burn 9 Ibs.
of carbon to CO. The total weight of air required to
obtain this oxygen will be 12X4.35 = 52.2 Ibs. and it will
contain 52.2-12 = 40.2 Ibs. of nitrogen.

By simple arithmetical calculations of the type just
given all the weight relations involved in the combustion
of C to CO can be determined. The volume of air required
in any given case can be found by multiplying the weight
of air by the specific volume as given in Table VIII.
Thus, in the illustrative problem already considered, it
was found that 52.2 Ibs. of air would be required to burn
9 Ibs. of C to CO. The volume of this air at 62° F. would
be 52.2X13.14 = 685.9 cu.ft.

It is found that a quantity of heat equal to about
4500 B.t.u. is liberated per pound of carbon burned to CO;
that is the calorific value of C burned to CO is 4500 B.t.u.

Returning now to Eq. (81), which was said to be more
useful than the simpler form given as Eq. (80), it will be
necessary to consider a rather simple law of gases. It
has been shown experimentally that equal volumes of all
gases contain the same number of molecules when at the same
temperature and pressure. This statement is known as
Avogadro's Law.. It has also been shown that the mole-
cules of gaseous oxygen contain two atoms.

The equation in question,

282

STEAM POWER

can therefore be read, two atoms of carbon combine with
one molecule of oxygen to form two molecules of carbon mon-
oxide. But, if every molecule of O yields two molecules
of CO it follows from Avagadro's law that the CO formed
will occupy twice the volume of the oxygen used if measured
at the same temperature and pressure. If the equation be
imagined as containing a numeral 1 before the O2, it
will be obvious that the coefficients of the terms represent-
ing gas molcules give volume relations directly. This equa-
tion therefore gives both volume and weight relations.

TABLE VIII
PROPERTIES OF AIR

Considering it to Consist only of nitrogen and oxygen.

Relative Proportions.

Ratio of N to O.

Ratio of Air to O.

Exact.

Approx.

Exact.

Approx.

Exact,

Approx.

By weight . .
By volume. .

/0.766N

10.234O

/0.791N
10.209O

0
0.

77 N
23 O

79 N
21 O

3.27

3.78

3.35
3.76

4
4

27
76

4.35
4.76

Spec. wt. at Atmos. Press.
(Lbs. per Cu.ft.)

Spec. Vol. at Atmos. Press.
(Cu.ft. per Lb.)

At 32° F.

At 62° F.

At 32° F.

At 62° F.

0.08072

0.07609

12.39

13.14

Weight of air containing one pound of oxygen is approximately
4.35 Ibs.

123. Combustion to CO2. The combination of carbon
and oxygen to form the dioxide is represented by the equa-
tion

(83)

COMBUSTION 283

which shows that one atom of carbon (twelve times heavier
than hydrogen) combines with two atoms of oxygen (each
sixteen times heavier than hydrogen) to form a molecule
of CO2, which is forty-four times heavier than an atom of
hydrogen. Therefore the weight of carbon and oxygen

12 31

combining are as ^TTTTH = o = o2> so tnat 2t H>s. °f oxygen
ZXIO o Z;j

are required to burn a pound of carbon to carbon dioxide.
Writing this in the form of an equation, gives

1 weight of C-f 2| weights of O = 3f weights of CO2. . (84)

The weight of air required can readily be found by
multiplying the required oxygen by the number 4.35,
previously shown to be the number of pounds of air con-
taining one pound of oxygen. Thus, the required air is
2f X4. 35 = 11. 57 pounds per pound of C burned to C02.
This number is commonly rounded out to 12 in engineering
literature.

The equation given shows volume relations directly.
It is evident, therefore, that one molecule of O yields one
molecule of CCb, and hence that the volume of the product
is exactly equal to the volume of the oxygen used in forming
it if measured at the same temperature and pressure.
This is a very important relation, and is often made use
of in engineering calculations.

Experiment shows that when carbon burns to the
dioxide about 14,600 B.t.u. are liberated per pound of
carbon burned, that is, the calorific value of C burned to CO2
in 14,600.

124. Combustion of CO to CC>2. Since carbon which
has burned to carbon monoxide has not combined with the
greatest possible quantity of oxygen, the monoxide can
take up more oxygen by burning to the dioxide. This
process is represented by the formula

2CO+02 = 2CO2, (85)

284 STEAM POWER

which shows that two molecules of monoxide combine with
one molecule of oxygen to form two molecules of the dioxide.
The volume of CO2 formed is therefore equal to that of
the CO burned.

So far as the ultimate result is concerned, at makes no
difference whether carbon is burned directly to CO2 or is
first burned to CO and then the CO is burned to CO2
The total oxygen used per pound of carbon burned to CO2
and the total heat liberated per pound of carbon burned
to C02 are the same in both cases.

Thus, for the oxygen, one pound of C burned to CO2
requires 2f Ibs. of oxygen; but one pound of C burned
to CO requires 1J Ibs. of oxygen, and 1J Ibs. more will be
required when this CO is burned to C02. The result
is therefore the same.

For heat liberated, one pound of C burned to C02
liberates about 14,600 B.t.u.; but one pound of C burned
to CO liberates about 4500 B.t.u. and 10,100 B.t.u. are
liberated when this CO is burned to C02. Since the sum
of 4500 and 10,100 is equal to 14,600 the result is again
the same.

Data on the combustion of C to CO and CO2 and the
combustion of CO to CO2 'are collected in convenient
form in Table IX.

125, Conditions Determining Formation of CO and CO2.
Excluding certain complicated considerations which are
not of great importance in steam-power engineering, it may
be said that when carbon is being burned at a certain rate
(pounds per unit of time) the amount of oxygen brought
into contact with the carbon determines whether the caibon
burns to CO or to CO2. If enough or more than enough
oxygen to burn the carbon to CO2 is brought into contact,
that oxide will be formed. If there is not enough to burn
all the carbon to the dioxide, both oxides are formed in cer-
tain proportions, which can be calculated.

Since combustion to CO yields only 4500 B.t.u. per

COMBUSTION

285

pound of C and combustion to €62 yields 14,600 B.t.u.
per pound of C, the importance of supplying sufficient
oxygen to burn all carbon to the dioxide in cases where
the liberation of the maximum quantity of heat is desirable
is obvious. In actual practice the oxygen is furnished
by supplying air and it is found necessary in most cases
to supply more than the amount of air theoretically re-
quired in order to insure burning all, or even nearly all,
of the carbon to the dioxide. This comes from the great
difficulty met in obtaining contact between the oxygen of the
air and the carbon which is to be burned, that is, in bringing
all the oxygen of the air into intimate contact with the
fuel being burned in real apparatus.

TABLE IX
COMBUSTION DATA FOR CARBON

(Per pound of carbon )

Oxygen Required.

Nitrogen Accompanying
Oxygen.

T>«, ^a Cu.ft. at 62° F.
Pounds. and 14 7 j bs

Pounds.

Cu.ft. at 62° F.
and 14.7 Lbs.

1.333 16.0

4.46

60.1

CO2 from C.

2 667 32 0

8 92

120 2

CO, from CO

1.333 16.0

4.46

60.1

Air Required.

Quantity of Product
(N not included).

Product.

Cu.ft. at

Cu.ft at

ated.

Pounds.

62° F. and

Pounds.

62° F. and

14.7 Lbs.

5 79

76.1

2.33

32.0

4,500

CO2 from C

11.58

152.2

3.67

32.0

14,600

10,100 per Ib.

CO2fromCO. ...

5.79

76.1

3.67

32.04

of C in CO
4,300 per Ib.

of CO

286 STEAM POWER

The air in excess of that theoretically required to burn
all the carbon completely is spoken of as excess air. In
the form of an equation, this statement is equivalent to

Air supplied — air theoretically required = excess air. (86)

It is customary to express the quantity of excess air in
terms of a numerical factor known as the excess coefficient.
This coefficient is defined as the number by which the quantity
of air theoretically required must be multiplied to give the
quantity of air actually used. In the form of an equation
this gives

Excess coefficient X air theoretically required

= air actually used. . (87)

ILLUSTRATIVE PROBLEM

Taking data from the illustrative problem previously considered,
assume that 9 Ibs. of carbon are burned in air to C02. Each pound
theoretically requires 11.57 Ibs. of air, so that the theoretical
air-supply for this case would be 9X11.57=104.13 Ibs. If in a
real case 150 Ibs. of air are supplied, the excess coefficient is equal
to 150^ 104. 13 =1.44.

126. Flue Gases from Combustion of Carbon. The

gases resulting from the combustion of fuels are known
in engineering as the products of combustion or flue gases,
because they are the gases passing through the flues or
passages leading from furnaces in which fuel is burned and
to the stacks which serve to carry off the gases.

It has already been shown that the CO2 formed by the
combustion of carbon has the same volume as the oxygen
which is used in forming it. Therefore, if the air supplied
in a given case just equaled that theoretically required
for combustion to CO2 and if all of the oxygen were used,
the CO2 formed would merely replace the oxygen in the
air. The theoretical proportions of the flue gas would
then be 0.21 of CO2 and 0.79 of N by volume,

COMBUSTION

287

If real flue gases obtained by burning carbon in air are
found to contain less than 21 per cent of CO2, the combustion
has evidently not yielded theoretically perfect flue gases.
The trouble may be due to an excess or to a deficiency of air.
If there is an excess of air there will be oxygen present in
the flue gases; if there is a deficiency there will be CO
present in the flue gases. An analysis of these gases for
oxygen and for CO would therefore indicate the source of
trouble and the remedy to be provided.

The curve to the right of the central vertical line in
Fig. 179 shows the theoretical decrease in volume per

oj30

£2°

8 15

>
I.

°S
D

[Xx

- —

- — , ,

0 40 30 20 10 0 50 100 150 200 250 301
eflciency (in per cent) Excess (In per cent)
1234

FIG. 179.— Effect of Air Supply on Flue Gas Analysis.

cent of C02 in flue gases as the excess air increases. The
single numbers 1, 2, 3 and 4 indicate the excess coeffi-
cients corresponding to the various percentages of excess
air.

The curves to the left give the theoretical decrease in
volume per cent of CO2 and the theoretical increase in
volume per cent of CO as the air supplied is decreased below
that theoretically required for complete combustion.

127. Combustion of Hydrogen. Hydrogen combines
with oxygen, or burns, to fonn water. The equation for
this reaction is

2H2+02 = 2H20, ...... (88)

288 STEAM POWER

which indicates that two molecules of hydrogen combine
with one molecule of oxygen to form two molecules of
water. In terms of volumes, two volumes of hydrogen
combine with one of oxygen to form two of gaseous water,
that is, water in the form of highly superheated vapor.
4s the water is cooled down it will obviously approach
and finally reach the liquid condition, with a rapid de-
crease in volume quite different from that experienced
by a gas under similar conditions, so that the volume rela-
tions hold only at high temperatures.

The weight relations can be calculated as in ether
cases, starting from the fact that four weights of hydrogen
combine with thirty-two weights of oxygen to form 36
weights of water. The weights of hydrogen and oxygen
are therefore in the relation of -g^ = J.

The heat liberated when one pound of hydrogen burns
to water is equal to about 62,000 B.t.u. This is the quantity
of heat which could be obtained if one pound of hydrogen
at, say, room temperature, and mixed with the requisite
quantity of oxygen, were ignited and the resultant water
were then cooled down to the initial temperature. During
the cooling of the water it would partly or entirely condense
and thus give up some or all of its latent heat of vaporization.
This heat would obviously be included in the calorific
value just given.

In many pieces of engineering apparatus in which
hydrogen is burned the products of combustion are not
cooled to such an extent that the water is condensed. The
latent heat of vaporization would not be liberated under
such conditions, but would remain bound up with the water
vapor. When the water is not condensed the heat liberated
is only about 52,000 B.t.u. per pound of hydrogen. 'This
number is known as the lower calorific value of hydrogen,
while 62,000 is known as the higher calorific value.

Data on the combustion of hydrogen are given in
Table X.

COMBUSTION

TABLE X
COMBUSTION DATA FOR HYDROGEN

(Per pound of hydrogen)

289

Product.

Oxygen Required.

Nitrogen Accompanying Oxygen.

Cu.ft at 62° F.
and 14.7 Lbs.

Pounds.

Cu.ft. at 62° F
and 14.7 Lbs.

H2O

8 96

26.8

361

Product.

Air Required.

Quantity of Product (N
not included).

Heat
Liberated.

Pounds.

Cu.ft. at 62°
F. and 14.7
Lbs.

Pounds.

Cu.ft. at 62°
F. and 14.7
Lbs.

H2O .-.-'

34.8

457

Liquid
0.144

/ 62,000
I 52,000

128. Combustion of Hydrocarbons. Many of the fuels
used by the engineer contain compounds of hydrogen
and carbon which are called hydrocarbons. One of the best
examples is methane (CH4), which forms the greater part
of all the so-called natural gas.

All of these hydrocarbons burn to CO2 and H^O if
the supply of oxygen is great enough. If there is a deficiency
of oxygen, combustion is incomplete and generally results
in the formation of CO2, H^O, CO, C in the form of soot,
and other products which need not be considered here.

For complete combustion the requisite oxygen and
air can be determined as in previous cases by means of
chemical equations. Thus for methane the equation is

= CO2+2H2O,

(89)

which shows that sixteen (12+4) weights of methane
combine with sixty-four (2X2X16) weights of oxygen to
form forty-four (12+32) weights of carbon dioxide and
thirty-six (4+32) weights of water.

290 STEAM POWER

The calorific value of hydrocarbons is generally assumed
to be equal to the sum of the heat values of the carbon
and hydrogen contained in one pound of the material.
Thus, if C represent the fraction of a pound of carbon
contained in one pound of the hydrocarbon and if H
represent the fraction of a pound of hydrogen contained
therein, the common assumption would make the higher
calorific value of the hydrocarbon

(CX 14,600) + (HX 62,000) B.t.u. . . (60)

The results obtained in this way do not generally check
well with the experimentally determined values, and it is
best to use the latter when they are available.

129. Combustion of Sulphur. Sulphur forms several
different oxides, but when burned under engineering con-
ditions it is generally assumed to form only the dioxide
S(>2. The chemical equation for such combustion is

(91)

and since the atomic weight of sulphur is 32, this equation
shows that equal weights of sulphur and oxygen combine
to form the dioxide.

The combustion of sulphur to 862 liberates about
4000 B.t.u. per pound of sulphur.

130. Combustion of Mixtures. It is often necessary
to obtain approximate calorific values of combustible
materials which, without great error, can be considered
as mixtures of combustible and non-combustible elements.
If there is oxygen present in the mixture it is assumed
to be combined with hydrogen in the form of water, so
that the uncombined or available hydrogen per pound of
material is given by the expression

AvailableH = H--, ..... (92)
8

COMBUSTION 291

in which H and O respectively represent the fractions of
a pound of hydrogen and oxygen in one pound of material.
The calorific values of such a mixture containing car-
bon, hydrogen and sulphur would then be given approxi-
mately by the equation

Higher B.t.u. - 14,6000+ 62,000 H- - +4000S, (93)

\ o/

in which the letters stand respectively for the fractions
of a pound of each of the elements present in one pound of
the mixture. Similarly the lower calorific value would
be (approximately)

Lower B.t.u. = 14,600C+ 52,000 ( H-^-) +4000S, (94)

\ P/
and the oxygen required will be

Pounds of O = 2iC+8(H-^\ +S. (95)

131. Temperature of Combustion. If combustion of
any material could be carried on inside of an ideal vessel
which did not absorb nor transmit heat, the heat liberated
during the combustion could not escape from the space
within the vessel.

If the vessel contained initially only the combustible
and the oxygen or air required to burn it, the products of
combustion would be the only material contained within
the vessel after the completion of combustion. Under
such circumstances the heat would be used in raising the
temperature of the products of combustion, and the process
could be pictured as though all of the combustion occurred
first, forming the products of combustion without change
of temperature, and then the liberated heat raised the
temperature of these products.

Knowing the weight of each of these products and the
quantity of heat required to raise the temperature of one

292 STEAM POWER

pound of each of them one degree, the amount of heat
required to raise all of them one degree could be found by
multiplying the two known values. Thus, if carbon had
been burned in oxygen to CCb with the theoretical oxygen
supply, the vessel would contain only carbon dioxide.
To raise the temperature of one pound of carbon dioxide
one degree requires an amount of heat equal to the specific
heat of that gas. Therefore, if W represents the weight
of CC>2 formed and C represents its specific heat, the amount
of heat required to raise the temperature of all of the CC>2
one degree would be W-C B.t.u. If Q B.t.u. were liberated
by the combustion, the temperature rise in degrees would
therefore be given by

Temp, rise- ^, ..... (96)

and if the initial temperature had been to degrees, the final
temperature would be

....... ' (97)

A final temperature figured in this way is called the theo-
retical temperature of combustion. It can never be attained
in practice because of heat lost to surroundings and because
of other losses which need not be considered here.

Theoretical temperatures of combustion are, moreover,
nearly always calculated on the assumption that the specific
heats of gases are constants, whereas they really increase
with the temperature. It therefore follows that tempera-
tures determined on the assumption of constant specific
heat will be too high for this reason also.

When gases are heated there are two distinctly different
limiting possibilities; the volume occupied by the gases
may remain constant or the pressure exerted by the gases
may remain constant while the volume increases. In
the case of constant volume all the heat added to the gases

COMBUSTION 293

must be used for raising the temperature; the amount
of heat required per pound per degree under these con-
ditions is known as the specific heat at constant volume
and is designated by Cp.

When, however, the volume is allowed to increase at
such a rate as to keep the pressure constant the heat sup-
plied must not only raise the temperature, but must also
do whatever external work is done in displacing (pushing
out of the way) surrounding mediums. The heat required
per pound per degree under these conditions is known as
the specific heat at constant pressure and is represented
by Cp. It is always greater than Cv by the amount of heat
required to do the external work accompanying a rise of
temperature of one degree.

Thus, in the case assumed above, had the vessel been
so constructed that its internal volume did not change,
the specific heat at constant volume would be used. On
the other hand, had the vessel been fitted with a movable
piston arranged to move outward at such a rate as to main-
tain constant pressure within the vessel as the temperature
rose, the specific heat at constant pressure would be
used.

In most cases the material burned is not pure carbon,
but a fuel containing carbon, hydrogen and sulphur, and as
air is generally used to furnish the oxygen, the products
of combustion will contain not only the oxides of carbon,
hydrogen and sulphur, but inert nitrogen as well. The
temperature rise is determined in the same way, however,
by dividing the heat liberated by the amount required to
raise the temperature of the products one degree. Thus
if Wi, W2, Ws . . . Wn stand for the weights of the various
products and Ci, €2, C3 . . . Cn for their respective specific
heats, the theoretical temperature rise is given by

' (98)

294 STEAM POWER

and the theoretical temperature of combustion is given by

3CS . . . WnCn'

if to stands for the initial temperature.

PROBLEMS

1. Assume 10 Ibs. of C burned to CO. Determine the quan-
tity of oxygen required, the quantity of air required, the quantity
of nitrogen in this air, and the quantity of heat liberated.

2. What will be the volume of the CO formed as above if
measured at 62° F. and 14.7 Ibs. pressure?

3. Assume 15 Ibs. of C burned to C02. Determine the quan-
tities of oxygen and air required, the quantity of nitrogen con-
tained in this air, and the quantity of heat liberated.

4. What will be the volume of the C02 formed from 15 Ibs.
of carbon if measured at 62° F. and 14.7 Ibs.?

5. What will be the volume of the flue gases formed by the
combustion of 11 Ibs. of carbon to C02 with the theoretical air
supply?

6. The quantity of CO obtained by the combustion of 8 Ibs.
of carbon is burned to C02 with the theoretical amount of oxygen.
Determine the quantities of oxygen and air required, the amount
of nitrogen contained in this air, and the quantity of heat liberated.

7. Assume 5 Ibs. of C burned in air to C02 with an excess
coefficient of 1.5. Determine the quantities of oxygen and air
supplied, the heat liberated and the composition of the flue gases.

8. The composition of flue gases resulting from the combus-
tion of carbon in air is found to be 21% of C02 and 79% of N
by volume. What is the value of the excess coefficient?

9. An analysis of flue gases resulting from the combustion
of carbon in air shows 12% of C02 by volume and no CO. The
gases are not analyzed for 0 or N. What can you say with regard
to the air supply?

10. Three pounds of hydrogen burn with theoretical oxygen
supply. Determine the wreight of oxygen and air used, the weight
of the resultant water and the weight of the flue gas.

11. Determine the heat liberated in the preceding problem if
the water vapor is condensed and if it is not condensed.

12. How much hydrogen would have to be burned to obtain
20 Ibs. of water?

COMBUSTION 295

13. The chemical formula of methane is CH4. If one pound
of methane is burned with theoretical air supply, what weight
of air will be used, and what will be the weight of the flue gases?

14. What would be the percentage composition of the flue
gases of the preceding problem on a weight basis?

15. The chemical formula of ethane is C2H6. Determine the
calorific value of this material by means of the formula given
in the text.

16. A certain material is found to have the following analysis
on a weight basis: C, 85%; H, 12%; 0, 1%; S, 2%. Determine
the calorific value of this material by means of the formula given
in the text, assuming that all the oxygen present is in combination
with hydrogen.

17. Determine the amount of oxygen required to completely
burn 3 Ibs. of the material described in the preceding problem.

18. One pound of carbon is burned to CO2 in pure oxygen
in a vessel so arranged as to maintain constant internal pressure.
The specific heat of C02 at constant pressure and at ordinary
temperatures is about 0.21. Calculate the theoretical temperature
rise and the temperature of combustion, using this value of the
specific heat and assuming an initial temperature of 60° F.

19. Make the same calculations as called for in the preceding
problem, but using the value 0.27 for the specific heat of C02.
This is more nearly the average value of the specific heat over
the range of temperature existing in such a case.

20. The hydrocarbon ethylene is represented by the chemical
formula C2H4. Assume that one pound of this material is burned
in air within a vessel arranged to maintain the products at con-
stant pressure and that the excess coefficient is 1.5. Determine
the theoretical temperature of combustion if the initial temperature
is 60° F., the mean specific heat of CO2 is 0.27, that of H20 is
0.61, that of N is 0.27, and that of 0 is 0.24.

CHAPTER XVI
FUELS

132. Commercial Fuels. In engineering practice any-
thing which is combustible and which can' be procured in
large quantities at a reasonable cost is called a fuel. The
principal commercial fuels are:

f(l) Coal.

a. Solid | (2) Wood and wood wastes.
1(3) Vegetable wastes.

{(1) Crude petroleum or natural oil.
(2) Various products made from petroleum.
(3) Methyl and ethyl alcohol.

c Gaseous j(1) Natural Sas'

[ (2) Artificial or manufactured gases.

Coal is by far the most extensively used fuel because of
its abundance and relative cheapness in most localities.
However, in oil-producing regions the crude oil and some
of the products made from it are more often the commonly
used fuel, particularly if -good coal is not mined in the
immediate vicinity.

Wood is, in general, too valuable to:be used exclusively
as a fuel excepting on the frontiers where wooded terri-
tory is being opened up and where coal cannot be pro-
cured excepting at prohibitively high cost. Wood wastes,
on the other hand, are very often used for fuel in the indus-
tries producing them.

Vegetable wastes, like wood wastes, are essentially of
local value, being practically entirely consumed by the
industries producing them.

296

FUELS 297

Natural gas is in many respects an ideal fuel, and is
extensively used for power production in some localities.
The diminution in the quantity available, the consequent
rise in the price, the great economy achieved in burning
this gas in gas engines and the increased use of the gas for
domestic purposes are, however, gradually eliminating this
fuel from the steam-power field.

Artificial gases have never been extensively used for the
generation of steam, as it is generally cheaper to burn the
materials from which the gases are made, rather than to
convert them into gas and then to burn the gas under
boilers. This condition may change in the future when
better markets have been opened up for some of the by-
products which can be obtained from artificial gas plants.

133. Coal. The word coal is used as the name of a
great group of natural fuels which consist of more or
less metamorphosed vegetable remains. At one end of the
group is the material known as peat, which is only slightly
changed from the original vegetable substance; at the
other end is the graphitic anthracite which has undergone
such radical metamorphosis that practically all of the
original vegetable material excepting carbon and ash has
been eliminated.

A common, rough classification of the coals in the order
of age, or of completeness of carbonization is,

1. Peat or turf.

2. Lignite (brown or black).

3. Sub-bituminous coal.

4. Bituminous coal.

5. Semi-bituminous coal.

6. Semi-anthracite.

7. Anthracite.

8. Graphitic anthracite.

The divisions are not at all exact, as they depend partly upon
chemical composition and partly upon physical properties.

298

STEAM POWER

Another classification of a more exact, variety is that
given in Table XI and partly illustrated in Fig. 180, which
gives what is known as Mahler's curve. It is for United
States coals only. The terms used in this classification
are explained in subsequent paragraphs.

16000

15000

60 70 80 90

#Fixe.d Carbon In the Combustible

100

FIG. 180.— Mahler's Curve for United States Coals.

TABLE XI

CLASSIFICATION OF COALS

Division.

Per cent of
Fixed Carbon
in Combustible.

Per cent of
Volatile Matter
in Combustible.

Calorific Value,
B.t.u. per Pound of
Combustible.

Graphitic
Anthracite
Semi-anthracite
Semi-bituminous
Eastern bituminous. . .
Western bituminous. .
Lignite

100 to 97
97 to 92.5
92.5 to 87.5
87.5 to 75
75 to 60
65 to 50
under 50

0 to 3
3 to 7.5
7.5 to 12.5
12.5 to 25
25 to 40
35 to 50
over 50

14,600 to 14,900
14,900 to 15,300
15,300 to 15,600
15,600 to 15,900
15,800 to 14,800
15,200 to 13,700
13,700 to 11,000

The graphitic anthracite occurs in very small quantities
and mostly in Rhode Island. With a few minor exceptions
the anthracites occur only in Eastern Pennsylvania and the

FUELS 299

semi-anthracites are almost entirely confined to the western
edge of this field.

The semi-bituminous coals are found on parts of the
eastern border of what is known as the Appalachian coal
field, extending from central Pennsylvania through the
intermediate States to the northern part of Alabama. The
greater part of this enormous bed consists of eastern bitu-
minous coal. Western bituminous coals are found in large
beds in the central part of the United States, principally in
the States of Illinois, Indiana and Kentucky on the east of
the Mississippi River, and in Iowa, Kansas and Texas to
the west of that river.

Lignite is found in small quantities in nearly all of the
western half of the United States and in large beds in the
Dakotas, Texas, Arkansas, Louisiana, Mississippi and
Alabama.

Peat is distributed in small beds throughout practically
all of the United States and is continually forming in many
marshes and on low-lying lands.

134. Coal Analyses. Two different coal analyses are
in use, the simpler being known as the proximate analysis
and the more exhaustive being called the ultimate analysis.
Both are made and reported on a weight basis.

The proximate analysis assumes coal to contain four
different and separable things, which are called fixed carbon,
volatile hydrocarbon or volatile matter or volatile, moisture
and ash.

Moisture is determined by maintaining a small quantity
of finely ground coal at a temperature of about 220° F.
for one hour. The material lost during this time is
assumed to be moisture only and is reported as such.
Coal from which the moisture has been driven in this way
is called dry coal.

Volatile matter is determined by heating a sample
from which the moisture has been driven, or a fresh sample.
The coal is maintained at a red to white heat with exclu-

300 STEAM POWER

sion of air until there is no further loss of weight. In
the case of a previously dried sample the loss under these
conditions is called volatile hydrocarbon. If the sample
was not previously dried a separate moisture determina-
tion is made on a similar sample and the weight of volatile
is found by difference.

Fixed carbon is found by combustion of a sample from
which the moisture and volatile have been driven, the
loss under these conditions being assumed to be entirely
due to the combustion of carbon.

Ash is the name given to the incombustible material
left behind after determining the fixed carbon.

The volatile hydrocarbons and the fixed carbon as
determined in the proximate analysis are assumed to be
the only combustible parts of the coal and their sum is
called the combustible.

Proximate analyses are reported in three different
ways: On coal as received, on dry coal, and on combustible.

Since the water content of a sample of coal received
at any plant is largely a matter of the weather conditions
during shipment, the best idea of the character of a coal
can be obtained by excluding the consideration of its
moisture content. It is generally best, therefore, to convert
analyses to a dry coal basis, that is, recalculate the per-
centages of volatile, fixed carbon and ash on the assumption
that the analysis was made on the weight of coal which would
result from drying the sample that was actually used. Ex-
cessive moisture is, however, undesirable for steam-raising
purposes, and the amount of moisture should therefore be
determined in every case.

Ash is also more or less a matter of accident in that the
amount contained is largely determined by the care used
in mining and subsequent cleaning of the coal. While
it has a very appreciable effect upon the character of the
material as a fuel it really has little connection with the
combustible part of the fuel. For purposes of classifica-

FUELS 301

tion, therefore, the ash should also be eliminated and the
analysis given on the basis of combustible.

Sulphur is sometimes reported with a proximate analy-
sis. In making such an analysis the greater part of the
sulphur is really driven off with, and regarded as, part of
the volatile, so that when the sulphur content is desired it
must be determined by a separate analysis.

The ultimate analysis attempts to separate the dry
combustible into the various elements of which it is com-
posed. The percentages of carbon, hydrogen, oxygen,
nitrogen and sulphur are determined as well as the per-
centage of ash in dry coal. Such analyses show the carbon
contents of coal to vary from about 98 per cent in the
graphitic anthracite through about 97 per cent in the
semi-anthracite, 87 per cent in semi-bituminous, 80 per cent
in bituminous and 74 per cent in lignites to as low as 61
per cent in peats. The corresponding figures for hydrogen
run from about 1 per cent through a range in the neigh-
borhood of 5 per cent for semi-bituminous to about 6 per
cent in the case of peat.

Oxygen varies from about 2 per cent or less in the
case of good anthracite to as high as 33 per cent for peat;
nitrogen generally forms about 1 per cent of the dry fuel
and sulphur from 1 to 3 per cent.

135. Calorific Value of Coals. The calorific value of
coals on a basis of combustible has been shown to vary
approximately according to a smooth curve, but the local
variations are so great that no generally applicable formula
for calorific value has yet been proposed. The formula
commonly used is based upon the ultimate analysis and
is similar to that suggested as approximately applicable
in the case of mixtures of combustibles. It is known as
Dulong's formula, and is

f 62,000] / 0\

B.t.u. per Ib. = 14,600C + or ( H- ~ ) +4000S, (100)

1 52,000 J N

302 STEAM POWER

in which the letters refer to the weight of the various ele-
ments contained in one pound of dry coal.

When an accurate knowledge of the calorific value of a
fuel is desired it should be obtained by means of a fuel
calorimeter. There are many varieties of this instrument,
but practically all operate on the same general principle.
A known weight of fuel is completely burned within a vessel
and the heat liberated is absorbed by water or similar
liquid. From measurements of liquid temperatures the
heat absorbed by the liquid can be determined, and this
with some additions for losses of various kinds must be the
heat liberated by the fuel.

136. Purchase of Coal on Analysis. Until quite recently
it was customary to buy coal from the lowest bidder pro-
vided the material supplied could be made to give satis-
factory results in the plant. Obviously the purchaser knew
nothing regarding his purchase, and often bought quantities
of ash and moisture at the price of combustible. Now,
however, the larger power plants and many of the smaller
are buying on the basis of analyses and calorific values as
determined in calorimeters.

A certain desirable standard analysis is set and cer-
tain variations are allowed from it. Wide variations are
penalized by deducting so many cents per ton for each
variation of a certain degree, and, finally, outside limits
are set for moisture and ash beyond which the fuel need not
be accepted. In some cases limits are also set for sulphur.

This is the logical method of purchasing coal in large
quantities, and is sure to come into very general use as its
advantages become known.

137. Petroleum. This material is obtained from drilled
wells and has been found in many widely separated sections
of the country. The oil wells of Pennsylvania and neigh-
boring States, of Oklahoma, Texas and California have been
the most productive and are hence the most widely known.

Natural petroleum, as it occurs in the United States, is

FUELS

303

generally a dark, rather thick, oily liquid with a char-
acteristic odor. It varies widely in composition so far as
the compounds contained are concerned, but the variations
in ultimate composition, specific gravity and calorific value
are comparatively small.

The ultimate analysis of crude oil generally shows about
83 to 85 per cent of carbon, 13 to 15 per cent of hydrogen
and small quantities of oxygen, nitrogen and sulphur.

The specific gravity generally lies between 0.80 and 0.90
and in most cases is nearer the upper figure. It is common
practice to express the gravity in terms of the Beaume
scale, an arbitrary scale developed for an instrument known
as the Beaume hydrometer. This device is arranged to
float in liquids and measures the gravity by the distance to
which it sinks. Various corresponding values of the Beaume
scale and specific gravity are given in Table XII for the
region most used in connection with petroleum.

TABLE XII
CORRESPONDING BEAUME READINGS AND SPECIFIC GRAVITIES

Beaum6 Reading.

Specific Gravity.

Beaum6 Reading.

Specific Gravity.

0.9333

0.8536

0.9210

0.8433

0.9090

0.8333

0.8974

0.8235

0.8860

0.8139

0.8750

0.8045

0.8641

0.7954

The higher calorific value varies between 19,000 and
20,000 B.t.u. per pound and the lower value is generally
1000 to 1500 B.t.u. lower.

Crude oil is sometimes used for fuel, but this is unde-
sirable, for two reasons. First, the crude oil contains
many highly volatile constituents which can be distilled

304 STEAM POWER

off and which have a high market value in the forms of
gasoline and allied distillates. Second, the presence of
these highly volatile constituents in the oil makes it more
dangerous, as combustible vapors are given off in large quan-
tities at low temperatures and the mixtures formed with
the- oxygen of the air are often highly explosive.

As a consequence, the material generally sold as fuel
oil is a residuum left after distilling off the more volatile
constituents of the crude oil. It has practically the same
properties as the crude, excepting that dangerous vapors
are not given off at so low a temperature.

PROBLEMS

1. A sample of coal gives the following proximate analysis:
moisture, 5%; volatile, 4.25%; fixed carbon, 80.75%; and
ash, 10%. Determine the percentage of combustible and the
percentages of fixed carbon and of volatile in the combustible.

2. What variety of coal is indicated by the values obtained
in Prob. 1?

3. The following results were obtained in making a proximate
analysis of a sample of coal; moisture, 7%; fixed carbon, 56.7%;
volatile, 24.3%; ash, 12%. Determine the percentage of com-
bustible and the percentages of fixed carbon and of volatile in the
combustible. What variety of coal is indicated by these values?

4. The ultimate analysis of a sample of dry coal gave the
following results: carbon, 79.12%; hydrogen, 4.14%; oxygen,
1.84%; sulphur, 0.92%; nitrogen, 0.74%; ash, 13.24%. Recal-
culate these values for an ash-free coal.

5. Determine by means of Dulong's formula the upper and
lower calorific values of the coal described in Prob. 4.

6. The ultimate analysis of a sample of crude petroleum from
which all water was removed gave the following results: carbon,
85%; hydrogen, 13%; sulphur, 1.0%; oxygen, 0.25%; nitrogen,
0.12%; ash (sand and similar material), 0.63%. Determine the
upper and lower calorific values by means of Dulong's formula.

CHAPTER XVII
STEAM BOILERS

138. Definitions and Classification. The term boiler
is generally applied to the combination of a furnace in which
fuel may be burned continuously and a closed vessel in which
steam is generated from water by the heat liberated within
the furnace.

Boilers are classified in many different ways, the more
important being given in the following schedule :

CLASSIFICATION OF BOILERS

(a) Plain cylindrical,

(6) Flue,

(1) According to form / \ «Ct i

(d) Sectional, etc.

(2) According to location off (a) Externally fired, and

furnace } (6) Internally fired.

(3) According to use

(a) Stationary,
(6) Portable (as on trucks,
or rollers),

(d) Marine.

ff(a) Horizontal,

(4) According to direction of ; ' T ,. ,

. . . { (6) Inclined,

principal axis , : Tr , . ,

[(c) Vertical.

(5) According to relative posi- f , x ,Tr , , ,

° I (a) Water tube,

tions of water and hoU ' ^. ,

(b) Fire tube,
gases

305

306

STEAM POWER

Examples of boilers of the different types mentioned are
given in subsequent paragraphs.

139. Functions of Parts. It has been shown that there
are two essentially different parts in the apparatus commonly
known as a steam boiler, the furnace and the boiling vessel.
A simple form of boiler known as a horizontal, return tubu-
lar boiler, or an H.R.T. boiler, is shown in Figs. 181 and
182 with the two essential parts and their components

Pressure Regulator

FIG. 181. — Sectional Elevation of H.R.T. Boiler and Furnace.

indicated. The furnace consists essentially of the combina-
tion of grates, bridge wall, fire and ash doors, the ash pit
and the space above the grates. It is the function of the
furnace to so burn the fuel that the maximum amount
of heat will be made available for absorption by the water
within the boiling vessel.

It is the function of the boiling vessel to transmit to
the water within it the greatest possible quantity of the
heat thus made available and to resist successfully the

STEAM BOILERS

307

tendency to rupture under the action of the high internal
pressure, that is, the pressure of the steam.

In the type of boiler shown the fuel is " fired " by
hand, that is, it is
spread on the grate by
being thrown from a
scoop shovel through
the opened fire door.
Air enters through both
doors in regulated pro-
portions and in such
quantities as best to
approximate complete
combustion.

The hot gases re-
sulting from the com-
bustion pass over the
bridge wall, along the FIG. 182.

lower part of the boiler Section through Furnace of H.R.T. Boiler,
shell and then through

the fire tubes, or flues, toward the front of the boiler as
shown by arrows in the figure. From the front end of the
tubes the products of combustion pass up through the
smoke box to " breechings " or " flues," which carry them
to the stack.

Heat is received by the water within the vessel in two
different ways:

(1) The hot fuel bed on the grate radiates energy in the
same way that the sun or any other glowing body radiates
energy. Some of this energy traverses the space between
fuel bed and boiler shell and ultimately passes through that
shell to the water within. The rest of the radiated energy
passes into the walls surrounding the furnace and heats
them and the surrounding atmosphere.

(2) The hot gases of combustion pass over the heating
surface of the boiler, as shown, and transmit part of their

308 STEAM POWER

heat to the water on the other side of those surfaces. The
rest of the heat which they carry is either lost to the surround-
ing walls or is carried up the stack by the gases which
leave the boiler at a comparatively high temperature.
This temperature ordinarily ranges from about 500° to 700°
F. and in extreme cases goes even higher.

140. Furnaces and Combustion. In most forms of
boiler the water within the boiler has practically the same
temperature as the steam being generated, and this is
generally from 320° to 400° F. Obviously the products
of combustion cannot be cooled by the water to a tem-
perature below that of the water, so that the gases leaving
the boiler in an ideal case would have a comparatively
high temperature. Practically, it is found undesirable to
attempt to reduce the temperature of the gases to a value
even approximating that of the water and, as indicated
above, they are discharged at a temperature several hundred
degrees higher. In order that the maximum amount of
heat may be made available for the boiling vessel the prod-
ucts of combustion must therefore leave the furnace with
the highest possible temperature, and the ideal furnace
would completely burn the cheapest fuel available in such
a way as to give this highest possible temperature and not
to generate smoke.

Real furnaces fall far short of this ideal performance,
for numerous reasons. The more important of these are
given in the following paiagraphs:

(a) Complete Combustion of Carbon. In a real furnace
the combustion of the carbon of the fuel may be incomplete
in two senses; first, some of the carbon may remain entirely
unoxidized and pass off with the ash, and second, some of
the carbon may be burned to CO instead of to CO2.

Imperfect combustion of the first kind can result from
fuel falling through the openings in the grate before it has
been ignited or when only partly burned, or it can result
from failure to get air to some of the carbon in sufficient

STEAM BOILERS 309

quantities to burn it completely before all of the surrounding
fuel has been converted into ash and the locality cooled
down to such an extent as to allow the unburned carbon
in its midst to cool below the temperature of ignition.

Imperfect combustion of the second kind, resulting
in the formation of CO, generally results either from a
lack of sufficient air above the fuel bed or from an excessive
quantity of air above the bed. In any furnace there is a
tendency toward the formation of CO within the bed of fuel,
and the deeper the bed the greater this tendency. If the
CO thus formed meets sufficient air after leaving the fuel,
and if the temperatures of CO and air be sufficiently high,
it will burn to CO2. Part of the air for what may be called
the secondary combustion will always work its way through
the fuel bed, because it is impossible to bring all oxygen in
the air passing through into contact with carbon of the fuel.
The remainder of the air required is generally admitted
through the fire door and is heated by passing over the front
part of the fuel bed. If too great a quantity of air is ad-
mitted in either way its temperature may be so low as to
cool the CO below its temperature of ignition and thus fail
to accomplish tfie object sought.

It has been shown that the combustion of pure carbon
with the theoretical air supply would give gases containing
about 21 per cent by volume of CO2. If the combustible of
real coal be assumed to consist entirely of carbon, the same
proportion of CO2 would result from ideal combustion.
Practically, it is so difficult to bring the oxygen of the air
into contact with the carbon of the fuel that a large excess
is always used, the excess coefficient ranging from about
1.3 to over 2 and averaging about 1.5 to 1.7 under very
good conditions. The latter figures correspond to percent-
ages of CO2 of 14 and 12 respectively, but a value as low
as 10, which corresponds roughly to an excess coefficient
of about 2, is not at all uncommon and is generally regarded
as a very good result except in the largest plants.

310 STEAM POWER

(b) Complete Combustion cf Hydrocarbons. The hydro-
carbons which appear as volatile matter in the proximate
analysis are practically all distilled from the fuel, as it is
heated in the furnace before ignition in the same way as
when making a proximate analysis. If they are to be
completely burned they must be mixed with the requisite
quantity of air after distillation and both the vapors and the
air must be maintained at a sufficiently high temperature
until combustion is complete. Part of the air for the com-
bustion of distilled volatile filters through the fuel bed
and the rest must be admitted through the fire door or in
some equivalent manner.

If the flame formed by burning hydrocarbons is allowed
to come in contact with cold surfaces, as, for instance,
the heating surfaces of the boiler, the gases are cooled
below the temperature of ignition and combustion ceases.
This results in the deposit of soot (unburned carbon) upon
the heating surfaces of the boiler and in the carrying of
soot and unburned hydrocarbons up the stack. The soot
and some of these hydrocarbons form the unsightly smoke
so familiarly associated with some stacks.

Or, if the air supply is at a sufficiently high temperature,
but is insufficient in quantity, the hydrocarbons are in-
completely burned and smoke results.

The formation of smoke can be conveniently studied
by means of the ordinary kerosene lamp. Such a lamp
operates by burning hydrocarbons of the same general
character as those distilled from solid fuels. The hydro-
carbons are drawn up by the wick in the form of liquids,
are vaporized by heat near the top of the wick and then
combine with oxygen from the atmosphere to give the
luminous kerosene flame.

If the flow of kerosene and the air supply are properly
adjusted and if the temperature is high enough, the com-
bustion results in the formation of invisible and practically
odorless gases. If, however, the air supply be decreased

STEAM BOILERS 311

or be greatly cooled, a very smoky and very odorous combus-
tion ensues. The same result could be obtained by the use
of too great a quantity of air, a condition often attained
when the supply of kerosene in the bowl of the lamp is almost
exhausted.

The effect of a cold surface is easily seen by inserting
a cold metallic or porcelain surface into the tip of the flame
and then withdrawing it. It will be found covered with
soot.

(c) Advantages and Disadvantages cf Excess Air. It
has been shown that excess air is practically necessary in
the real furnace in order to insure against a deficiency at
any point, and it is thus advantageous in that it makes the
combustion more nearly complete than would otherwise
be the case. On the other hand, excess air represents just
so much excess material to be heated at the expense of heat
liberated by combustion and hence decreases the maximum
temperature attained. A sufficiently great supply of excess
air could so reduce the temperature that even if combus-
tion were complete very little heat would be made available
for absorption by the boiling vessel, because the temperature
attained by the products of combustion would be too low.

Excess air in large quantities may also result in cooling
unburned gases before combustion to such an extent as to
make the completion of combustion impossible.

141. Hand Firing. The commonest type cf furnace is
that shown in Figs. 181 and 182, and the commonest method
of hand firing consists in spreading a layer of fuel as evenly
as possible over the entire surface of the fuel bed as often
as required to replace the fuel burned away. At such inter-
vals as experience shows to be necessary the fire is cleaned,
that is, the ashes are worked out from under the fuel by
means of slice bars, so that practically nothing but live
fuel resting on a thin layer of ash remains behind.

This method is open to many serious objections; the more
important are:

312 STEAM POWER

1. There is a gradual increase in thickness of fuel bed
from the time of one cleaning until the time of the next.
This gives a constantly changing set of requirements for
the proper proportions of air entering below and above
the fuel bed and a constantly changing resistance to flow
of air through the bed, so that great skill is necessary if the
best conditions are to be maintained throughout.

2. There is always a tendency for a fuel bed to burn
faster at some points than at others, due to the accidental
distribution of fuel, ash and air. Where " holes " are
formed in this way large quantities of comparatively cold
air can pass through with the consequences already enumer-
ated. It takes considerable skill and watchfulness on the
part of the fireman to prevent the formation and continued
existence of such holes.

3. The firing door must be opened wide every time
that fuel is to be fired, that is, at intervals varying from two
or three minutes to fifteen or more, depending on load,
character of fuel, etc. While the door is open large quanti-
ties of cold air readily flow into the furnace and cool down
all parts of it, and a proportionately smaller amount will
ordinarily pass through the fuel bed. The result of this on
the flue gases and operation of the boilers has already been
considered, but there is another result of equal or greater
importance. As a consequence of this action the volatile
hydrocarbons distilled off from the freshly fired fuel, which
are themselves at a comparatively low temperature, are
surrounded on all sides by cooled walls and come in contact
with cold air only. The chances of their burning completely
are very slight, and a great part of these volatilized materials
passes off unburned as invisible gas and as smoke. Ob-
viously the greater the volatile content the greater the dif-
ficulty, so that anthracite causes least trouble in this way,
while most bituminous coals give heavy black smoke when
burned under these conditions.

The cooling down of the interior of the furnace during

STEAM BOILERS 313

firing is accompanied by the covering of the fuel bed with
cold fuel, so that, for the time being, very little radiant
heat enters the boiling vessel, and the gases which come in
contact with its surface are comparatively cool. The
maintenance of a constant steam pressure under these con-
ditions is practically impossible, but the difficulties can be
partly overcome by very frequent firing of small quantities,
so that the door is open a very short time and also that the
layer of fuel is very thin and does not cut off much heat.

4. The cleaning of the fire necessitates keeping the
fire door open for several minutes, with results of the same
variety as -those just enumerated.

Summing up these difficulties, they divide themselves
into two classes — those which can be almost or entirely
eliminated by skill of a very high order and those which are
inherent and cannot be eliminated by skill. It will also
be observed that all should give more trouble with fuels
high in volatile than with those of the anthracite variety,
both as to incomplete combustion and to the formation
of smoke.

Several other methods of hand firing have been proposed,
particularly for use with bituminous coals, and some of
them have been successfully utilized in isolated instances.
Nearly all depend upon covering only part of the fuel bed
at one time and, by alternating the parts covered in this
way, fresh fuel on one part of the bed is coked while air is
heated by coming in contact with the uncovered incandescent
part of the bed and is therefore in proper condition to burn
more perfectly the volume of hydrocarbons being distilled
off. These methods are all good, but they involve a great
deal of careful work and a high degree of skill on the part of
the fireman.

Other methods of eliminating some of the difficulties
depend upon modifications of the furnace and air supply.
Most attempt to entirely surround the fuel and the gases
given off with heavy masses of brick work and tile, so that

314

STEAM POWEE

enough heat will be stored during incandescent periods
to tide over the periods of cooling. Some forms have
combined with this idea a series of air ducts in the brick
work so arranged that air on its way to the furnace passes
through these ducts and is heated. In some cases the air
supply is automatically controlled and more air is supplied
above the fire during the period of distillation, or coking,
as it is called, than during the following period, when the
coked coal is brightly incandescent and little volatile
matter is present.

MM -3yAT.CuVy.v\v- -f^a^' . -.O-

'uiMfiUi'

pra3S»Wi^:t
If WS

ewm.

FIG. 183.

In some hand-fired furnaces which are intended for use
with bituminous coals that give a long flame the parts of the
boiling vessel within range of the flames are covered with
tiles. This prevents impingement of unburned gases upon
cool surfaces and thus tends to prevent the formation of
smoke and incomplete combustion.

Carrying this principle to its logical conclusion results
in the installation of the grate in a firebrick chamber in
front of the boiler-setting proper, as shown in Fig. 183.
Such a device is known as a Dutch oven and is often very
efficient in totally or partially preventing the formation of

STEAM BOILERS 315

smoke. It does not, however, give as high an economy
as might be expected, because a great part of the radiant
heat of the fire does not reach the boiler surfaces and because
the large external surface results in great radiation losses
to atmosphere.

Another interesting modification consists of reversing
the direction of the draft, that is, the direction in which the
air passes through the fuel bed. The type of furnace al-
ready described is known as an updraft furnace, because the
air passes upward in flowing through the bed. The modi-
fied type here referred to is called a downdraft furnace,
because the air flows downward in passing through the fuel.

In downdraft furnaces the coal is fired on top of the
grate as in other types, but the air is admitted above, flows
downward toward what would normally be the ashpit,
and from there on over the heating surfaces of the boiler.
Fresh coal fired on top of the incandescent bed in such a
furnace distills as in other types, but the volatiles are mixed
with the entering air and are carried downward through the
hot bed so that ideal conditions for combustion are more
nearly attained. In some forms there is a second updraft
grate beneath the downdraft grate. This second grate
receives partly burned coals falling through from the upper
grate arid holds them until combustion is practically com-
pleted.

In downdraft furnaces the grate bars are generally made
of pipes, and water, from the boiler or on its way to the
boiler, is circulated through them. If this were not done
the grates would quickly warp out of shape and ultimately
burn away because of the high temperatures to which they
are subjected.

142. Mechanical Grates. In order to overcome the dif-
ficulties arising from opening the doors for the purposes of
cleaning the fire, numerous so-called rocking, shaking,
self -cleaning, or dumping grates have been developed.
These are generally built up of grate bars which have a

316

STEAM POWER

rough T or an inverted L section with the upper horizontal
branch of the T or inverted L slightly rounded, as shown in
Fig. 184. These bars are arranged in groups with their
longitudinal axes running across the grate, and they are so
supported that they can be rocked about a point in the verti-
cal leg of the T or L by means of levers located at the front
of the boiler. By rocking the bars the lower part of the fuel
bed which has been burned to ash can be dropped into the
ash pit, while the upper part is sufficiently agitated to close
up holes which may have formed, and this can all be done

FIG. 184.

with the doors closed. Or, if desired, part or all of the fuel
bed can be dropped into the ash pit by a similar rocking
motion.

143. Smoke and Its Prevention. An idea of the reasons
for the formation of smoke will have been obtained from the
preceding paragraphs. A reasonably skillful fireman should
have little difficulty in burning anthracite coals in the simpler
forms of furnaces without smoke, but it is almost impossible
to commercially burn many of the varieties of bituminous
coals in this way without the formation of excessive volumes
of dense black smoke at intervals immediately following-
each firing.

STEAM BOILERS 317

Aside from all aesthetic and sanitary considerations,
smoke is undesirable because it represents poor furnace
conditions and waste. The actual loss of carbon in visible
smoke is generally almost negligible in comparison with
the other losses in the form of unburned hydrocarbons,
the lowered initial temperature, etc. All of these losses
combined represent a waste of considerable magnitude.

The proper method of smoke elimination is not the
combustion or removal of smoke already formed, but it
is the burning of fuels in such ways as not to form any
appreciable quantity in the first place. To accomplish
this end the following must be achieved:

1. Coal must be fired continuously and uniformly
without the. opening of doors which admit cold air to the
furnace.

2. Volatiles must be distilled continuously and uni-
formly and in such a place that they are given ample oppor-
tunity to mix with proper proportions of air and to bum
completely before coming in contact with cool surfaces.

3. The air supply must be properly controlled and
tempered to meet the demands of the fuel both in and
above the bed.

4. The fire bed must be worked continuously and
uniformly so as to eliminate ashes as rapidly as formed
and to maintain a bed of uniform depth and condition.

Some of these necessary conditions can be attained by
the use of the various forms of hand-fired furnaces already
described but, even in the hands of skillful and industrious
men, it is impossible to meet all of them. Mechanical
stokers which more nearly approach the ideals set have
therefore been developed and are widely used.

144. Mechanical Stokers. These mechanical devices are
useful for two reasons — they eliminate a great deal of labor
and they make possible the burning of many varieties of
refractory fuels without the formation of excessive quanti-
ties of smoke.

318 STEAM POWER

Despite the good results which can be achieved by
their use, mechanical stokers are not installed in small
plants as often as might be expected. This is because
good stokers are very expensive in comparison with hand-
fired furnaces and, despite economy of fuel, do not generally
show a financial saving unless their use eliminates the
services of several firemen.

It is generally assumed that one man can care for water,
coal and ashes for about 200 boiler horse-power or can
handle coal only for about 500 boiler horse-power. Expe-
rience has shown that one man can care for about 2000 to
5000 boiler horse-power when the boilers are equipped with
good stokers and coal-handling apparatus.

Financial calculations will generally show stokers to
be profitable investments for plants of 2000 or more boiler
horse-power. Where they are installed in smaller plants,
the absolute necessity of eliminating smoke or the use of
very poor varieties of coal have generally dictated their use.

Mechanical stokers can be roughly divided into two
types, those which duplicate hand spreading of fuel and
are known as sprinkler stokers, and those which supply
fuel at one or more points and work it progressively toward
the ash end of the apparatus as it burns. The first type
has not been widely installed, though it is possible that
it may meet with more popular approval after further
development.

Stokers of the second type may be roughly divided
into five classes, which are

1. Chain grates.

2. Inclined stokers.

3. Underfeed stokers.

4. Combinations of above.

5. Powdered coal stokers.

A chain grate, as made by the Illinois Stoker Company,
is illustrated in Figs. 185, 186, 187, and 188. It consists

STEAM BOILERS

319

320

STEAM POWER

of a broad chain made up of a great number of small links
and carried on toothed wheels and roller wheels supported
in a frame which can be wheeled into position within the

FIG. 186.— Sprocket and Links of Illinois Chain Grate.

TOP VIEW OF CHAIN

SHOWING DISTRIBUTION OF AIR SPACES

BOTTOM VIEW OF CHAIN

SHOWING ROLLERS FOR DRIVING-SPROCKET
ENGAGEMENT

FIG. 187.

boiler setting. The general arrangement of the chain and
rollers is shown in Fig. 185; details of the front or driving
rollers and of the links are shown in Fig. 186; a top and bot-
tom view of part of the chain is given in Fig. 187; and
Fig. 188 is a perspective view of the frame showing the

STEAM BOILERS 321

tracks on which it may be rolled into and out of the boiler
setting.

The chain is driven slowly in the direction indicated
by the arrows in Fig. 185 by power applied, through worm
gearing, to the shaft of the toothed wheels at the front
of the stoker. Coal feeds automatically from the hopper
by gravity and is carried into the combustion space by
the moving chain, the thickness of the bed being controlled

FIG. 188. — Framework of Illinois Chain Grate.

by the height of the adjustable gate shown. As the fuel
enters the furnace it passes under the coking arch, which
spans the entire front part of the grate and which is main-
tained at a high temperature by heat radiated from the in-
candescent fuel nearer the inner end of the grate. The
volatiles are distilled from the fresh coal by heat received
from this arch and are heated and mixed with air at this
point. The coked fuel is then carried on into the furnace
and burned, the refuse being discharged at the bridge wall.
If the thickness of bed and speed of chain travel are
properly adjusted, all of the fuel can be coked before pass-
ing out from under the arch and can be burned almost

322

STEAM POWER

completely before reaching the bridge wall, so that prac-
tically ashes only will be discharged.

The apron shown at A in Fig. 185 is used to prevent
the free passage of air to the part of the chain carrying

STEAM BOILERS

323

practically nothing but ash, as this would result in excessive
dilution of the products of combustion.

A stoker of this type installed under a horizontal return-
tubular boiler is shown in Fig. 189. In the illustration part
of the side frame of the stoker is broken away in order to
show the chain and its roller guides. The eccentric shown

FIG. 190. — Details of Feed Mechanism, F.onsy Stoker.

near the top of the front of the boiler drives the chain
through an arm of adjustable length, which makes possible
the control of the speed of chain travel.

An inclined stoker with front feed and a step grate,
known as the Honey stoker, is shown in Figs. 190 and 191.
The fuel is fed out of the hopper and onto the dead plate
by means of the reciprocating pusher. From the dead
plate it is pushed down upon the grate bars by the follow-
ing fuel. These bars are rocked mechanically so that their

324

STEAM POWER

STEAM BOILERS

325

tops alternately assume horizontal and inclined positions,
and this action feeds the fuel downward until it is dis-
charged onto the dumping grate. The material collect-
ing on this grate is periodically dropped by hand into the
ashpit.

The fuel is coked while passing under the coking arch

FIG. 192. — Transverse Section of the Murphy Stoker.

and the coked material is practically completely burned
by .the time it has traveled down the grate. The volatiles
are mixed under the coking arch with heated air which has
passed through the grate and with heated air forced in
above the fuel by steam jets.

An inclined stoker of the side-feed type with bar grates,
known as the Murphy stoker, is illustrated in Figs. 192,
193 and 194. This stoker is provided with two coal-

326

STEAM POWER

magazines or hoppers which are placed horizontally in
the side walls of the boiler setting and feed fuel onto the
inclined grate bars, Fig. 192, which carry it downward
toward the lower point of the V formed by the grates.
The grate bars, Fig. 194, are alternately fixed and mov-
able, the movable bars being hung from above and their

%%pilfe^^

r^^'^/fU^?:^:^^::;^

FIG. 195.— Longitudinal Section, Murphy Stoker.

lower ends being moved up and down by power furnished
by a small steam engine.

A toothed bar arranged for rotation by hand or by power
is located at the bottom of the V and is used for grinding
up ash and clinker which is too large to fall through into
the ash pit. This bar is kept cool by making it hollow

STEAM BOILERS

327

and connecting one end to the smoke flues or stack so
that air is constantly drawn through it.

The location of the coking arch and the method used
j"or supplying warm air should be evident from the figures.

A stoker of this variety is shown in place under a hori-
zontal water-tube boiler in Fig. 195.

An underfeed stoker made by the American Stoker Com-

Movable
Grate Bar

FIG. 194.— Grate Bars of Murphy Stoker.

pany is shown in Figs. 166 and 197. Coal is fed from the
hopper onto the reciprocating bottom B by means of the
reciprocating pusher P. The latter forms the bottom of
a trough as shown in Fig. 197, and its reciprocating motion
feeds the coal upward and out of this trough so that it
spills over onto the inclined grate bars. The reciprocating
motions are all obtained from the direct-acting steam cylin-
der shown.

The inclined grate bars are alternately fixed and mov-

328

STEAM POWER

able, the movable bars sliding back and forth at right angles
to the trough .under the action of horizontal rocking

bars 7?. This action gradually feeds the fuel downward
and toward the side of the furnace, the refuse finally land-
ing on the dumping trays shown.

STEAM POWER

329

FIG. 196. — Longitudinal Section of Class E American Stoker.

FIG. 197. — Cross Section of Class E, American Stoker.

330

STEAM BOILERS

331

Air enters the duct below the trough through the
adjustable gate G, controlled by crank C, and part of it
passes out through holes H near the top of the trough
Fig. 197. The remainder passes down through the hollow

FIG. 199.— Taylor Stoker Under Horizontal Water-tube Boiler.

grate bars and into the heated air box from which it flows
upward between the grate bars.

It will be observed that the coal is fed onto the grate
from below, so that all volatiles distilled off must pass up-
ward through the incandescent fuel before entering the
space above the fuel bed. Part of the air which is to burn

332 STEAM POWER

this volatile matter also passes through the fuel bed and
the remainder flows over the incandescent fuel from the
opening shown near the hopper in Fig. 196. The air and
volatiles are thus raised to a high temperature and well
mixed, and the operation is continuous and uniform, all
tending to facilitate smokeless combustion.

Another variety of underfeed stoker known as the
Taylor stoker is shown in Fig. 198 (a), (6) and (c).
This stoker is built up of alternated retorts and air
boxes, the proper number to give the desired width
of stoker being used. Coal is fed from the hoppers
into the retorts by the upper ram or plunger shown
in Fig. 198 (b) and part of it is again pushed forward by
the lower ram or plunger. The stroke of the lower
plunger can be regulated and in this way the relative
quantities of coal pushed forward in the upper and lower
parts of the retorts can be controlled. The coal spreads
over the tuyere blocks which form the inclined tops of
the air boxes and forms a comparatively even, inclined
layer of fuel.

Coking proceeds under the incandescent fuel which
forms the upper surface of this layer, and the volatiles
mix with air entering through the hot tuyeres and pass
upward through the hot fuel above.

In this stoker advantage is often taken of the fact that
the draft (pressure of air) required with underfeed stokers
is so great that it can be more economically attained by the
use of a fan than by the use of a stack. The fan and the
coal-feeding plungers are both connected to one engine
and the speed of this engine is automatically controlled by
the steam pressure within the boiler. As this pressure
decreases the engine speeds up, thus delivering more coal
and air and as the pressure increases the engine slows down
with opposite results. By properly fixing the travel of the
plungers initially, the best relative proportions of air and
coal are set for the entire range of loads to be carried and the

STEAM BOILERS 333

variation of both is thereafter in approximately the same
proportions.

A stoker of this type in position under a horizontal
water-tube boiler is shown in Fig. 199. A double-ended
arrangement of Taylor stokers as used under very large
water-tube boilers is shown in Fig. 200.

Powdered-coal stokers have been invented in great
number and are successfully used in several of the indus-
tries. They have not, however, "been extensively used
under steam boilers, although isolated installations have
been reported as giving satisfactory results.

In all cases, bituminous coal is crushed to a fine powder
and injected into the furnace with the necessary air for com-
bustion, the air under pressure generally being made to
mechanically entrain the coal dust and carry it into the fur-
nace. The mixture of fine coal and air gives an intensely
hot blow-pipe type of flame, and firebrick and tile are
generally used to prevent it from impinging directly upon
metallic parts.

Oil firing is essentially a mechanical, rather than a manual
process, and while oil burners are not ordinarily under-
stood as belonging to the class of mechanical stokers, they
have all the essential characteristics of such apparatus.

To burn oil successfully under a boiler it must be finely
atomized and mixed with the necessary quantity of air,
and there must be sufficient open space within the furnace
for the free development of the flame and the completion
of combustion before impingement on cool surfaces.

Oil-burning furnaces are generally given a rather large
volume; considerable firebrick is used in such ways as to
give incandescent walls and baffles to assist ignition and
combustion, and all heating surfaces are arranged so that
they are not in the direct path of the flame.

The atomization of the oil is effected in two distinctly
different ways. In some forms of burners it is brought
about by mechanical means, the oil being pumped through

334

STEAM POWER

FIG. 200. — Double-ended Arrangement of Taylor Stoker under Sterling
Type W. Boiler.

STEAM BOILERS 335

a nozzle of some sort which is so shaped that the issuing
jet breaks up into a great number of very small particles.
In other forms, steam is used to break up the jet, the steam
and oil entering the body of the burner separately and later
coming into contact in such a way that the oil is literally
torn apart by the steam. This form of burner has been
more extensively used in the United States than has the
former.

Oil burning shares with the burning of powdered coal,
the property of permitting very accurate regulation of the
air supply to suit the quantity of fuel being burned. The
excess coefficient may therefore be maintained at a low value
and the initial temperature may be made correspondingly
high. Part of the advantage thus gained over the com-
moner methods of coal firing is, however, counterbalanced
by the quantity of steam used for heating and pumping the
oil and for atomizing in some forms of burners.

Both oil burning and powdered-coal burning can be
easily made to give smokeless combustion in properly
designed furnaces and both yield readily to forcing. That is,
the temporary consumption of excessive quantities of fuel
to tide over short demands for excessive amounts of steam
is comparatively easily effected.

145. Rate of Combustion. The rate at which coal is
burned in a given furnace or on a certain grate is generally
given in terms of pounds cf coal fired per square foot of grate
surface per hour and is referred to as the rate of combustion.

The rate at which coal can be consumed is largely
dependent on the intensity of draft available, that is, on
the air pressure available for driving air through and over
the bed of fuel. The higher the pressure available, the
greater will be the quantity of air which can be supplied
and the greater will be the quantity of coal that can be
burned. If it were not for the cost of creating the draft,
the only limit to increasing the rate of combustion would
occur when the velocity of the air became so great that the

336 STEAM POWER

fuel would be picked up from the grate and carried
onward into the flues in a partly burned condition. Com-
mercial drafts give pressure differences above and below
the fuel bed which range from about 0.1 inch of water to
as high as 8 ins. In stationary plants the pressures generally
range from 0.1 to about 0.5 in cases where hand firing is
employed, and are carried as high as 5 or more inches
of water with some forms of mechanical stokers.

The best rate of combustion varies with the type and
size of fuel, the type and size of furnace, the type and size
of boiler, the draft and many other considerations. In ordi-
nary power-plant practice the rates of combustion com-
mercially used generally fall within the following limits:
with anthracite, 15 to 20 Ibs.per square foot per hour; with
semi-bituminous, 18 to 22 Ibs.; and with bituminous, 24 to
32 Ibs. There is a rapidly growing tendency to exceed these
values, particularly in the case of large plants.

As practically all of the volatile is consumed above the.
grate, the fixed carbon content is practically the determining
factor, since it is this constituent that is burned on the
grate. This explains the high rate possible with fuels with
high volatile content. The most economical results are
generally obtained when from 12 to 16 Ibs. of fixed carbon
are consumed per square foot of grate per hour.

The figures given above do not represent limiting con-
ditions. In torpedo-boat practice, where high-draft pres-
sures are used (from 4 to 8 ins. of water), rates of from 50
to 120 Ibs. are attained. On locomotives, which also use
high-draft pressures, rates of combustion greatly in excess
of stationary practice are generally used.

The capacity of a given boiler, that is, its ability to
generate steam, increases as the rate of combustion is
increased, since more heat is thus made available. The
economy of the combination, that is, pounds of steam
generated per pound of coal fired, increases until some best
rate of combustion for the fuel in question is reached, and

STEAM BOILERS

337

thereafter decreases. The variation of economy is, however,
not very great for a comparatively wide range of combustion
on either side of the best rate.

Curves giving approximate draft pressures required fc.v
different rates of combustion when different kinds and sizes
of fuel are hand fired are given in Fig. 201. The sizes
referred to are explained in Tables XIII and XIV. Table
XIII also shows the relative increase of ash content as the

0 5 10 15 20 25 30 35 40

Pounds of Coal per Sq. Ft. of Grate Surface per Hour

FIG. 201.— Draft Required for Different Rates of Combustion with
Different Sizes and Kinds of Fuel.

size decreases, there being a tendency toward the concen-
tration of the ash in the smaller sizes.

146. Strength and Safety of Boiler. Attention has
already been called to the fact that the boiling vessel has to
be designed with two different requirements in view: it
must be mechanically strong to resist internal pressure and
it must transmit the maximum amount of heat to the con-
tained water.

Spherical and cylindrical surfaces with the pressure act-

338

STEAM POWER

TABLE XIII

SIZES OF ANTHRACITE COAL

(Sizes larger than pea coal generally too costly for power-plant use.)

Through Screen

Over Screen

Ash Content

Name.

with Mesh.

(Average) .

(Inclusive.)

Run of mine

unscreened

Broken.

Eetr

Stove

Chestnut

Pea

Buckwheat No. 1

Buckwheat No. 2 or rice. . .

1
4

TABLE XIV

SIZES OF BITUMINOUS COALS

(Considerable variation in commercial practice in naming and sizing.)

Through Bars

Over Bars Spaced

Name.

Spaced Apart.

Apart.

(Inches.)

Lump

Nut

Slack

ing on the inside of the curve are best adapted to resist
such pressures, as they already have the shape which the
pressure would tend to give them. Boilers are, therefore,
constructed as far as possible of vessels having only spherical
and cylindrical surfaces.

Flat surfaces which are poorly adapted to resist such
pressures as act within a boiler must often be used despite
their weakness. When incorporated in a boiler they are
invariably " stayed," that is, braced by being fastened to
other surfaces by stay bolts and other forms of fastenings.
Examples will be given later.

Most of the early designs of boilers and many of the

STEAM BOILEES

339

modern types consist of large cylindrical vessels made by
riveting together properly shaped steel plates. These shells
are often traversed from end to end by flues or tubes for
carrying hot gases and generally have flat ends more or
less perfectly braced by these tubes and by long tie rods

FIG. 202. — Lap Joint.

FIG. 203.— Butt Strap Joint.

and other braces. Such boilers when in operation are almost
entirely filled with water and often hold many tons.

Boilers of these types have been responsible for many
disastrous boiler explosions, and this fact has led inventors
to the development of models which should be less dangerous.
It seems practically impossible to develop a commercial
boiler which cannot be made to explode to a certain extent

gritndinal

FIG. 204.— Riveted Plates of Boiler Shell.

FIG. 205.

if sufficiently mistreated and mishandled; but much can be
done to minimize the danger.

The great weakness of the older forms lies in the riveted
joints, which can never be made as strong as the plates
which they fasten together. Two types of joint are in use;
they are known respectively as the lap joint and the butt
strap joint. These are shown in Figs. 202, 203 and 204.

340 STEAM POWER

So far as a circumferential seam, that is, one running around
the cylinder as shown in Fig. 204, is concerned, the lap joint
is perfectly satisfactory and is universally used. With
longitudinal seams, however, this is not the case. A lap
joint throws the joined edges out of a true cylindrical
surface as shown in Fig. 205, and when the vessel is subjected
to pressure there will be a tendency for the plates to assume
a cylindrical contour as nearly as possible. This causes
local bending of the plates on each side of the lines of rivets,
and the continued repetition of this action ultimately causes
failure. The conditions are often made still worse by calking
the joint on a line indicated by a in Fig. 205, that is, by
hammering the metal at the inner surface of the edge of
the outer plate into firmer contact with the outer surface
of the inner plate for the purpose of making a tight joint.

The butt-strap joint can obviously be made so that the
joined plates more nearly form a true cylindrical surface.

Other weaknesses of the older forms lie in the flat surfaces
used; in constructions which render it possible for sediment
to collect on heated surfaces and thus permit local over-
heating of the plate; and, above all, in the very large
quantity of water contained.

The disastrous consequences of boiler explosions are
generally due to the action of the hot water contained within
the boiler and not to the steam contained at the time rupture
occurs. The water within the boiler is under steam pressure
and approximately at steam temperature. Removal of the
pressure by rupture of the container would enable a great
part of this water to flash suddenly into steam at the
expense of its own heat, and this is exactly what occurs in
the case of a boiler explosion. Local failure causes a sudden
lowering of pressure, and this results in the formation of
large volumes of steam which, blowing out through the
initial fracture, tend to enlarge it, to move the boiler and
surroundings, and, in general, to do all possible to further
the rupture and make conditions worse.

STEAM BOILERS 341

From the preceding discussion the requirements for
maximum safety can be deduced. They are:

1. The smallest convenient diameter of cylindrical ves-
sels, so as to decrease the total load on joints for any
given steam pressure.

2. The elimination of the greatest possible number of
riveted joints and the use of butt-strap longitudinal joints
on all large-diameter, cylindrical vessels.

3. The substitution of curved surfaces for all flat stayed
surfaces.

4. So shaping the boiler that the required extent of
heating surface may be obtained without enclosing a great
volume to be filled with hot water when the boiler is
steaming.

5. So shaping the boiler that such water as is contained
therein will be divided up into small masses contained
within separate vessels connected in such a way that rapid
flow of all water toward one point of failure is impossible.

6. So shaping the boiler that no riveted joints shall be
in the paths of flames and that no sediment can collect on
metal immediately over flames or exposed to very hot gases.

7. So shaping the boiler that it shall
be free to expand and contract with
changes of temperature, with the least
resultant strain on the different parts.

These various requirements are most
nearly met in the different forms of water-
tube boilers, some of which will be de-
scribed in succeeding paragraphs.

147. Circulation in Boilers. If a flask

of water, such as that shown in Fig. 206, *?G: 206'~ .

. . . lation in a Flask,

be heated in the manner indicated, the

water will gradually acquire motion and follow paths
such as those shown by the arrows in the illustration. The
heated water will rise in the center of the mass and the
cooler water will flow downward around the outer surface.

342 STEAM POWER

Such motion is called circulation. Rapid circulation within
a boiler is very desirable, since it brings the maximum quan-
tity of water in contact with the heating surfaces in a given
time and hence tends to increase the amount of heat taken
from those surfaces. It also tends to sweep along any
bubbles of steam or gas formed on such surfaces and to carry
away any sediment which may have collected, thus pre-
venting overheating of the surfaces.

Circulation can be expedited by providing free and
unrestricted paths for the water so as to guide it in the
proper directions and by applying the most intense heat
at the proper point along the path of the water. The tem-
perature of the water which is subjected to the most intense

Blow off Valve

FIG. 207. — Elementary Types of Boilers.

heat is naturally raised and the water at that point becomes
less dense than in other parts of the boiler. The formation
of steam at such points also materially lessens the density.
As a result of this lowering of density the heated water
rises and the cooler water descends to take its place. The
more rapid this exchange can be made, the more steam can
be generated from a given amount of surface in a given
time and hence, other things equal, the better the boiler.

The elements of two common forms of boiler are shown
in Fig. 207, the arrows indicating the direction of the cir-
culation and its effect upon the delivery of steam and
of sediment.

148. Types of Boilers. In a book of this scope it would
be impossible to describe all the types of boilers at present

STEAM BOILEES

343

in use. The more important varieties have therefore been
chosen for description and illustration.

Two types of internally fired, tubular boilers more
accurately described as internally fired, upright or vertical,
fire-tube boilers are shown in Fig. 208. The furnace is

344

STEAM POWER

Tube Sheet

Steam Space
Exposed Tubes

Water Level

Water Column
and Try Cocks

Feed Water
Connection

Pressure
Gauge

Hand Holes
(Closed by
hand hole covers
i iu operation)

FIG. 209.— Large Internally Fired Tubular Boiler.

STEAM BOILERS 345

contained within the shell of the boiler and is almost com-
pletely surrounded with water. The heat radiated from
the hot fuel is thus almost entirely received by the water
of the boiler. The hot gases, rising from the fuel bed,
pass upward through the tubes and, after giving up part
of their heat to the surrounding metal, enter the smoke
box and pass directly to the stack.

Boilers of the type shown in Fig. 208, (a) and (6) are
called exposed-tube boilers, because the water level is carried
below the tops of the tubes. The tubes, therefore, extend
through the steam space and act as imperfect superheaters.
Boilers of the type shown in Fig. 208, (c) , in which the tubes
do not enter the steam space, but are entirely covered by
water, are called submerged-tube boilers.

Upright tubular boilers of the types shown in Fig. 208
are built by a number of manufacturers in sizes ranging
from about 4 boiler horse-power to about 50 boiler horse-
power. They are self contained, require no setting of any
kind, and are shipped completely erected. Such boilers
are very often mounted on trucks or skids and used to
generate steam for small hoisting and other forms of con-
tractors' engines. They are also used on steam fire engines.

The pressure carried in these small tubular boilers is
generally under 100 Ibs. per square inch, but they can be
built for higher pressures if desired.

In Fig. 209 is shown a larger type of internally fired
tubular boiler as made by the Bigelow Company for station-
ary use. These boilers are similar to those just described,
but are made only in large sizes, in this case, in sizes ranging
from 40 boiler horse-power to 200 boiler horse-power. The
exposed tubes generally give a superheat of about 25° F.

These large upright boilers can be built to operate with
a pressure as high as 200 Ibs. per square inch and because
of the small area covered by even the largest sizes, they
are particularly adapted to locations in which floor space
is limited.

346

STEAM POWER

The locomotive type of boiler is shown in Fig. 210. It is
an internally fired, horizontal, tubular or fire-tube, boiler.

Such boilers are seldom used for stationary purposes,
but are universally used on steam locomotives and, in the
smaller sizes, are often mounted on trucks or skids arid used
for semi-stationary purposes by contractors and others.
Boilers of this type are built in sizes ranging from 10 boiler
horse-power or less up to over 100 boiler horse-power for
general power purposes, while those used on the largest
locomotives generate over 2000 boiler horse-power.

The Continental type of boiler, named from the Con-
tinental Iron Works, is shown in Fig. 211. These boilers

Handholes

FIG. 210. — Locomotive Type of Boiler.

may be described as internally fired, return tubular, with
semi-external combustion chamber, this chamber being out-
side of the boiler shell proper but being built as an integral
part of the boiler and transportable therewith. Boilers of
this type are built in sizes ranging from about 75 boiler
horse-power to 300 or more.

The grates, furnace and ash spaces, and bridge wall are
all carried within circular, corrugated flues, one flue being
used in the smaller sizes and two in the larger. The corru-
gations serve the double purpose of strengthening the flue
and of exposing added heating surface to fire and hot gases.

The steam pipe shown just below the steam connection
at the top of the boiler is commonly used on boilers for the

STEAM BOILERS

noipsg dox i

347

348

STEAM POWER

purpose of preventing the escape of excessive quantities of
moisture with the steam.

These boilers are very compact in shape and are short
for their capacity, but they contain a great volume of water.
They possess the advantages of having a large steam space
and a very extended liberating surface over which the steam
separates from the water.

Uptake

Tubes

Fire Doors

Ash Pits

Man Hole-

FIG. 212.— Scotch Marine Type Boiler.

The Scotch marine type of boiler is shown in Fig. 212.
It has the same general construction as that just described
excepting that the combustion chamber is entirely enclosed
within the water space of the boiler. This chamber is
built up of flat plates and is held against collapse by numer-
ous stay bolts. Boilers of this type were until recently
the standard for marine practice, but they are now being
replaced in many instances by water-tube boilers of more
recent design.

STEAM BOILERS

349

Scotcn marine boilers are very economical in the use of
fuel, are good steamers, and are absolutely self contained.
They are built in numerous sizes, the smallest having shells
with diameters of about 6 ft., while the largest diameter
used is about 16 ft. The largest boilers have three and
four corrugated furnaces.

Two types of externally fired, return-tubular (or
" H.R.T.") boilers are shown in Figs. 213 and 214. The

Manhole

Stay RocUr ~

Front ^lU=^r 3 - — "^ShBuck Stave

FIG. 213.— Horizontal Return-tubular Boiler with " Full Flush Front."

only essential differences in these two types are in the forms
of setting and in the methods of suspending the boilers.
The shell is generally rigidly supported at the furnace end
and arrangements made to allow for movement of the other
end with changes of temperature.

These boilers can be built very cheaply and are therefore
widely used when their limitations do not prevent. It has
been found inadvisable to build them in sizes larger than
200 boiler horse-power or for pressures higher than 150
Ibs. per square inch, and they are generally used in smaller

350

STEAM POWER

STEAM BOILEES

351

352

STEAM POWEK

FIG. 216.— Forged
Header for Bab-
cock & Wilcox
Boiler.

sizes and with lower pressures. These limitations are set
by permissible thickness of metal immediately above the
fire, experience having shown that the
plates deteriorate rapidly at this point if
made too thick.

One form of Babcock & Wilcox water-
tube boiler is shown in Fig. 215. This
boiler is built up of sections consisting of
several tubes joined at the ends by headers,
and the sections are connected side by side
at each end to a long horizontal drum.
The ends of this drum are closed with
" dished " heads, thus doing away with
flat surfaces and the necessity for stays
within the drum.

A detail of the forged header is shown in
Fig. 216. It may be regarded as a long
box of rectangular section with opposite
walls pierced by circular holes, which has been so distorted
as to give it a wavy shape. The distortion brings the holes
into such positions that the tubes when expanded into these
holes are " staggered," that is,
do not lie one above the other.
The general principle in-
volved in the arrangement of
these sections or elements and
the resulting circulation are
shown in Fig. 217. The location
of the feed-water inlet and other
details are shown in Fig. 218.
It will be observed that the feed FIG. 217.— Elementary Babcock
water enters in such a direction * Wilcox Boiler, Showing

,,,.,. ,., Circulation,

and position that it is readily

picked up by the current of water circulating in the boiler,
carried toward the rear and down the rear header. During
this travel it is heated by contact with the hot water in

STEAM BOILERS

353

Hand hole opposite encL
of tube, closed by hand
hole cover when in
operation.

End of tube
expanded into
hole of header.

FIG. 218. — Details of Babcock & Wilcox Boiler Construction.

354

STEAM POWER

the boiler and most of its impurities are separated out and
settle in the mud drum at the bottom of the rear header.

The boiler is suspended by stirrups from beams carried
by the brickwork as shown in Fig. 215, the tube sections

STEAM BOILERS

355

simply hanging from the drum by the nipples at each end.
The various parts of the structure are thus free to expand
and contract independently as their temperatures change
and are not bound in any way by the brick setting.

The steam is collected from a perforated steam pipe near
the top of the steam space. The baffle shown in Fig. 218
prevents the steam which rises from the front header from
carrying the water bodily into the steam space and makes
the greater part of the water surface in the drum act as
separating surface.

The scale which accumulates inside of the tube is removed
by tools inserted through the hand holes in the front headers
opposite the ends of the tubes. One of these hand holes
and its cover are shown in section in Fig. 218. Soot .and
dust which accumulate on the outer surfaces of the tubes
are blown off periodically by a steam jet, the necessary
nozzle and hose being "in-
serted through the tall and
narrow side cleaning doors
shown in Fig. 215 opposite
each " pass."

A section of the Heine
water-tube boiler is shown
in Fig. 219. This boiler con-
sists of a slightly inclined
drum with dished heads, two
sheet-steel headers and nu-
merous tubes connecting
these headers. The shape
of the header is shown in
Fig. 220, which indicates the
positions occupied by the tubes and the way in which
the header is joined to the drum.

The products of combustion are generally made to pass
along the tubes by the longitudinal baffles shown, instead
of across the tubes as in the boiler last described.

<£>OOOOOOOOOOO

FIG. 220. — Front End Elevation,
Heine Boiler.

356 STEAM POWER

The mud drum in this type is located within the boiler
and consists of a sheet-steel box supported a few inches
above the bottom of the drum. The feed water enters at
the front end of this drum and gradually spreads out as it
is heated by the surrounding water. The greater part of
the impurities settles to the bottom and is blown off period-
ically. The warmed water rises and flows out of an opening
in the top of the box at the front end and there joins the
circulation of the boiler, traveling toward the rear, down
the rear header and up the tubes to the front header.

The interior of the tubes is cleaned of scale through
hand holes just as in the last boiler. The external surfaces
are freed of soot and dust by means of a steam jet which
is introduced through the stay bolts in the headers, these
bolts being made hollow for this purpose. Since it is not
necessary to use doors in the side walls for cleaning in this
type, Heine boilers are often set up in batteries of three or
more, each interior side wall serving as the side wall of two
settings. In the case of the boiler last described the neces-
sity for side cleaning doors makes it impossible to join more
than two boilers in this way.

The Heine boiler is supported by standing the front
and rear headers upon the brickwork of the setting and
it can therefore expand freely in all directions.

A section of the Sterling water-tube boiler is shown in
Fig. 221. This boiler consists of three upper horizontal
drums connected by short curved tubes and connected to
a single lower horizontal drum by means of long tubes which
are curved near the ends. The curves of all tubes are so
made that the tubes enter the drum surfaces radially, thus
giving a simple joint which is readily made tight by expand-
ing the tube into the sheet.

The feed water is introduced into the upper rear drum,
and is gradually heated and partly purified as it passes
downward to the lower drum, in which the greater part of
the material precipitated from the water is caught and

STEAM BOILEES

357

stored until blown off. From the lower drum the water
is supposed to pass upward through the front bank of tubes,
the steam formed passing to the central drum through the
upper set of short curved tubes, and the water which is
not evaporated passing to the central drum through the

Feed Wator Inlet

Steam Connection.

Smolc
Couucc (ion

Damper

Bottom Blow-off

FIG. 221

-Section of Sterling Boiler.

lower set of curved tubes. This water passes from the upper
central drum to the lower and returns through the front
bank of tubes. Any steam formed in the rear bank of tubes
or in the rear drum passes to the central drum through the
short curved tubes connecting the steam spaces.

The entire boiling vessel is hung from a frame of struc-
tural steel by means of the upper drums, so that the lower

358 STEAM POWER

drum hangs practically free on the tubes. Independent
expansion of all the members is insured by this method
of suspension and by the curvature of the tubes, which per-
mits each one of them to bend to the extent necessary to
equalize any strains caused by changing temperatures.

The interiors of the tubes are cleaned by means of tools
lowered from inside the upper drums and the exterior
surfaces are blown off by steam jets introduced through
doors in the brickwork of the setting.

The Wickes vertical water-tube boiler is shown in
section in Fig. 222. It consists of an upper and lower cir-
cular drum, connected by straight tubes expanded into the
lower and upper heads of the drums respectively. A
vertical baffle placed in the center of the bank of tubes
gives an upward path to the products of combustion when
passing over the front tubes and a downward path when
passing over the rear tubes.

The feed water is generally introduced at the rear of
the upper drum, the circulation being downward in the rear
tubes and upward in the front tubes.

The interior surfaces of the tubes are cleaned by tools
lowered into them by a man standing within the upper
drum, which is made high enough to make this possible.
The external surfaces are cleaned by steam jets inserted
through doors in the brickwork.

The entire boiler is supported on brackets riveted to
the lower or mud drum and is free to expand in all directions,
the brickwork simply enclosing but not confining it.

149. Boiler Rating. Practically all apparatus which is
connected with the development of power is given a horse-
power rating. In some cases such a method of rating is
convenient and simple, in others it is inconvenient, irra-
tional and complicated. The term horse-power, when used
as a measure of work or power, means very definitely the
equivalent of 33,000 ft.-lbs. per minute. When, however,
a certain number of horse-power is used as the rating of a

STEAM BOILERS

359

team Connection

*^Feed Water
«* Connection

ttom
Blow-off

FIG. 222.— Wickes Vertical BoUer.

360 STEAM POWER

particular piece of apparatus, it generally means that that
piece of apparatus, when working at about its best effi-
ciency, can do what is necessary to make available the
stated number of horse-power in the plant of which it forms
a part.

Thus a boiler rated at a certain horse-power was origi-
nally supposed to be able to supply the amount of steam
required by an average engine developing that quantity
of power and to do this when working at its best efficiency.
The water rates of engines are, however, so different that
there is no real connection between boiler horse-power and
engine horse-power, and it is best to consider the boiler
horse-power as a perfectly arbitrary unit defined in a certain
way.

The American Society of Mechanical Engineers has
defined the boiler horse-power as the equivalent of the evapora-
tion 0/34.5 /6s. of water per hour from and at 212° F. This
means the conversion per hour of 34.5 Ibs. of water at 212° F.
into steam at the same temperature and therefore at atmos-
pheric pressure.

Each pound of steam generated under these conditions
requires the expenditure of the latent heat of vaporization
rt atmospheric pressure, which is equal to 970.4 B.t.u.
according to the latest steam tables. The older tables gave
965.7. This quantity of heat is known as a Unit of Evapora-
tion and is abbreviated U.E. The boiler horse-power is,
therefore, the equivalent of 34.5 U.E. per hour or 34.5X970.4
= 33,479 B.t.u. per hour.

As practically no power-plant boilers receive their feed
water at a temperature of 212° F. and convert it into steam
at the same temperature, it is necessary to convert the
weight actually evaporated to what it would have been
from and at 212° F. and then to divide this figure by 34.5
in order to find the boiler horse-power developed.

The number of pounds which would have been evaporated
from and at 212° F. if the same amount of heat had been

STEAM BOILERS 361

transmitted is known as the equivalent evaporation, or as
the equivalent weight of water evaporated into dry steam from
and at 212° F.

The method of obtaining the equivalent evaporation
has been defined by the American Society of Mechanical
Engineers. The heat given to each pound of dry saturated
steam produced is to be determined; this is to be multiplied
by the total weight of dry saturated steam generated per
hour, and the product is to be divided by the latent heat
of vaporization at 212° F. Thus, for a boiler receiving
its feed water at some temperature tf above 32° F., the
water contains a quantity of heat equal to <// B.t.u. per
pound, g/ being found in the steam table opposite the tem-
perature tf. Each pound of dry saturated steam leaving
the boiler carries with it an amount of heat equal to X for
the existing temperature. The heat supplied each pound
in the boiler must therefore be X— qf and, for W pounds
per hour, the heat supplied would be W(\—q/). The
equivalent evaporation is then given by

Equiv. evap. = TF(=~n Ibs. per hour. . (101)

This expression may be regarded as consisting of two
factors, the weight of dry steam generated per hour, and
a fraction which will always have the same value for a given
combination of pressure and feed-water temperature. This
fraction is called the factor of evaporation, and it is cus-
tomary to tabulate the various values of the factor of
evaporation for different common combinations of pressure
and feed-water temperature.

It should be noted that the equivalent evaporation as
defined above gives the boiler no credit for heat given to
water which leaves the boiler as water, nor does it give
credit for any superheating. The former may be justified
by saying that the boiler, as a commercial piece of apparatus,
is not intended to supply hot water; but many commercial

362 STEAM POWEE

boilers are expected to supply superheated steam and should
be given credit for heat used in that way.

Returning now to the boiler horse-power, its value can
obviously be found for any given boiler by dividing the
equivalent evaporation per hour by the number 34.5.

Boilers are supposed to be so rated that they will
develop their rated horse-power when operating at about
their best efficiency and will do it with moderate draft
and reasonably good firing with average fuel. Experience
has shown that for most boilers the best efficiency is ob-
tained when an equivalent evaporation of from 3 to 3.5 Ibs.
of water occurs per square foot of heating surface. The
heating surface is generally taken as the total surface in
contact with hot gases excepting in the case of tubes. The
outer surfaces of tubes are generally counted even if they
be in contact with the water. An equivalent evaporation
of 3 to 3.5 Ibs. per square foot would call for a heating
surface of from 12 to 10 sq.ft. per boiler horse-power.

Most water-tube boilers are given 10 sq.ft. of heating
surface per rated boiler horse-power, and most return-
tubular boilers are supplied with 11 to 12 sq.ft. Scotch
marine boilers are generally designed on a basis of about
8 sq.ft. per rated boiler horse-power.

The quantity of water which can be evaporated per
square foot seems to depend to a great extent upon the rate
at which hot gases can be passed over the heating surface,
and experiments have shown that from five to eight times
the ordinary rates of evaporation can be attained if suf-
ficient fuel can be burned. As the rate of evaporation per
square foot is increased above the commonly accepted
value, the efficiency decreases, but the decrease is generally
small for a considerable increase in rate of evaporation.
Most power-plant boilers can give from 150 to 200 per cent
of their normal rating, and some are now being installed to
operate for long periods at about 200 per cent of what
would be considered a normal rating.

STEAM BOILERS 363

150. Boiler Efficiencies. There are a great many pos-
sible efficiencies which may be considered in connection
with boiler tests. The two most commonly used are defined
by the A.S.M.E., and are:

1. Efficiency of the boiler

Heat absorbed per pound of combustible burned
Calorific value of 1 Ib. of combustible

2. Efficiency of boiler and grate

_ Heat absorbed per pound fuel
Calorific value of 1 Ib. of fuel*

The names used are not very well chosen, and it is better to
call the first the efficiency based on combustible and the
second the efficiency based on coal. The weight of com-
bustible burned is calculated by subtracting from the coal
fired the total weight of moisture and the total weight of
refuse in the ash pit.

The heat absorbed is by definition the heat absorbed
by the dry steam made by the boiler, but it seems probable
that this will also be modified in the near future as suggested
in a preceding paragraph.

It is also possible to determine the efficiency of the grate,
of the furnace, and of the boiling vessel, and this is some-
times done.

The best commercial operating values for the efficiency
of the boiler as a whole, that is, the boiler and grate on the
basis of total fuel fired, are about 75 per cent for good
qualities of coal and 80 per cent for oil, but such values
are generally obtained only in well-equipped plants operat-
ing on comparatively constant loads. Average commercial
values generally range from 60 to 70 per cent on a yearly
basis in well-equipped plants which are carefully operated,
and many boiler plants are operated at an efficiency of 50
per cent and less.

The pounds of water evaporated per pound of coal

364

STEAM POWER

fired generally ranges between 6 and 10, and the equiva-
lent evaporation per pound of combustible burned will
generally fall between 8 and 12 pounds.

151. Effects of Soot and Scale. The flue gases in real
boilers are seldom clean mixtures of the products of com-
bustion and nitrogen, as theory would indicate. They
always contain more or less soot and unburned hydro-
carbons, as well as some finely powdered ash and fuel. With
strong draught, very large particles of ash and fuel may be
carried by the flue gases.

These materials are partly carried up the stack by the
gases and partly deposited on the heating surfaces of the
boiler. Such deposits decrease the conductivity of the
heating surfaces, and if the deposits are heavy the loss may
be very great. The results of one investigation on the
effect of soot are given in Table XV, the values being
taken from an article published in the Proceedings of the
Institute of Marine Engineers for the year 1908. These
values are probably too high, particularly for the thicker
deposits, but they serve to bring out the fact that a very
appreciable loss does occur from the presence of such
deposits.

TABLE XV

EFFECT OF SOOT DEPOSITS ON BOILER HEATING SURFACES

Thickness of Deposit
in Inches.

Loss of Conductivity
in Per cent.

0.0

9.5

26.2

45.2

69.0

The effect of soot deposits in decreasing the efficiency
of boilers was used for a long time as a basis for argument
in favor of certain types of boilers in which the heating
surfaces were so shaped and located that such deposits

STEAM BOILERS 365

formed to a minimum degree and against other types
less favorably designed from this point of view. Prac-
tically, however, the removal of such deposits by means of
steam jets applied at regular intervals is so simple that this
consideration need be given little weight in the selection
of a boiler. Provision should always be made, however,
for the easy use of the jets for cleaning purposes.

152. Scale. Practically all water available for boiler feed
contains various salts in solution and it often contains solid
matter in suspension as well. This material is all deposited
within the boiler as the water is heated and converted into
steam. There is thus a gradual collection within the boiler
of all the solid material brought in by the water.

In well-designed boilers the greater part of such deposits
is carried to a part of the boiler in which the metallic sur-
faces are not exposed to high temperature gases, as, for
instance, the mud drums in water-tube boilers. It can then
be drawn off periodically in the form of a thin mud sus-
pended in water. In practically all boilers, however, some
of the solid material will be carried to the heating surfaces
exposed to high temperature gases and deposited there.
Under the action of heat, the mud-like material gradually
changes until, in many instances, it forms a very hard,
stone-like coating on the heating surface. This is known
as boiler scale.

Such deposits may cause two kinds of trouble: They
may decrease the conductivity of the heating surfaces and
thus decrease the efficiency of the boiler; and, because of
their location on the water side of the metal, they may per-
mit the hot gases to overheat that metal, thus weakening
it. Such overheated metal often " bags " under the high
internal pressure and may eventually give way with disas-
trous results. The mechanical structure of the scale seems
to be the determining factor; scales which are easily pene-
trated by water have little effect, while those which are very
dense and non-permeable may cause serious trouble.

366 STEAM POWER

Boilers should be blown down periodically to keep them
as free as possible of scale-forming material, and they should
be so constructed that scale which has been formed can be
removed easily. Very efficient tools have been developed
for removing scale from the interior and exterior surfaces
of tubes, so that boilers using tubular heating surfaces are
readily cleaned of scale.

153. Scale Prevention. Much of the solid material
carried by water is deposited when the water is heated to
a temperature of from 150° to 200° F., so that heating feed
water before it is admitted to the boiler is at least a partial
preventive in most cases.

Nearly all of the salts which are soluble in hot water
and therefore are not deposited when the feed is heated,
can be made to form insoluble compounds by the addition
of comparatively cheap chemicals. By the addition of such
chemicals in the feed-water heaters, or in other apparatus
specially designed for that purpose, the greater part of the
solid content of the water can be precipitated before it is
admitted to the boiler.

There are a great many " boiler compounds " on the
market which are intended to be mixed with the water as
it is fed to the boiler and are supposed to prevent the for-
mation of scale on the heating surfaces. All they can
possibly do is to change the chemical composition of the
solids; they cannot prevent the deposit of these solids within
the boiler. They are therefore, at best, only an imperfect
remedy.

154. Superheaters. Many boiler plants are now ar-
ranged to supply steam superheated 25 to 200 degrees Fahr.
It was shown in an earlier chapter that the use of super-
heated steam greatly improves the economy of reciprocating
engines and turbines, and there are also other advantages
which accrue from its use.

Superheaters are of two kinds — separately fired and
built-in superheaters. The separately fired superheaters are

STEAM BOILERS 367

enclosed in a brick setting fitted with grate and furnace
similar to that of an ordinary boiler. The built-in super-
heaters are installed within the boiler setting so that the
products of combustion pass over them in flowing through
the boiler.

In either type the steam passes through the superheater
on its way from the boilers to the engines. In the case
of separately fired superheaters, the temperature of the
superheated steam is controlled by regulation of the fire
on the grate of the superheater, but in the built-in type
regulation in this way is practically impossible, as the fire
under the boiler must be controlled to suit the demand for
steam. The control of such superheaters is therefore effected
either by locating them in such a position that the natural
variation in the temperature of the gases reaching them
gives an approximate regulation, or they are installed in
a separate chamber and hot gases passed over them in such
proportions as necessary to give the required temperature.

The Babcock & Wilcox superheater as applied to the
boiler of the same make is shown in Fig. 223. The steam
collected in the dry pipe within the drum passes downward
to the upper manifold of the superheater and from there
it flows through the U-shaped tubes into the lower manifold.
From the lower manifold it flows through the superheater
stop valve to the engine or turbine.

The superheater is so located that the hot gases pass
over it between the first and second passes and there is no
way of shutting off these gases. Provision, as shown in
the illustration, is therefore made for flooding the super-
heater during starting, or when superheated steam is not
desired. When flooded it becomes heating surface similar
to that of the tubes below, the steam made passing into
the drum through the dry pipe.

The Heine superheater as applied to a Heine boiler is
shown in Fig. 224. It consists of a sheet-metal header or
box into which U-shaped tubes are expanded. The steam

368

STEAM POWER

Safety Val

Drain
Valve

Front

Gases
from Furnaces

FIG. 223.

-Superheated Steam from
Superheater

Saturated

Steam to

Superheate

Dam
Cont

FIG. 224. — Heine Superheater.

STEAM BOILERS

369

enters the bottom of the header and is guided by dia-
phragms in such a way that it passes through the lower set
of U-tubes, returns to the header, passes through the upper
set of tubes, and then leaves the superheater at the top.

FIG. 225.— H.R.T. Boiler and Foster Superheater.

This apparatus is installed in a brick chamber built into
the boiler setting and connected with the furnace by a flue
(not shown) in the brick side wall. A damper controls the
flow of hot gases to this chamber and the degree of superheat
is controlled by the position of this damper.

FIG. 226. — Element of Foster Superheater.

The Foster superheater is shown installed in the setting
of an H.R.T. boiler in Fig. 225 and the details of the con-
struction of one element are shown in Fig. 226. The core
is used to spread the steam in a thin stream, thus bringing
it into better contact with the heating surface. The fins

370 STEAM POWER

on the exterior of the element are used for the purpose of
getting a more extended metallic surface in contact with
the hot gases.

155. Draft Apparatus. Attention was called in a pre-
ceding paragraph to the fact that there must be a difference
of pressure between the spaces below and above the fuel
bed in order to cause the necessary air to flow through the
bed. This difference of pressure is called the draft.

As a matter of fact, a slight difference of pressure is
required to cause the flow of gases through any part of the
boiler and the drop in pressure through the fuel bed is only
part of the total draft required.

The draft may be created in two distinctly different
ways. It may be caused by a chimney or stack, and is
then known as natural draft, or it may be produced by fans
or blowers, in which case it is called mechanical draft.

(a) Chimneys or Stacks. Stacks are practically always
used in small plants because of the simplicity resulting from
their use and because the interest on the investment com-
pares favorably with interest on investment plus cost of
operation for mechanical draft. In large plants fitted with
some types of mechanical stokers, or where fuel is to be
burned at a high rate, or where the flue gases are to be used
for heating feed water, mechanical draft is generally installed.
A stack of some sort is necessary even though mechanical
draft be used, because the products of combustion must be
discharged at a sufficient elevation to prevent their being
a public nuisance.

A chimney serves to carry away the hot products of
combustion arid when in operation is filled with a column of
gases with higher average temperature than that of the
surrounding air. As a result the density of gases within
the stack is less than the density of the outer air and the
gas pressure at the bottom of the structure is less inside
the stack than it is outside. If an opening is made at this
point, the external air will therefore flow in. By arranging

STEAM BOILERS

371

the apparatus as shown in Fig. 227, the temperature
of the air flowing into the bottom of the stack is raised
as it passes through the furnace and the flow is thus made
continuous.

The height of the chimney determines the draft created
by it with flue gases of a given temperature, and, with any
given height, the area determines the quantity of gas which

Atmospheric Pressure
i mount caused by
low dtnbity of hot gases
in fatatk.

FIG. 227. — Diagrammatic Arrangement of Stack.

can be carried off in a given time. The proportions of
chimneys can be determined from rational formulas based
on theoretical considerations, but it is necessary to assume
values for a number of constants and a proper choice de-
pends largely upon experience.

As a result, all but the more important chimneys are
generally designed on an empirical basis and many formulas
have been developed for this purpose. One of the most
common methods of design is to choose the height in accord-

372 STEAM POWER

ance with the values given in Table XVI, and then to
determine the sectional area according to an empirical
assumption or formula.

TABLE XVI

COMMON HEIGHTS OF CHIMNEYS

{Applicable to plants smaller than about 700 H.P. Larger installations should
have stacks of from 150 to 175 feet in height unless local conditions call for
greater height.)

Character of Fuel.

Height above Grate in Feet.

Free-burning bituminous . .

Anthracite, medium and large sizes
Slow-burning bituminous

100
120

Anthracite pea size.

130

Anthracite, buckwheat sizes

150

Thus, some designers simply assume the sectional area
at the top of the stack equal to about one-ninth of the grate
area for anthracite coal and equal to about one-seventh of
the grate area for bituminous coal. Others use a formula
developed by William Kent, which is based upon the
assumption that the stack should be large enough to carry
away all the gases resulting from the combustion of 5 Ibs.
of coal per rated boiler horse-power per hour. This formula
gives the boiler horse-power which the stack can serve
and is

HJ\ = 3.33(A-0.6VZ)V#, . . . (102)
in which

H. P. = Rated boiler horse-power;

A = Internal sectional area in feet of circular or square

chimney;
H = Height above grate in feet.

(6) Mechanical Draft. Fans can be so used as to force
air into the ash pit, that is, to raise the pressure on the

STEAM BOILERS 373

entering side of the fire. In such cases the equipment is
said to give forced draft. Or fans may be installed at the
discharge end of the flues and may "draw " the gases
through the boiler by lowering the pressure within to a value
below that of the external atmosphere. Such an instal-
lation is said to give induced draft.

Forced draft suffers from the disadvantage that the
pressure within the furnace is greater than atmospheric and
hot gases may therefore be blown out when the fire
door is opened. On the other hand, the fan handles only
cool air instead of hot products of combustion as in
the case of induced draft and its useful life is therefore
much longer. Forced draft is much more common than
induced draft.

Several arrangements giving balanced draft have been
developed. With such apparatus a pressure equal to atmos-
pheric is maintained above the fuel bed and no hot gases
are blown out through the firing door.

PROBLEMS

1. The equivalent evaporation of a boiler during a certain
test was 3450 Ibs. per hour. What boiler horse-power was de-
veloped?

2. A water-tube boiler with 5000 sq.ft. of heating surface
and rated in the ordinary way gave an equivalent evaporation
of 25,875 Ibs. per hour. At what per cent of rating was the
boiler operating?

3. A certain boiler produced 3500 Ibs. of dry steam in one
hour from feed water at a temperature of 50° F. The steam
pressure was 200 Ibs. per square inch gauge. What was the
equivalent evaporation?

4. A boiler receiving water at a temperature of 250° F. con-
verts it into superheated steam at a pressure of 210 Ibs. per square
inch gauge and a temperature of 580° F. The boiler produces
26,000 Ibs. of steam per hour. What is the equivalent evaporation
if the boiler is given credit for all the heat given the material
passing through it? What boiler horse-power is developed!

5. A boiler produces 7.5 Ibs. of dry steam per pound of coal
fired. The feed-water temperature is 80° F. and the steam pres-

374 STEAM POWER

sure is 125 Ibs. per square inch absolute. What is the equivalent
evaporation per pound of coal?

6. A boiler is supplied with coal which has a calorific value
of 13,520 B.t.u. per pound. It produces 8 Ibs. of dry saturated
steam at a pressure of 150 Ibs. per square inch gauge per pound
of coal. The feed-water temperature is 70° F. What is the
efficiency of the outfit?

CHAPTER XVIII
RECOVERY OF WASTE HEAT

156. Waste Heat in Steam Plant. There are two great
heat wastes in the steam plant — the waste in the hot gases
going up the stack and the waste in exhaust steam. The
magnitude of the stack loss can best be appreciated by
determining an approximate value for assumed conditions.
For this purpose assume the fuel to be pure carbon, the
excess coefficient 1.5, average atmospheric temperature 60°
F., average stack temperature 600° F., and no moisture in
the air. The specific heat of the flue gases may be taken
as constant and equal to 0.24.

With an excess coefficient of 1.5, the total weight of
flue gas per pound of carbon burned would be about 18.4
Ibs. and the heat carried up the stack figured above room
temperature would be

Stack loss = 18.4X0.24 (600-60).

= 2380 B.t.u. per pound of C burned (approx.)

With a calorific value of 14,600 B.t.u. per pound of carbon
this loss would be equivalent to a little over 16 per cent
of the total heat in the fuel.

It would be more correct to use the temperature of
the steam in the boiler instead of room temperature, because
the lowest temperature theoretically attainable by gases
passing through a boiler would be equal to that of the steam
and water on the other side of the heating surface. Under
ordinary conditions of operation, this method of figuring
would give a theoretically avoidable stack loss equal to
about 50 per cent of the figure obtained above.

375

376 STEAM POWER

The magnitude of the exhaust loss can be similarly
approximated. Assume for this purpose an engine receiving
dry saturated steam at 115 Ibs. absolute per square inch and
exhausting it with a quality of 90 per cent at a pressure of
15 Ibs. absolute per square inch.

The heat above 32° in the entering steam is 1188.8 B.t.u.
per pound and the heat exhausted per pound is 1053.7.
The heat in the exhaust represents therefore about 89 per
cent of all the heat supplied when calculations are made
above a temperature of 32° F. If a feed-water tempera-
ture of 60° be assumed and heat quantities be figured
above that datum the results are practically the same.

There are always numerous pieces of auxiliary apparatus
in steam plants such as boiler-feed pumps, circulating pumps,
vacuum pumps, etc. These are often steam driven and are
generally very uneconomical in the use of heat, so that they
throw away in their exhaust steam large quantities of heat
originally transferred from fuel to water and steam in the
boiler.

157. Utilization of Exhaust for Heating Buildings. It
often happens that steam-power plants are located within
or in the neighborhood of buildings requiring artificial heat
during part of the year. In such cases the exhaust steam
from main and auxiliary engines can generally be advan-
tageously used for this purpose. Under particularly favor-
able circumstances, the weight of steam required by the
plant may equal approximately that required for heating,
and the greater part of the exhaust could then be turned
directly into the heating system.

The engines in plants of this character may be regarded
as reducing valves for the heating system, receiving steam
at high pressure and reducing the pressure to the value best
adapted to the heating system installed. If the com-
paratively small losses arising from radiation from the
engine, from friction and from the presence of hot water
in the exhaust be neglected, all heat received by the engine

RECOVERY OF WASTE HEAT 377

and not turned into useful mechanical energy is made use
of in the heating system. The engine may therefore be
very uneconomical in the use of steam and still not cause
a waste of fuel, provided always that the heating system
can absorb all heat exhausted.

Since the demands of a heating system vary from day
to day and since there is generally no demand for heat
during several months of each year, it follows that a high
degree of skill is necessary in choosing the character of the
apparatus installed. A compromise is generally made
between the cheap and uneconomical engine allowable during
the coldest months and the more expensive and more
efficient engine desirable when no heating is to be done.

There are other cases of somewhat similar character.
In many industries use can be made of exhaust steam for
the heating of evaporating pans, dye vats, kilns and other
apparatus. Steam plants of an uneconomical character may
be very economical financially in connection with such
industries if all or nearly all of the heat in the exhaust can
be utilized industrially.

158. Feed-water Heating. An examination of the steam
table will show that the total heat above 32° F. per pound
of steam varies between 1180 and 1200 B.t.u. for such pres-
sures as are commonly used in boilers. The average tem-
perature of water as it occurs on the surface of the earth
is probably somewhere in the neighborhood of 60°, so that
the heat above 32° per pound would roughly average 27
B.t.u. A boiler receiving water at 60° and converting it
into steam at any of the ordinary pressures must therefore
supply over 1100 B.t.u. per pound of water.

This immediately suggests a use for heat in exhaust
steam. Steam exhausted into very low vacuums has a
temperature only 10° to 30° higher than the assumed average
natural feed temperature, but steam exhausted at atmos-
pheric pressure has a temperature of 212° F. and could
therefore impart large quantities of heat to water at 60° F.

378 STEAM POWER

Since the boiler must supply over 1100 B.t.u. per pound
of steam made, raising the feed temperature about 11° or
12° should effect a saving of about 1 per cent in fuel con-
sumption. By raising the temperature from 60° to 212°
there should therefore result a saving of approximately 13
to 14 per cent.

Other advantages which would accrue from this pre-
liminary heating of the feed water would be (1) the deposit,
outside of the boiler, of a large amount of the solid matter
carried by the water, (2) the use of fewer or smaller boilers,
and (3) the reduction of the strains which occur in the metal
of some designs when very cold feed water is used.

Exhaust steam feed-water heaters are divided into two
types, open and closed heaters. In open heaters the steam
and feed water are brought into intimate contact in the
form of jets, sheets and sprays within a vessel of appropriate
size and shape. They are often called contact heaters.
When the exhaust steam comes from reciprocating engines
it always carries in suspension some of the oil used for
lubricating the engine cylinders. If allowed to enter the
heater, this oil would mix with the feed water and eventually
reach the boilers, where it might cause serious damage by
depositing upon heating surfaces exposed to the fire or to
very hot gases. Such heaters are therefore always fitted
with oil or grease extractors when used with reciprocating
units. When receiving the exhaust from turbines, oil ex-
tractors are not necessary, as no lubricant is used within
the steam spaces of such units.

Closed heaters consist of tubes or coils enclosed within
a metal vessel. One medium passes through the tubes
and the other over their outer surfaces. Such heaters are
therefore often called non-contact heaters.

As oil is a poor conductor of heat, the exhaust steam
from reciprocating units should be passed through an oil
extractor before entering a closed heater in order that the
heating surfaces may be used to the best advantage.

RECOVERY OF WASTE HEAT 379

Exhaust steam feed-water heaters are often divided into
primary and secondary heaters. This distinction has nothing
to do with structure, being based entirely on position and
temperature. Thus there may be available exhaust steam
at a pressure below atmospheric, as from condensing main
units, and exhaust steam at atmospheric pressure from non-
condensing auxiliaries. The lower pressure steam could be
used to heat the feed water in a primary heater and the
higher pressure steam could then raise its temperature still
further in a second or secondary heater.

The other great waste, that in the stack gases, can also
be partly eliminated by using some of it to heat the feed
water. As the highest steam temperature ordinarily avail-
able in the exhaust system is about 212° F., and as the
products of combustion leaving the boilers generally have
temperatures in the neighborhood of 600° to 700° F., it
is evident that on a basis of temperature the hot gases
have a decided advantage as a heating medium. On the
other hand, the specific heat of the hot gases is low, while
exhaust steam can give up all of its latent heat with no
change in temperature, so that on a basis of heat avail-
able for transmission to the water, the steam has the
advantage.

The waste heat in the flue gases is used for feed- water
heating in devices known as economizers. These generally
consist of groups of tubes, joined at their ends by headers
and standing vertically within sheet-metal flues leading
from the gas passages of the boilers to the chimney. The
water to be heated is pumped through the tubes on its way
to the boilers and the hot gases flow over the tubes on their
way to the chimney. Mechanically operated scrapers are
arranged to travel up and down the tubes at intervals and
keep their external surfaces free of soot and dust, which
would seriously reduce their ability to transmit heat.

The feed water supplied the economizer is generally
first heated in an exhaust steam heater and arrives at the

380 STEAM POWER

economizer with temperatures between about 120° and
200° F., depending upon the kind of heaters used and upon
the relative quantities of exhaust steam and feed water.
The economizer discharges the water to the boiler at tem-
peratures which generally run from about 210° to 300° F.,
depending upon the amount of preliminary heating, the
extent of economizer surface and a number of other variables.
The gases which have passed through an economizer
often have temperatures as low as 250 to 350° F., which is
generally too low a value to give good chimney draft.
Plants making extensive use of economizers are therefore
generally fitted with some form of mechanical draft.

PROBLEMS

1. Determine the heat lost in the chimney gases per pound
of coal in a plant operating under the following conditions, and
express the loss as a percentage of the heat value of the coal.
The coal has a calorific value of 14,000 B.t.u. per pound; the
temperature of the gases leaving the boiler is 570° F.; 20 Ibs. of
gas result from each pound of coal burned; the mean value of
the specific heat of the gases is 0.245; and the temperature of
the air entering the furnace is 75° F.

2. Determine the quantity of heat which could be obtained
from the gases of Prob. 1 by using an economizer to reduce their
temperature to 250° F. What percentage of the heat value of
a pound of coal does this saving represent?

3. The boilers of a certain plant produce 100,000 pounds of
steam per hour when the plant is operating at full load. The
steam-driven auxiliaries consume 10% of this steam. Steam is
generated at a pressure of 175 Ibs. per square inch gauge, and
is superheated 150° F. The main units operate condensing and
the condensate leaves the condensers at a temperature of 75° F.
The auxiliaries operate non-condensing and exhaust their steam
at atmospheric pressure and with a quality of 92%. The coal
used has a calorific value of 13,850 B.t.u. The boiler efficiency
/Heat given water and steamX , „„...

I ° . _,. ,. I -jo / ^^Y

\ Heat in fuel supplied /

(a) Determine the amount of coal which would have to be
burned per hour if the steam exhausted from the auxiliaries were
thrown away and make-up water at a temperature of 50° F. were

RECOVERY OF WASTE HEAT 381

used in its place. The condensate from the condensers of the
main unit is assumed to be returned to the boiler after being
mixed with the make-up water.

(6) Determine the amount of coal which would have to be
burned per hour if the auxiliary exhaust were used to heat the con-
densate from the main units in an open heater and if the operation
of the plant were so perfect that no make-up water had to be
added.

CHAPTER XIX
BOILER-FEED PUMPS AND OTHER AUXILIARIES

159. Boiler-feed Pumps. The pumps used for forcing
the feed water into boilers may be of reciprocating or
centrifugal construction and may be driven by reciprocating
steam cylinders, by small steam turbines or by electric
motors.

Steam-driven pumps are very wasteful, often using over
100 Ibs. of steam per horse-power hour. It would therefore
seem more economical!^ use motor-driven pumps in electric-
power stations, as the large power units will generate electric
power with a consumption of from 10 to 25 Ibs. of steam per
horse-power hour and the motor efficiency will generally be
over 80 per cent. There is, however, another point which
must be considered. The exhaust steam from small engines
operating boiler-feed pumps can be used for heating the
feed water as described in the last chapter, and thus the poor
economy of these units is of little significance ; practically
all heat exhausted can be returned to the boiler in the
boiler feed if desirable. As a result of this considera-
tion, coupled with others of less importance, nearly all
boiler-feed pumps and other similar auxiliaries are steam
driven unless there are so many that there would be
more exhaust steam than could be absorbed by the feed
water.

There is at present a marked tendency toward the use
of turbine-driven, centrifugal pumps for boiler feeding, in
place of those driven by reciprocating steam units. The
turbine type has several advantages, the more important
being:

382

BOILER-FEED PUMPS AND OTHER AUXILIARIES 383

(1) No oil in exhaust steam, so that latter is well adapted
to use in all forms of feed- water heaters;

(2) Higher speed because of continuous flow of water
and continuous rotation of mechanical parts, thus making
possible great decrease in size for a given amount of work,
and

(3) Better pump characteristics for this sort of work.

The Duplex Steam Pump. The great majority of re-
ciprocating steam pumps used for boiler-feed purposes are
of the duplex pattern, one design of which is shown in Figs.

Steam End

Water End

FIG. 228.— Duplex Steam Pump.

228 and 229. Two steam cylinders are arranged side by
side, their piston rods extending into similarly arranged
water cylinders and carrying water plungers or pistons as
shown in Fig. 229. As there is no rotating shaft in a pump
of this kind, the steam valves cannot be operated by eccen-
trics as is common with steam engines. For the purpose
of operating these valves, bell cranks, pivoted near the
center of length of the pump, are provided. These are
arranged so that the long arm of one bell crank engages a
collar on the piston rod of one steam cylinder and the short
arm operates the valve gear of the other steam cylinder.
The motion of the valve of one cylinder is therefore derived

384

STEAM POWER

from the piston motion of the other cylinder. The steam
pistons are practically 180° out of phase, one moving out
while the other moves in.

Practically no expansion of the steam is obtained in the
cylinders of pumps of this type. They operate on the
rectangular cycle described in an earlier chapter and are
correspondingly wasteful in their use of steam.

Slide Valves

Steam End

Water End

FIG. 229.— Duplex Steam Pump.

160. The Steam Injector. On steam locomotives and
in other portable steam plants, as well as in many small
stationary plants, a device known as a steam injector is used,
instead of a pump, for forcing feed water into the boiler.
A simple form of steam injector is shown semi-diagram-
matically in Fig. 230.

Steam from the boiler flows through the steam nozzle
and expands from boiler pressure to a very low pressure,
thus acquiring a high velocity at the expense of the heat

BOILER-FEED PUMPS AND OTHER AUXILIARIES 385

energy which it brings from the boiler. At the end of the
nozzle it mixes with water and imparts to that water some

of its kinetic energy, so that the mixture moves into the
small end of the delivery tube with a high velocity. By
the time it has reached that point, practically all the steam
has been condensed, and, as the sectional area of the delivery

386 STEAM POWER

tube increases, the velocity of the liquid decreases with a
corresponding increase in pressure according to Bernoulli's
theorem. In properly designed apparatus, the resultant
pressure is great enough to force the mixture of water and
condensed steam into the boiler against boiler pressure.

The space at the end of the steam nozzle is maintained
at a low temperature by the feed water flowing through
it and the pressure of the steam is therefore very low at this
point, being less than atmospheric in most cases. Atmos-
pheric pressure is therefore able to force water up the suction
pipe if the " lift " is not too great, and when once started
such a device can therefore " raise " its own water as well
as delivering it against pressure.

It is interesting to note that the efficiency of this appa-
ratus is almost 100 per cent on a heat basis. All heat not
radiated from the apparatus is returned to the boiler in the
mixture of condensed steam and feed water and, as the
external surface is very small, very little heat is lost by.
radiation.

161. Separators. Two kinds of separators are used in
steam plants: (a) the oil separators already referred to for
separating oil from exhaust steam, and (6) steam separators,
which separate water from steam.

As it is impossible entirely to prevent radiation from
steam pipes, it follows that condensation will occur in any
pipe line which carries saturated steam. Water is also
formed in the cylinders of reciprocating engines not supplied
with very highly superheated steam, and much of it is
generally present in the exhaust of the high and inter-
mediate cylinders of multiple-expansion engines.

A small amount of water can be passed through the
cylinder of a reciprocating engine without mechanical
damage, but it probably causes a loss of heat by cling-
ing to the walls and assisting in the heat interchanges
which always occur. Large quantities of water are apt
to cause mechanical damage, as water is inelastic, and if

BOILER-FEED PUMPS AND OTHER AUXILIARIES 387

Sieves

/Jacket of Insulating
Material to Decreasa
Radiation Loss.

Drain-)

FIG. 231. — Steam Separator.

388 STEAM POWER

more of it is trapped in a cylinder end than can be con-
tained in the clearance, something must give way when
the piston reaches the end of its stroke.

It is customary to separate as much as possible of the
water of condensation before admitting steam to the
cylinder. The separators used are built in many different
shapes and types, but practically all depend upon two
principles. These are:

(1) Water is much more dense than steam, and if a
stream of a mixture of water and steam be made to travel
in a curve, the water will therefore collect at the outside
of the curve, and

(2) Water brought into violent contact with metallic
surfaces " wets " them and has a tendency to adhere thereto.

In steam separators the stream of mixture is therefore
made to change its direction of flow suddenly and to impinge
upon baffles in such a way that the greater part of the
liquid is caught and drained off.

One form of separator is shown in Fig. 231. The mixture
impinges on sieves in the first part of its passage through
the separator, part of the water passing through the open-
ings and draining to the reservoir at the bottom of the
device. Ridges and troughs catch all water separated and
guide it to drains leading to the reservoir, so that no water
which is once deposited is again picked up by steam.

Another form of separator is illustrated in Fig. 232.
The steam impinges upon the inverted V-shaped casting
and water caught on the projecting ridges drains toward
the sides and then downward into the receiver, while the
steam passes on as shown.

162. Steam Traps. In the separators just described,
there is a constant accumulation of water which must be
periodically drained off if the entire device is not to fill
up and become inoperative. Similarly there is a constant
accumulation of liquid in steam jackets, in receivers of
multi-expansion engines and in low points in steam lines.

BOILER-FEED PUMPS AND OTHER AUXILIARIES 389

To drain all such accumulations by hand from time to
time, as necessary, would be both time consuming and
uncertain, and devices known as traps have therefore been
developed for doing this automatically. Traps are arranged
to collect condensation until they fill to a predetermined
point. When this occurs they automatically discharge in
such a way as to prevent the escape of steam and are then
ready to receive liquid again.

Traps may be arranged to receive condensation from
high or low-pressure mains or from spaces in which a vacuum

(<*) (*)

FIG. 232. — A Steam Separator.

is maintained, and to discharge it to a hot well or similar
receiver or even into the boiler itself.

The principles upon which traps work are very nu-
merous and there are many different designs. The more
common either make use of the weight of the accumulating
water to cause the trap to discharge, or they make use of
floats resting on the condensate and opening the discharge
valve when a certain height is reached, or they depend upon
expansion and contraction of certain parts when exposed
to steam of high temperature and cooler condensate respect-
ively.

390 STEAM POWER

163. Steam Piping. There is a great deal of piping of
various kinds in all steam plants and the financial success
or failure of a plant often depends upon this apparently
insignificant item. It is beyond the limits of a book of this
scope to consider the many different forms of piping and the
many different ways in which apparatus may be connected.
This is a study in itself and one of great importance.

It should be noted, however, that all of the following
points must be kept in view when designing and installing
piping and that that installation which most nearly meets
all these requirements may be regarded as the best.

(1) The various lines should conduct the materials flow-
ing through them with the minimum loss of pressure and
with the minimum loss (or gain) of heat.

(2) The pipe lines should be so constructed as to make
failure of a dangerous sort, from expansion and contraction,
water hammer and such, most unlikely if not impossible.

(3) All connections should be so made that the careless
manipulation of valves cannot cause an accident.

(4) The number of flange and screw connections and the
number of valves and fittings should be reduced to the
minimum, as they are often sources of weakness and are
always costly.

(5) The entire layout should be so arranged that inter-
ruption of service because of pipe, or valve, failure is (as
nearly as possible) impossible.

(6) The cost of the system should be as small as it can
be made, consistent with the other requirements.

It is almost unnecessary to say that all of these desirable
ends are never attained in any plant. A compromise must
always be made in order to bring the cost within reasonable
limits, but most of the recent installations show a tendency
toward better design in this part of the plant and a con-
sideration of reliability and safety far in excess of what was
formerly customary.

TABLES

392

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SATURATED STEAM TABLE

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394

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396

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SATURATED STEAM TABLE

397

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398

STEAM POWER

PROPERTIES OF ONE POUND OF SUPERHEATED STEAM

[Condensed from Marks and Davis's STEAM TABLES AND DIAGRAMS, 1909, by
permission of the publishers, Longmans, Green & Co.]

Sp.V. = specific volume in cu.ft.; AQ = B.t.u. total heat above
32° F.; A0 = total entropy above 32° F.

Absolute
Pressure.
Lbs. Sq.in.

Degrees of Superheat.

Sat. Temp.
0 F.

100

150

200

250

300

15 |

(213) }

Sp.V.
AQ

A<£

26.27
1150.7
1 . 7549

28.40
1174.2

1.7886

30.46
1197.6
1.8199

32.50
1221.0
1.8492

34.53
1244.4

1.8768

36.56
1267.7
1.9029

38.58
1291 . 1
1.9276

60 |

(281) )

Sp.V.
AQ
A0

8.51
1173.6
1 . 6581

9.19
1198.8
1 . 6909

9.84
1223 . 4
1.7211

10.48
1247.7
1.7491

11.11
1271.8
1 . 7755

11.74
1295.8
1.8002

12.36
1319.7
1.8237

100 j

(327. 8) |

Sp.V.
AQ

4.43
1186.3
1.6020

4.79
1213.8
1.6358

5.14
1239 . 7
1.6658

5.47
1264.7
1 . 6933

5.80
1289.4

1.7188

6.12
1313.6
1 . 7428

6.44
1337.8
1 . 7656

110 f

(334. 8) 1

Sp.V.
AQ

4.05
1188.0
1 . 5942

4.38
1215.9
1 . 6282

4.70
1242.0
1.6583

5.01
1267.1
1.6857

5.31
1291.9
1.7110

5.61
1316.2
1 . 7350

5.90
1340.4
1.7576

120 f

(341. 3) 1

Sp.V.
AQ
A0

3.73
1189.6
1.5873

4.04
1217.9
1.6216

4.33
1244.1
1.6517

4.62
1269.3
1 . 6789

4.89
1294.1
1 . 7041

5.17
1318.4
1.7280

5.44
1342.7
1.7505

130 f

(347. 4) j

Sp.V.
AQ
A<£

3.45
1191.0
1.5807

3.74
1219.7
1.6153

4.02
1246.1
1 . 6453

4.28
1271 A
1.6724

4.54
1296.2
1.6976

4. 80
1320.6
1.7213

5.05
1344.9
1 . 7437

140 j
(353. 1)|

Sp.V.
AQ
Atf>

3.22
1192.2
1.5747

3.49
1221 .4
1 . 6096

3.75
1248.0
1.6395

4.00
1273.3
1.6666

4.24
1298.2
1.6916

4.48
1322.6
1.7152

4.71
1346.9
1.7376

150 f

(358. 5) |

Sp. V.
AQ

3.01
1193.4
1 . 5692

3.27
1223.0
1 . 6043

3.51
1249 . 6
1 . 6343

3.75
1275.1
1.6612

3.97
1300.0
1.6862

4.19
1324.5
1.7097

4.41
1348.8
1 . 7320

SUPERHEATED STEAM TABLE

399

PROPERTIES OF ONE POUND OF SUPERHEATED STEAM

(Continued)

Absolute
Pressure.
Lbs. Sq.in.

Degrees of Superheat.

250

300

100

150

200

Sat. Temp.
0 F.

160 {
(363.6)}

Sp. V.
AQ

A</>

2.83
1194.5
1.5693

3.07
1224.5
1.5993

3.30
1251.3
1 . 6292

3.53
1276.8
1.6561

3.74
1301.7
1 . 6810

3.95
1326.2
1.7043

4.15
1350.6
1.7266

170 f

(368.5)}

Sp. V.
AQ
A0

2.68
1195.4
1 . 5590

2.91
1225.9
1 . 5947

3.12
1252.8
1.6246

3.34
1278.4
1.6513

3.54
1303 . 3
1.6762

3.73
1327.9
1.6994

3.92
1352.3
1.7217

180 f

(373.1)}

Sp. V.
AQ
A<£

2.53
1196.4
1.5543

2.75
1227.2
1.5904

2.96
1254.3
1.6201

3.16
1279.9
1 . 6468

3.35
1304.8
1.6716

3.54
1329.5
1 . 6948

3.72
1353.9
1.7169

190 {

(377. 6) 1

Sp. V.
AQ

A<£

2.41
1197.3
1.5498

2.62
1228.6
1 . 5862

2.81
1255.7
1.6159

3.00
1281.3
1.6425

3.19
1306.3
1.6627

3.37
1330.9
1.6904

3.55
1355.5
1.7124

200 j

(381. 9) \

Sp. V.
AQ
A<£

2.29
1198.1
1.5456

2.49
1229 . 8
1.5823

2.68
1257.1
1.6120

2.86
1282.6
1 . 6385

3.04
1307.7
1 . 6632

3.21
1332.4
1 . 6862

3.38
1357.0
1 . 7082

300 f

(417.5)}

Sp. V.
AQ

A</>

1.55
1204.1
1.5129

1.69
1240.3
1.5530

1.83
1268.2
1.5824

1.96
1294.0
1.^082

2.09
1319.3
1.6323

2.21
1344.3
1.6550

2.33
1369.2
1.6765

500 f

(467.3))

Sp. V.
AQ
A<£

0.93
1210.0
1.470

1.03
1256
1.519

1.11
1285
1.548

1.22
1311
1.573

1.31
1337
1.597

1.39
1362
1.619

1.47

1388
1.640

INDEX

PAGE

Absolute pressures 41

Absolute temperature scale 12

Action of steam, in cylinder 24

on impulse blades of steam turbine 234-236

Adiabatic expansion 58

Advance angle 167

Advantages of condensing 251, 252

Advantages, relative, of contact and non-contact condensers. . 274, 275

Air, excess, combustion 286

Advantages and disadvantages of 311

Analogy, hydraulic 26

Analyses of coal (see fuels) 299-301

Purchase of coal on analysis 302

Angle of advance 167

Ash in coal 300

Atmospheric line on indicator diagram 119

Atoms 278

Avogadro's Law 281

Babcock & Wilcox superheater 367

' ' " " water-tube boiler 352-355

Balanced slide valves 184

Barometer, conversion of readings from inches mercury to pounds

per square inch 255, 256

Barometric Condenser 261-266

Baume scale to express gravity 303

Bearings Ill, 112

Bilgram diagram 168-182

Angularity of connecting rod 179

Diagram for both cylinder ends 177

Exhaust and compression 175-177

Indicator diagram from 180-183

Piston positions 177-182

Blades, impulse, action of steam on, in impulse turbine 234-236

401

402 INDEX

PAGE

Boiler-feed pumps and other auxiliaries 382-390

Boiler, generation of steam in 38, 39

Boilers, steam 305-374

Circulation in 341, 342

Classification according to —

(1) form; (2) location of furnace; (3) use; (4) direction
of principal axis; (5) relative positions of water and hot

gases 305, 306

Draft apparatus 370-373

Chimneys or stacks 370-372

Mechanical draft 372, 373

Effects of soot and scale 364, 365

Efficiencies 363, 364

Functions of parts 306-308

Furnaces and combustion 308-311

Hand firing 311-315

Mechanical grates 315, 316

Mechanical stokers 317-335

Rate of combustion 335-337

Rating 358-362

Boiler horse-power 359

Equivalent evaporation 361

Scale 365, 366

Prevention of 366

Smoke and its prevention . . . 316, 317

Strength and safety 337-341

Superheaters —

Built in 366, 367

Separately fired 366, 367

Babcock & Wilcox 367

Foster 369, 370

Heine 367, 368

Types of boilers 342-359

Babcock & Wilcox, water tube ...;.. ... 352-355

Continental 346-348

Externally fired, return tubular 349-352

Heine water-tube 355, 356

Internally fired, tubular 342-345

Locomotive • 346

Scotch marine 348, 349

Sterling water-tube 356-358

Vertical fire tube 343-345

Wickes vertical water-tube 358, 359

INDEX 403

PAGE

British Thermal unit 3, 13

Buildings, heating of, by exhaust steam 376, 377

Built-in superheaters . 366, 367

Calorific value of coals —

Dulong's formula 301, 302

Fuel Calorimeter 302

Calorific value of, petroleum oils 303, 304

Calorimeter, fuel 302

Carbon, combustion of 279

CO, combustion to 279-282

CO. combustion to 282, 283

CO and CO., conditions determining formation of. . . 284-286

CO to CO2, combustion of 283, 284

complete combustion of 308-310

flue gases from combustion of 286, 287

Card factors and conventional diagram 125-128

Centigrade scale 10, 11, 12

Chain grate stokers . . 318-322

Chart —

Mollier, for steam 230

temperature-entropy, for steam 62-65

Chimneys or stacks 370-372

Circulation in boilers 341-342

Classification of boilers, according to —

(1) form; (2) location of furnace; (3) use; (4) direction of prin-
cipal axis; (5) relative postion of water and hot gases 305-374

Classification of steam engines 92, 93

Clearance — steam engine —

mechanical and volumetric 84, 85

Clearance volume determined from diagram 131, 132

Closed and open feed-water heaters 377-380

Coal-fuels 297-299

Analyses of — proximate and ultimate 299-301

Purchase of, on analysis 302

Coefficient, excess in combustion 286

Combined indicator diagrams 155-158

Combined type turbine 247

Combustion and furnaces; steam boilers 308-311

Combustion 277-295

Definitions — Compounds, elements, heat or calorific value,
atoms, molecules, etc . 277-279

404 INDEX

PAGE

Combustion —
Combustion of —

Carbon 279

Hydrocarbons 289, 290

Hydrogen 287-289

Calorific value of 290

Mixtures 290, 291

Sulphur 290

Combustion to —

CO 279-282

CO, 282, 283

CO to CO2 283, 284

Conditions determining formation of CO and CO3 284-286

Excess Air and excess coefficient 286

Flue gases from combustion of carbon 286, 287

Rate of, in boiler furnaces 335-337

Temperature of combustion 291-294

Theoretical temperature 292

Commercial fuels — solid, liquid and gaseous 296, 297

Complete expansion cycle 55-58, 72

Complete 7Vchart for steam 68-70

Compound engine 149-151

Compounding 141-158

Combined indicator diagram 155-158

Compounding 144-149

Cylinder ratios 151-153

Gain by expansion 141-144

Indicator diagrams and mean pressures 153-155

The compound engine 149-151

Compounds — combustion 277-279

Compression and exhaust — Bilgram diagram 175-177

Condensation, cylinder, methods of decreasing 89-92

Condensation, initial 81, 82

determination of 86-89

Condensers and related apparatus 251-276

Advantages of condensing 251, 252

Conversion of readings from inches of mercury to

Ibs. per square inch 255, 256

Cooling towers 275, 276

Measurement of vacuum 252-255

Principle of 256-258

Types of —

Contact.. . 258-268

INDEX 405

PAGE

Condensers (continued) —
Contact-
Barometric 261-266

Jet, Parallel flow 259, 260

Siphon 266

Westinghouse — Leblanc 267, 268

Non-contact 268-271

Surface 268-270

Two-pass or double flow 270-271

Relative advantages 274, 275

Water required by contact condensers 271-273

Water required by non-contact condensers 273, 274

Condensing, advantages of 251, 252

Condensing plants 23

Conditions determining formation of CO and CO2 284-286

Connecting rod 109, 110

Angularity of 179

Conservation of Energy, law of 2

Conservation of Matter, law of 1

Constant-quality lines on T</>-chart 66, 67

Constant volume lines, on TV-chart 68

Constant speed governing 215, 216

Contact condensers 258-268

Continental type boiler 346-348

Conventional diagram and card factors 125-128

Conversion of barometric readings, from inchs mercury to pounds

per square inch 255, 256

Cooling towers 275, 276

Corliss and other high-efficiency engines 196-212

Locomobile type 210-212

Non-detaching Corliss gears 201-205

Poppet valves 205-208

Trip-cut-off Corliss 196-201

Unaflow engine 208-210

Corliss engine, trip-cut-off 196-201

Corliss gears, non-detaching 201-205

Crank end of engines 98

Cross-head and guides 107, 108

Cushion steam and cylinder feed 85, 86

Cut-off governing 215

Cut-off ratio 128

Cycle, area on 7>-chart representative of work 73

Complete expansion 55-58, 72

406 . INDEX

PAGE

Cycle, incomplete expansion 58-60, 74, 75

Modifications for wet and superheated steam 73, 74

Of events in simple steam power plant 22

Theoretical, of steam turbine 225-228

Cycles, desirability of various, in engines 55

Cylinder, action of steam in 24

Condensation, methods of decreasing 89-92

Efficiency 139

Feed and cushion steam 85, 86

Ratios 151-153

Cylinder and steam chest 101, 102

Decreasing cylinder condensation 89-92

De Laval impulse turbine 236-238

Density, specific, of dry saturated steam 38

Description and method of operation of D-slide valve 159-165

Design of nozzle, steam turbine 228-234

Determination of clearance volume from diagram 131, 132

Determination of I.h.p 120-124

Developed horse-power 137

Developed thermal efficiency 138

D-slide valve 159-195

Angle of advance 167

Angularity of connecting rod 179

Bilgram diagram • 169

Description and method of operation 159-165

Diagram for both cylinder ends 177

Exhaust and compression 175

Exhaust lap 168

Indicator diagram from Bilgram diagram 180

Lead 166, 167

Limitations of D-slide valve 183-185

Piston positions 177

Reversing engines 185-187

Steam lap— outside lap 165, 166

Valve setting 187-195

D-slide valve engine, simple 96-98

Diagram, Bilgram, for both cylinder ends 177

Bilgram, indicator diagram from 180-183

Indicator 24

Indicator and mean pressures for compound engines . 153-155

combined 155-158

Indicator, conventional and card factors 125-128

INDEX 407

PAGE

Diagram, water rate 86, 132-136

Diagrams from real engine ; . . . 192, 193

Double acting engines 55

Double-flow condenser 270

Downdraft furnace 315

Draft apparatus 370-373

Chimneys or stacks 370-372

Mechanical draft 372, 373

Dry-air pump 264

Dry-saturated steam, total heat of 33

Specific density of 38

Specific volume of 36-38

Dry-vacuum pump 264

Dulong's formula — combustion 301, 302

Duplex steam pump 383, 384

Eccentric 160-165

Economy of turbines 247-249

Effective pressure, mean, methods of varying 215

Efficiency 52, 53

Cylinder ,. ./ . . .' 139

Developed thermal 138

Effect of temperature range on 75

Indicated thermal 138

Mechanical and thermal 137-140

Of boilers 363, 364

Relative ' 139

Elements — combustion 277

Energy —

Conservation of energy, law of 2

Heat 2

Mechanical 2

Units of 3

Engine —

Application of theory for an ideal to a real 54, 55

Compound, triple, quadruple, quintuple 148

Receiver type 149

Tandem and cross-compound 151

Woolf type 149

Desirability of various cycles 55

Double acting 55

Efficiency 52, 53

Heat quantities involved 50-52

408 INDEX

PAGE

Engine (continued) —

Ideal steam 43-60

Operation of 45, 42

Operation of the real steam 77-80

Reversing 185-187

Steam —

Classification —

1) On basis of rotative speed; (2) Ratio of
stroke to diameter; (3) Valve gear; (4)
Position of longitudinal axis; (5) Num-
ber of cylinders; (6) Cylinder arrange-
ment; (7) Use 92, 93

Clearance, volumetric and mechanical 84, 85

Crosshead and guides 107, 108

Cushion steam and cylinder feed 85, 86

Diagram water rate 86

Cylinder and steam chest 101, 102

Determination of initial condensation 86-89

Initial condensation 81

Losses, in real installations 80-84

Methods of decreasing cylinder condensation. .. 89-92

Nomenclature 98

Principal parts .......... 98-114

Bearings Ill, 112

Connecting rod 109, 110

Crosshead and guides 107, 108

Cylinder and steam chest 101, 102

Flywheels 112, 113

Frame 99, 100

Piston 102-106

Piston rod and tail rod 106, 107

Shaft .. 110, 111

Re-evaporation in 83, 84

Rotative and piston speed 93-96

Simple D-slide valve 96-98

Throttling or wire-drawing 82, 83

Work done by 46-60

Engines, Corliss and other high efficiency 185-187

Locomobile type 210-212

Non-detaching Corliss gears 201-205

Poppet valves 205-208

Trip-cut-off Corliss 196-201

Unaflow. . , 208-210

INDEX 409

PAGE

Entropy diagram 61-71

of liquid, vaporization, and dry saturated steam 61-63

TV-chart for steam 62-65

Complete 68-70

Constant quality lines 66, 67

Diagram for a real engine 136

Heat from 68

Quality from 65-68

Saturation curve 63

Superheating lines 63, 64

Volume from 68

Water line 63

Entropy, diagrams of steam cycles 72-76

Equivalent evaporation, boilers 361

Excess air — combustion 286

Advantages and disadvantages 311

Excess coefficient 286

Exhaust and compression — Bilgram diagram 175-177

Exhaust lap 166, 168

Exhaust steam, utilization of, for heating buildings 376, 377

Expansion, adiabatic 58

Cycle, the complete 55-58, 72

the incomplete 58-60

Gain by, in compounding 151-144

Ratio of, apparent and real 128-130

External latent heat of vaporization 31

Externally fired, return tubular boiler 349-352

Fahrenheit scale 11, 12

Feed-water heating 377

Open and closed heaters 377-380

Firing boilers by hand 311-315

Fixed carbon in coal 300

Flywheel 97, 98, 112, 113

Regulation 213, 214

Flue gases from combustion of carbon 286, 287

Foot-pound, definition 3

Forward stroke of engines 98

Foster superheater . . 369, 370

Frames of engines 99, 100

Front end of engines 98

Fuel calorimeter 302

Fuels... . 296-304

410 INDEX

PAGE

Fuels (continued) —
Commercial —

Solid, liquid, gaseous 296, 297

Coal 297-299

Analyses 299-301

Calorific value of —

Dulong's formula 301, 302

Fuel Calorimeter ' 302

Petroleum 302

Baume scale to express gravity 303

Calorific values 303, 304

Purchase of coal on analysis 302

Functions of boiler parts 306-308

Furnaces — and combustion 308-311

— Updraft and downdraft 315

Gases, and vapors, steam 27

— Flue, from combustion of carbon 286, 287

Gaseous fuels 296, 297

Gauge pressure 39-41

Gearing and staging — turbines 238-243

Gears, Corliss, non-detaching. 201-205

Generation of steam in real steam boiler 38, 39

Generation of steam or water vapor 28

Governing — throttle and cut-off 215

Coefficient of regulation 216

Constant speed 215, 216

Governor 97, 98

Regulation 214, 215

Governors —

Pendulum 217

Rites inertia 218-220

Shaft 217,218

Grates, mechanical 315, 316

Gridiron valve 185

Guides and crosshead 107, 108

Hand firing — steam boilers 311-315

Heat 9

Absorption, reversal of process 38

Energy 2

Unit of 13

From 7>-ehart 68

tNDEX m 411

PAGE

Heat (continued) —

Latent, of vaporization 30, 32

Internal and external 30, 31

Of liquid, q or h 31, 32

Of superheat 34, 35

Quantities in rectangular cycle 50-52

Quantity of 16

Specific 14

Total, of dry saturated steam 33

Of superheated steam 36

Of wet steam 33, 34

Value of elements and compounds 277-279

Heat, waste — in steam plant 375-381

Feed-water heating 377

Open and closed heaters 377-380

Utilization of exhaust for heating buildings 376, 377

Heaters, feed-water, open and closed 377-380

Heine superheater 367, 368

Heine water-tube boiler 355, 356

Horizontal, return, tubular boiler 306-308

Horse-power 17

Developed 137

Hour, definition 18

Of steam boilers 359

Hydraulic analogy ; 26

Hydrocarbons, combustion of 289, 290, 310

Calorific value of . . 290

Hydrogen, combustion of 287-289

I.h.p. — determination of 120-124

Impulse steam turbine . 221-225

De Laval type 236-238

Inclined stokers 322-325

Incomplete expansion cycle 58-60, 74, 75

Indicated thermal efficiency 138

Indicator 115

Indicator diagram 24, 115-140

Atmospheric line 119

Conventional and card factors 125-129

Cut-off ratio 128

Determination of clearance volume from diagram. 131, 132

Determination of I.h.p 120-124

Diagram factor or card factor 126-129

412 INDEX

PAGE

Indicator diagram (continued) —

From Bilgram diagram 180-183

Mean effective pressure 122

Ratio of expansion 128-131

Reducing mechanism 118

Scale of spring 118

The planimeter 123

Indicator diagrams and mean pressures for compound engines 153-155

Combined 155-158

Indicator diagrams from real engine 192, 193

Injector, steam 384-386

Inertia governor, Rites 218-220

Initial condensation 81, 82

Determination of 86-89

Inside lap, negative 168

Internal latent heat of vaporization 30

Internally fired, tubular boilers 342-345

Jet condensers 259, 260

Joule's equivalent 14

Joule, the 3

Kinetic mechanical energy 8

Lap angle 166

Lap, steam 165, 166

Negative inside .- 168

Outside and exhaust 166, 168

Latent heat of vaporization 30, 32

Internal and external 30, 31

Lead 166, 167

Leblanc — Westinghouse condenser 267, 268

Liquid fuels 296, 297

Liquid, heat of, q or h 31, 32

Entropy of 61

Limitations of D-slide valve 183-185

Balanced slide valves 184

Gridiron valve 185

Piston valve 184

Riding cut-off valves 185

Locomobile type of high efficiency engines 210-212

Locomotive type boiler 346

Low-pressure or exhaust steam turbines 249

INDEX 413

PAGE

Matter 1

Law of conservation of matter 1

Units of matter 3

Mean effective pressure 122

Methods of varying 215

Mean pressures and indicator diagrams for compound engines . 153-155

Measurement of temperature 10

Measurement of vacuums 252-255

Mechanical and thermal efficiencies 137-140

Mechanical clearance, steam engine 84, 85

Mechanical draft 372, 373

Mechanical energy : 2, 3, 7

Potential and kinetic 7, 8

Mechanical grates 315, 316

Mechanical stokers 317-335

Mercury readings, conversion to pounds per square inch 255, 256

Mercury thermometers 10-12

Method of operation and description of D-slide valve 159-165

Mixtures, combustion of 290, 291

Moisture in coal 299

Molecular activity 9

Molecules * 278

Natural draft, chimneys 370

Negative inside lap 168

Non-condensing plants 23

Non-contact condensers 21 3-271

Surface (Wheeler) 268-270

Non-detaching Corliss gears 201-205

Nozzle design, steam turbine 228-234

Oil firing 333-335

Open and closed feed-water heaters . 377-380

Operation of simplified steam engine 45, 46

Operation of real steam engine 77-80

Outside steam lap 166

Outstroke of engine 98

Parallel-flow condenser 259-261

Parson's type turbine .* 246

Pendulum governors 217

Petroleum 302

Baume* scale to express gravity of 303

Calorific values. , 303, 304

414 INDEX

PAGE

Piping, steam 390

Piston, engine . . : 102-106

Piston positions for Bilgram diagram 177-182

Piston rod and tail rod 106, 107

Piston speeds of steam engines 93-96

Piston valve 184

Planimeter 123

Plant, steam power 20

Plants, condensing, non-condensing 23

Poppet valves 205-208

Positions of piston for Bilgram diagram 177-182

Potential mechanical energy 7

Powdered coal stokers 333

Power and work 17

Power, unit of, horse power 17

Pressure, absolute 41

Gauge 39-41

Mean effective 122

Methods of varying 215

Pressures, mean, and indicator diagrams for compound engines. 153-1 55

Prevention of smoke 316, 317

Prime-mover 20

Principal parts of engines 98-114

Principle of condenser 256-258

Properties of steam 27

Proximate analysis of coal 299

Pump, dry air or dry vacuum 264

Vacuum 259

Pumps, boiler feed 382-384

Purchase of coal on analysis 302

Quality from 7>-chart 65-68

Constant, lines 66, 67

Quantity of heat 16

Rate, diagram water 86, 132-136

Rate of combustion in boiler furnaces 335-337

Rating of steam boiler 358-362

Ratio, cut-off 128

Ratio of expansion — apparent and real 128-130

Ratios, cylinder 151-153

Reaction type turbine 243-247

Receiver engine 149

INDEX 415

PAGE

Recovery of waste heat 375-381

Reducing mechanism 118

Re-evaporation 83

Regulation 213-220

Coefficient of governor 216

Constant speed governing 215, 218

Governors —

Pendulum 217

Rites inertia. . 218-220

Shaft 217, 218

Kinds — flywheel and governor 213-215

Methods of varying mean effective pressure — Throt-
tling and cut-off 215

Relative advantages of contact and noo-contact condensers. . 274, 275

Relative efficiency 139

Return tubular boilers, horizontal 306-308, 349-352

Reversal of process of heat absorption 38

Reversing engines 185-187

Riding cut-off valve 185

Rites inertia governor 218-220

Rotative speeds of steam engines 93-96

Safety and strength of boilers 337-341

Saturated steam, dry, specific volume of 36-38

Saturated vapor 31

Saturation curve, temperature entropy chart for steam 63

for compound engine cards 156

Scale 365, 366

Prevention of 366

Scale of spring, indicator 118

Scotch marine type boiler 348, 349

Separately fired superheaters 366, 367

Separators 386-388

Setting, valve ...... , . . . 7,. . . , , , , , , , 187-195

Shaft governors 217, 218

Shaft of engine. '. 110, 111

Simple D-slide valve engine 96-98

Siphon condensers 266

Slide valves 184

Balanced 184

Gridiron valve 185

Piston valve 184

Riding cut-off valve 185

416 INDEX

PAGE

Smoke and its prevention 316, 317

Solid fuels 296, 297

Soot and scale, effects of, in boilers 364, 365

Specific density of dry, saturated steam 38

Specific heat 14

Specific volume of dry saturated steam 36-38

Speeds, rotative and piston, of steam engines 93-96

Spring, scale of, indicator 118

Stacks or chimneys 370-372

Staging and gearing, steam turbines 238-243

Steam, action in cylinder : 24

Action of, on impulse blades of turbine 234-236

Boiler, generation of steam in 38, 39

Cushion, and cylinder feed 85, 86

Diagram water rate 86

Cycles, T$-diagrams of, 72-76

Steam engine the ideal 43-60

Bearings Ill, 112

Classification 92, 93

Connecting rod 109, 110

Crosshead and guides 107, 108

Cylinder and steam chest 101, 102

Determination of initial condensation 86-89

Flywheel and governor 97, 98

Flywheels 112, 113

Frame 99, 100

Losses in real installations 80-84

The real 77-114

Initial condensation 81

Re-evaporation 83, 84

Throttling 82

Wire-drawing 82, 83

Methods of decreasing cylinder condensation .... 89-92

Nomenclature of 98

Operation of, 77-80

Piston 102-106

Piston rod and tail rod 106, 107

Principal parts. 98-114

Rotative and piston speeds 93-96

Simple D-slide valve 96-98

Steam, entropy of dry saturated 61, 62

Generation of 28-39

Heat of superheat 34-35

INDEX 417

PAGE

Steam, lap, D-slide valve 165, 166

Modification of T$-chart for wet and superheated 73, 74

Properties of 27

Specific density of dry saturated 38

Specific volume of dry saturated 36-38

Temperature-entropy chart for 62-71

7>-chart complete 68-70

Total heat of dry saturated 33

Total heat of wet 33, 34

Vapors and gases 27

Wet, effect of 53

Steam injector i 384

Steam piping 390

Steam power plant 20-22

Steam trap 388

Steam turbine (see Turbine)

Stephenson link gear 186

Sterling water-tube boiler 356-358

Stokers, mechanical 317-313

Chain grate 318

Inclined, overfeed 323-327

Powdered coal 333

Sprinkler 318

Underfeed 327-333

Strength and safety of boilers 337-341

Sulphur, combustion of 290

Sulphur in coal 301

Superheat, heat of 34, 35

total heat of 36

Superheaters —

Built in 366, 367

Separately fired 366, 367

Babcock and Wilcox 367

Foster. .' 369,370

Heine 367, 368

Superheating 31

Lines, on temperature-entropy chart for steam . . 63, 64
Surface condensers 268, 269

Tail rod and piston rod of engine 106, 107

Temperature 9

Measurement of 10

Pressure relations 29

418 INDEX

PAGE

Temperature of combustion 291-294

Temperature rise 293

Theoretical 292

Temperature-entropy chart for steam 62-71

Complete chart 68-70

Heat from 68

Quality from 65-68

Volume from 68

T (^-diagram for a real engine 136

T(£-diagrams of steam cycles 72-76

Complete expansion cycle 72

Area of cycle representative of work 73

Modifications for wet and superheated steam 73

Temperature range, effect on efficiency 75

Temperatures of vaporization 29

Theoretical cycle of steam turbine 225-228

Thermal and mechanical efficiency 137-140

Developed thermal efficiency 138

Indicated thermal efficiency 138

Thermometers, mercury 10-12

Throttle governing 215

Throttling or wire-drawing 82

Towers, cooling 275, 276

Traps, steam 388, 389

Trip-cut-off Corliss engine 196-201

Triple expansion 148

Tubular boiler, horizontal return 306-308

Turbine, steam 221-250

Action of steam on impulse blades 234-236

Combined type 247

De Laval impulse type 236-238

Economy of 247-249

Gearing and staging 238-243

Impulse .' 221-225

Nozzle design 228-234

Reaction type 243-247

Theoretical cycle 225-228

Types of boilers 342-359

Babcock & Wilcox, water-tube 352-355

Continental 346-348

Externally fired, return tubular 349-352

Heine water-tube 355, 356

Internally fired, tubular 342-345

INDEX 419

PAGE

Types of boilers (continued) —

Locomotive 346

Scotch marine 348, 349

Sterling water-tube . . . : 356-358

Wickes vertical water-tube 358, 359

Types of condensers —

Contact 258-268

Barometric 261-266

Jet, parallel flow type 259, 260

Siphon ' 266

Westinghouse-Leblanc 267, 268

Non-contact —

Surface 268-270

Two-pass or double flow 270, 271

Ultimate analysis of coal 299-301

Una-flow engine 208-210

Underfeed stokers 325-333

Unit of heat energy 13

Units of matter, energy and work 3

Updraft furnace 315

Utilization of exhaust steam for heating buildings 376, 377

Vacuum 252-255

Measurement of 253, 254

Pump 259

Valve, D-slide — (see D-slide valve) 159-195

Setting 187-195

Valves. . .' 159-195, 205-208

Balanced slide 184

Gridiron 185

Piston -v 184

Poppet k 205-208

Riding cut-off 185

Vapor, saturated 31

Pressure-temperature relations ; saturated water vapor .... 29

Water or steam, generation of 28

Vaporization, entropy of 61

Latent heat of 30, 32

Internal and external 30, 31

Temperatures of 29

Vapors and gases 27

Volatile matter in coal . . . 299

420 INDEX

PAGE

Volume, clearance, determined from diagram 131, 132

Constant, line 68

From T0-chart 68

Volumetric clearance, steam engine 84, 85

Waste heat, in steam plant 375, 376

Recovery of 375-381

Water line, temperature entropy chart for steam 63

Water rate, diagram, steam engine 86, 132-136

Water required by contact condensers 271-273

Water-tube boilers 352-359

Water vapor or steam, generation of 28

Water vapor, saturated, pressure-temperature relations 29

Weight of water required by non-contact condensers 273, 274

Westinghouse-Leblanc condenser 267, 268

Westinghouse-Parsons turbine 245

Wet and superheated steam; modifications of 1 (£-chart for. ... 73, 74

Wet steam, total heat of 33, 34

Wickes vertical water-tube boiler 358, 359

Wire-drawing or throttling 82

Woolf type engine 149

Work 3, 17

Area of cycle on T$-diagram representative of work 73

Done by the engine 46-50

Unit of.. 3

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