Engineering Science Series

ENGINES AND BOILERS

ENGINEERING SCIENCE SERIES

EDITED BY
DUGALD C. JACKSON, C.E.

PROFESSOR OF ELECTRICAL ENGINEERING

MASSACHUSETTS INSTITUTE OF TECHNOLOGY

FELLOW AND PAST PRESIDENT A.I.E.E.

EARLE R. HEDRICK, Ph.D.

PROFESSOR OF MATHEMATICS, UNIVERSITY OF MISSOURI
MEMBER A.S.M.E.

ENGINES AND BOILERS

THOMAS T. EYRE

DEAN, COLLEGE OF ENGINEERING

STATE UNIVERSITY OF NEW MEXICO,

FORMERLY ASSISTANT PROFESSOR OF

MECHANICAL ENGINEERING, PURDUE UNIVERSITY

J?eto gorfe

THE MACMILLAN COMPANY
1922

PRINTED IN THE UNITED STATES OF AMERICA

Set up and electrotyped. Published August, 1922.

. - FORESTRY

Press of J. J. Little & Ives Co.
New York

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, 1

Yo ^es^vv

PREFACE

This text book on Engines and Boilers is intended for use in
engineering schools which offer an elementary course in Heat
Engines. No attempt has been made to cover the more advanced
work in Thermodynamics, or to give an exhaustive treatment of
the subject of Heat Power.

This work is the result of the author's experience during the
several years that he taught classes in Engines and Boilers and in
allied subjects at Purdue University. Much of the material was
given to the students first in lectures, and later in the form of
mimeographed notes. It is now presented in book form with the
hope that it may be of value in other engineering schools.

At the end of the book a list of representative problems is given.
It has been the author's experience that the student obtains a
better understanding of the subject if he is required to work
problems related to the matter in the text.

The author wishes to thank Professor C. H. Lawrence for valu-
able suggestions made in regard to the form of presentation of some
of the work.

THOMAS T. EYRE.

University of New Mexico.

M512399

CONTENTS

CHAPTER I

PAGE

PRESSURE, TEMPERATURE AND HEAT UNITS ... 1

CHAPTER II

FUEL . ... 4

Anthracite Coal Bituminous Coal Lignite and Peat
Natural Gas Oil Coal Fields of the U. S. Coal Storage -
Determination of the Heating Value of Fuel Combustion
Composition of Flue Gas Flue Gas Analysis Heat Lost in Flue
Gas Value of CO 2 for Best Efficiency CO 2 Recorders.

CHAPTER III

STEAM ..... .... 16

Use of Steam Tables Throttling Calorimeter.

CHAPTER IV

BOILERS ........... 23

Requirements Rated Horsepower Heating Surface Rules
for Finding the Heating Surface Superheating Surface Size
of Boiler Tubes The B. & W. Boiler The Sterling Boiler The
Wickes Boiler The Return Tubular Boiler The Internally-
fired Return Tubular Boiler The Scotch Marine Boiler The
Vertical Fire Tube Boiler The Locomotive Boiler Superheaters

Horsepower of boilers Factor of Evaporation Efficiency of
Boilers A. S. M. E. Boiler Test Code.

CHAPTER V

BOILER ACCESSORIES AND AUXILIARIES ...... 45

Grates The Plain Grate The Rocking Grate Mechanical
Stokers The Chain Grate The Roney Stoker The Under-
feed Furnace Smoke Prevention Settings Draft Dam-
pers Safety Devices The Pressure Gage The Safety-valve

Safety-valve Capacity Napier's Formula Safety-valve For-
mula The Water Glass or Gage Glass High-water and Low-
water Alarm The Fusible Plug Boiler Feedwater Treatment

Scale Prevention and Removal Oil Separators Boiler Feed
Pumps The Injector Boiler Feed by Returning Trap The
Steam Line The Steam Trap Expansion Joints Steam

vii

Vlll CONTENTS

PAGE

Separators Steam-pipe Covering Feedwater Heaters Econ-
omizers Condensers The Surface Condenser The Jet Con-
denser Cooling of Circulating Water.

CHAPTER VI

THE STEAM ENGINE 76

History The Plain Slide-valve Engine Parts of the Steam
Engine Piston Displacement Clearance Steam Back of
Piston During Stroke The Indicator and its Purposes Events
of Stroke Location of Events on Diagram Equation of Expan-
sion and Compression Curves Hypothetical Indicator Diagram
Determination of Clearance from Card Determination of the
Mean Effective Pressure Indicated Horsepower Brake Horse-
power Mechanical Efficiency Thermal Efficiency Cylinder
Condensation Steam accounted for by the Indicator Diagram
Valve Setting from the Indicator Diagram.

CHAPTER VII

COMMON TYPES OF STEAM ENGINES 100

Slide-valve Engine The Corliss Engine The Four-valve
Engine The Compound Engine The Tandem-compound
The Cross-compound Cylinder Ratio The Combined Indi-
cator Diagram The Diagram Factor Ratio of Expansion
The Unaflow Engine.

CHAPTER VIII

VALVES 112

The D Slide-valve Relative Motion of Crank and Piston
Valve Diagrams The Valve Ellipse The Bilgram Diagram
The Zeuner Diagram Types of Slide-valves Valve with Pres-
sure Plate The Piston Valve Double-ported Valves The '
Gridiron Valve The Riding Cut-off Valve Effect of Rocker Arm
on the Location of Eccentric Oscillating Valves Poppet Valves
Reversing The Stephenson Link The Walschaert Valve
Gear The Joy Valve Gear Setting the Slide-valve.

CHAPTER IX

GOVERNORS ........... 137

General Classification of Governors The Gravity-balanced
Spindle Governor The Spring-balanced Governor Governing
by Changing Position of Eccentric Governing by Changing Angle
of Advance Governing by Changing Both Angle of Advance and
Valve Travel Centrifugal and Inertia Governors.

CHAPTER X

STEAM TURBINES 150

General History Fundamental Principles Available En-
ergy in Steam Velocity Due to Expansion Impulse and Reac-

CONTENTS lx

PAGE

tion Bucket Shapes Single-stage Turbine Staging Multi-
stage Impulse, Velocity-staging Multi-stage Impulse, Pressure-
staging Multi-stage Impulse, Combination Pressure-staging and
Velocity-staging Multi-stage Reaction Change of Area of
Steam Passage Leakage Loss due to Running Partial Capacity
Summary of Losses in the Steam Turbine Common Commer-
cial Types DeLaval Single-stage Multi-pressure-stage Im-
pulse Turbine Curtis Turbine Parsons Turbine Westing-
house Turbine Other Types Low-pressure Turbine Mixed-
pressure Turbine Use of Superheated Steam The Marine
Turbine.

CHAPTER XI

GAS ENGINES .......... 193

General Historical Cycles of Operation The Four-stroke
Cycle The Two-stroke Cycle Classification from Fuel Used
Efficiency Fuels The Gasoline Carburetor The Gas Pro-
ducer Cooling of Cylinders Ignition Valves Governing
Determination of Horsepower Multi-cylinder Engines.

PROBLEMS 213

ENGINES AND BOILERS

CHAPTER I
PRESSURE, TEMPERATURE AND HEAT UNITS

1. Pressure Units. In steam engineering, pressure is meas-
ured in the following units:

(1) In pounds per square inch.

(2) In inches of mercury.

(3) In inches of water.

In this country boiler pressures are ordinarily measured in
pounds per square inch above atmospheric pressure. Condenser
pressures are commonly measured from atmospheric pressure in
inches of mercury, i.e. the difference between the pressure in
the condenser and atmospheric pressure is read on a mercury
column. Draft pressures are usually measured in inches of water.
Pressure gages and vacuum gages are so constructed that they
read zero at atmospheric pressure. The atmosphere exerts a vari-
able pressure, which is about 14.5 pounds per square inch at ordi-
nary altitudes and under ordinary conditions. At sea-level, the
standard is taken as 14.7 pounds per square inch, which is equiv-
alent to 29.92 inches of mercury.

As the atmospheric pressure is slightly variable, it is necessary
in accurate work that pressures be reduced to an absolute basis.
Since the zero reading of a boiler pressure gage means atmospheric
pressure, the absolute pressure will be the sum of the gage pressure
and atmospheric pressure.

A partial vacuum usually exists in a condenser, since the abso-
lute condenser pressure is usually less than atmospheric pressure.
Since the vacuum gage as well as the boiler pressure gage reads
zero at atmospheric pressure, the absolute pressure in the con-
denser is the difference between the atmospheric pressure and the
vacuum-gage pressure.

Figure 1 shows diagrammatically the relation between these
pressures. In this figure, a is the boiler-gage pressure, b the

2 ENGINES AND BOILERS

atmospheric pressure, and c the absolute boiler pressure. Like-
wise, d represents the vacuum-gage pressure, which is measured
downward from the atmospheric pressure; and e, the difference
between 6 and d, is the absolute condenser pressure.

a 1

: r

t i

l_L_ Condenser pressure

V f

Z^/ pressure

FIG. 1

Barometers used in engineering work in this country are usu-
ally graduated in inches. The mercury barometer is used be-
cause mercury is the most convenient fluid for this use. A cubic
inch of mercury weighs very nearly 0.49 of a pound. Hence the
pressure in pounds per square inch is the barometric reading in
inches multiplied by 0.49 Vacuum gages also are commonly
graduated to read in inches of mercury. Therefore the absolute
condenser pressure in pounds per square inch is 0.49 times the
difference between the barometer and the vacuum-gage readings.

EXAMPLE 1. Find the absolute boiler pressure when the pressure gage
reads 110 pounds and the barometer reads 29.4 inches.

SOLUTION. When the barometer reads 29.4 inches, the atmospheric pres-
sure is 29.4 X0.49 = 14.4 pounds per square inch. The absolute boiler pressure
is then 14.4 + 110 = 124.4 pounds per square inch.

EXAMPLE 2. Find the absolute pressure in a condenser when the barometer
reads 29.8 inches and the vacuum gage reads 27.3 inches.

SOLUTION. The absolute condenser pressure is the difference between the
barometric and the vacuum-gage pressures, which in inches of mercury is
29.8-27.3 = 2.5. This reduced to pounds per square inch is 2.5. X49 = 1.22.

2. Temperature Units. Ordinary temperatures are measured
by means of the mercury thermometer. For higher tempera-
tures, such as those that occur in furnaces, special thermometric
devices called pyrometers are used.

Three thermometer scales are in use, the Fahrenheit, the Cen-
tigrade, and the Reaumur. In the Fahrenheit scale, the differ-
ence between the temperatures of melting ice and boiling water
at sea level is divided into 180 divisions or degrees; the freezing
point is 32 and the boiling point 212. This makes the zero

PRESSURE, TEMPERATURE, AND HEAT UNITS

Fahrenheit Centigrade Reaumur

point come 32 below the freezing point of water. In the Centi-
grade scale, the freezing point is and the boiling point is 100.
In the Reaumur scale, the freezing point is and the boiling
point is 80. Figure 2 shows graphically the relation between these
scales. It is readily seen
how the reading on one
scale may be reduced to
that of either of the
others. It is serviceable

--!. -

| 67/.S-

80' i Boili/tf temperature

femperofare

\\ ''

-J \ A Absolute zero

FIG. 2

to remember that a
difference of temperature
equal to 5 on the Cen-
tigrade scale is equal to
9 on the Fahrenheit
scale.

Experiment shows
that a perfect gas under
constant pressure at 32 Fahrenheit expands 1/491.5 part of its
own volume for each degree (F.) that its temperature is in-
creased. Because of this we call the point 491. 5 below the freez-
ing point, or 459.5 on the Fahrenheit scale, the absolute zero.
This corresponds to a Centigrade temperature of 273.

3. Heat Units. In engineering work in this country, it is
customary to use English units. Our common heat unit is the
British thermal unit (B.t.u.), which is practically defined as the
amount of heat necessary to raise the temperature of one
pound of pure water from 62 to 63 Fahrenheit. In the metric
system the engineers' heat unit is the large calorie, which is the
amount of heat necessary to raise the temperature of a kilogram
of water from 15 to 16 Centigrade. As one kilogram =2.2046
pounds, and as 1 Centigrade =1.8 degrees Fahrenheit,

one calorie =3. 968 B.t.u., or one B.t.u. =0.252 calorie.

4. Mechanical Equivalent of Heat. Heat experiments have
been made to determine the relation between heat-energy and
mechanical work. The latest and most refined experiments show
that one B.t.u. as above defined is substantially equivalent to 778
foot-pounds of work. This relation is called the mechanical equiva-
lent of heat. 1

*For definition of the "mean B. t. u." and corresponding Mechanical Equivalent of Heat
see A. S. M. E. POWER TEST CODE, Edition of 1915, p. 28.

CHAPTER II
FUEL

5. Introduction. The source of the world's supply of energy
is the sun. In the use of water power we are drawing, in point
of time, almost directly from the sun. On the other hand, in
the use of such fuels as coal, gas, and oil, we make use of a store
of energy that has been accumulating for ages.

While the world's supply of coal, oil, and gas is limited, we have
used but a very small part of the known deposits. The past
century has seen a marvelous change in our manner of living and
in our ways of thinking. Our vast commercial system with its
perplexing problems has arisen during the past few generations.
One of the chief causes of this great change is our ability to util-
ize the vast stores of energy to be found in nature. The steam-
engine has been the chief means by which the energy stored in
our coal deposits has been tapped and forced to do the work of
man. What will become of this modern civilization of ours when
the fuel supply upon which it is based is exhausted, is an inter-
esting problem of the future. Already conservationists are call-
ing to us to stop the great waste of our natural resources.

Of all our fuels, coal is the most important. Coal is the re-
mains of vegetable matter deposited in remote geological ages.
It is well known that wood rots but little when kept under water.
If the water be fresh, -the wood is not eaten by the teredo worm
or other forms of aquatic life, and will be kept in a fair state of
preservation for thousands of years. If tree trunks and other
vegetable matter fall into a fresh-water swamp and are sub-
merged before they rot, and if this continues for many centuries,
there will be a great accumulation of it. Oar coal deposits are
the result of such an accumulation of vegetable matter. Under
tropical conditions accompanied by a large supply of carbon
dioxide in the atmosphere the growth was very rapid and a deep
bed would collect in a comparatively short time.

Geologists tell us that in the past, parts of the surface of the
earth have gradually risen while others have fallen. At a remote
time, the tops of our highest mountains may have been the bot-
tom of the sea. Suppose a former swamp with its accumulated
vegetable matter is now sunk, and that great quantities of silt or

FUEL 5

other material are deposited upon it. The weight of the material
above will compress the vegetable matter into a compact and
dense mass. It is also possible that it will be subject to a high
temperature, which will change its chemical composition. Vary-
ing conditions of pressure and heat are thought to be responsible
largely for the differences between the various kinds of coal.

The principal constituents of coal are carbon, hydrogen, oxygen,
nitrogen, sulphur, and refractory earths called ash. The wood-
fiber of the original vegetable matter was composed chiefly of
hydrocarbons. While under the influence of great pressure, it
has at some period of its history been subjected to considerable
heat and therefore undergone a process of destructive distiilization.
This has driven off much of the volatile matter of the original
vegetable material and left a considerable portion of uncombined
carbon, which is called fixed carbon. The remainder of the car-
bon exists in combination with hydrogen. These carbon and
hydrogen compounds are called hydrocarbons. They are easily
volatilized, and so comprise a part of the volatile matter of the
coal.

Oxygen and hydrogen are always present in coal in the form
of water. This water is volatilized, of course, when the coal is
burned. Since heat is required to evaporate and to superheat it,
water is a detriment if present in too large a quantity. A small
amount of water, however, seems to aid in the combustion of
some coals.

All coal then contains fixed carbon, volatile matter (hydrocar-
bons and water), and ash; and it may contain other substances
(e.g., sulphur). The combustible is the fixed carbon, the hydro-
carbons, and part of the sulphur that may be present. Excessive
sulphur is undesirable because it is harmful to the metal of the
boiler and the stack if moisture is present, since it may form
sulphurous or sulphuric acid; it combines with the ash to form
a fusible slag or clinker, which is commonly objectionable; and
it makes the fuel more liable to spontaneous combustion when
stored in deep piles.

Coals that have been subjected to the greatest pressure and
heat are composed mostly of fixed carbon, and contain only a
small amount of volatile hydrocarbons. Such coals are called
anthracite. Coals containing larger quantities of volatile hydro-
carbons are called bituminous. Since there is no definite divid-

6 ENGINES AND BOILERS

ing line between these two classes, but the two seemingly overlap,
the terms semi-anthracite and semi-bituminous are commonly used
to designate coals to which are ascribed certain properties of
each class.

6. Anthracite Coal. Anthracite coal contains but a small
amount of combustible volatile matter. While it is considered
better for some uses, its heating value is less than that of good
grades of bituminous coal. It burns slowly, with but a small
flame and practically no smoke. Due to its slow burning quali-
ties, a relatively large grate area is needed on which to burn it.
The supply of this coal that is easily obtained has been diminish-
ing in this country, and its demand for domestic use has greatly
increased during the past few years. This has led to a rapid in-
crease in price and a great diminution of its use for power purposes.

Anthracite is considered much superior to bituminous coal for
the production of producer gas. This is due to the fact that it is
so free from the hydrocarbons that produce tars. The formation
of tar has been the great objection to the use of bituminous coals
in the producer plant.

Average anthracite coal contains about 85% fixed carbon, 4%
volatile matter, 9% ash, and 2% water. The heat value averages
about 13000 B.t.u. per pound.

7. Bituminous Coal. Most of the coal used for power pur-
poses is bituminous. This coal contains a larger percentage of
volatile combustible matter. It burns at a lower temperature
than does anthracite, and with a much longer flame. The length
of the flame varies with the composition, some kinds being called
long-flaming and others short-flaming. The ordinary furnace is
not usually so constructed as to give the volatile hydrocarbons
a chance to be completely burned. This results in the formation
of smoke. Engineers have spent much time and study on the
prevention of smoke. Upon yielding up their volatile matter
some coals fuse and form a solid mass or cake of the nature of
coke. These are called caking coals. This action hinders the draft.
If rapid combustion is desired, the mass must be broken up.

The composition of bituminous coal varies greatly, but the
average of the better grades may be taken as 65% fixed carbon,
28% volatile combustible, 5% ash, and 2% moisture, with a heat
value of 14000 B.t.u. per pound.

FUEL 7

8. Lignite and Peat. Lignite or brown coal is high in vola-
tile combustible and also contains much moisture. The evapora-
tion of this moisture after mining causes the lignite to crumble or
slack. It is usually inferior to anthracite and bituminous coals,
but it is used where it is easily obtained and where better coal is
expensive. Abroad, lignite is often formed into briquettes.

Peat is the partly decayed remains of vegetation that accumu-
lates in bogs. While it is an inferior fuel, it is used to a consid-
erable extent abroad. It is sometimes pressed into briquettes.

9. Natural Gas. Natural gas is used to some extent for power
purposes in sections of the country within reach of the gas fields.
It contains about 90% marsh gas (CH 4 ) and has a heat value of
nearly 1000 B.t.u. per cubic foot.

10. Oil. Crude petroleum and fuel oil are used to a consid-
erable extent in parts of this country. The petroleum produced
in the eastern and middle states is of a paraffin base, while that
from Texas and California is of an asphalt base. Gasoline and
other light oils are distilled from petroleum, and the residue is
sold as fuel oil. Petroleum is composed principally of hydrogen
and carbon in the form of hydrocarbons and has a heat value of
about 20000 B.t.u. per pound.

11. Coal Fields of the United States. Our principal deposits
of anthracite coal are in eastern Pennsylvania. The deposits are
not of large extent and the best are rapidly becoming exhausted.
It is claimed that there is anthracite in Alaska. Only a small
proportion of the anthracite mined is being used for power pur-
poses, the rest going for domestic heating and like purposes.

The best of our bituminous and semi-bituminous coal is taken
from the field that includes western Pennsylvania, eastern Ohio,
a large part of West Virginia, and eastern Kentucky. Southwest-
ern Indiana and most of the state of Illinois are underlaid with coal
of a fair quality. There is also a field running north from Okla-
homa through eastern Kansas and western Missouri into Iowa.
The coal from the latter is generally of poor quality, and is used
only locally. Immense coal fields exist in southern Utah and
Colorado, and in New Mexico and Arizona. No complete survey
of these fields has been made as yet, and they have not been de-
veloped up to the present. There is also a coal and lignite field
in eastern Montana and western North Dakota.

8 ENGINES AND BOILERS

The peat beds of the country are principally in Minnesota,
Wisconsin, Michigan, New York, and the New England states.

Oil is produced in the territory occupied by the eastern coal
fields, in Kansas, Oklahoma, Texas, California, and, to a smaller
extent, elsewhere.

12. Coal Storage. In the operation of most steam-power
plants, it is essential that a constant supply of coal be available.
Owing to unsettled labor conditions at the mines and to uncer-
tain transportation facilities, it is necessary that there be some
storage capacity. With anthracite coal this is a simple problem,
but it is not so with certain grades of bituminous coal. Upon
exposure of bituminous coal to the air, there is a considerable
oxidation of the hydrocarbons with attendant heat production.
If the coal pile is large, this heat may start a fire which is costly
and hard to extinguish. The origin of such a fire is called spon-
taneous combustion. Even if fire does not start, there is a loss of
heat-value up to as high as 10% in some grades of coal. If the
moisture content is large, the weathering is accompanied by a
crumbling or slacking. Storage piles are often ventilated in order
to keep them cool. In some large plants, the storage is so ar-
ranged that it may be submerged in water. This obviates the
fire risk, and reduces the other losses to a minimum.

13. Determination of Heating Values of Fuel. In plants
where large amounts of fuel are used, it is quite common to buy
coal on the basis of its heating value. In accurate tests of power
plants it is also necessary to know the heating value of the coal
used. Care must be exercised in order to get a fair sample of the
fuel. The heating value may be determined in two ways, by
combustion in a calorimeter, or by chemical analysis.

In the calorimeter method a sample of the fuel is placed in a
steel bomb along with compressed oxygen. The bomb is placed
in a calorimeter containing water, and the fuel is ignited by means
of an electrically heated wire. Upon the firing of the fuel, heat
is given to the water. It is possible to determine from the rise
of the temperature of the water the amount of heat generated.
The value thus determined is known as the higher calorific value.
It must be remembered that it is seldom possible actually to get
this amount of heat by burning in a furnace. This is due to the
fact that the hydrogen in the fuel combines with the oxygen of

FUEL 9

the air to form water, which ordinarily passes from the furnace
in the form of steam, carrying with it the heat of vaporization
of the steam. The lower heat-value does not contain this heat
of vaporization. The higher value is the accepted standard.

There are two methods of chemical analysis, the ultimate, and
the proximate. The ultimate analysis may be made on the basis
either of moist or dry fuel. The latter is commonly accepted.
If the analysis is made on the basis of moist fuel, it may be con-
verted to the dry basis by dividing the percentage of the various
constituents by one minus the percentage of the moisture. The
ultimate analysis gives the percentage by weight of carbon, hydro-
gen, oxygen, nitrogen, sulphur, and ash. Knowing the composi-
tion of the fuel, the heat-value may be determined by a formula.
An accepted formula is a modification of that of DULONG :

B.t.u. per pound of dry fuel = 14600 C+62000 (H-O/8)-f 4000S,

in which C, H, O, and S represent the proportionate parts by
weight of carbon, hydrogen, oxygen, and sulphur. The heat-
value of pure sulphur is 4000 B.t.u. per pound, but the sulphur
in coal is mostly in a form that is noncombustible. The results
of the ultimate analysis agree closely with those obtained from
the calorimeter. The proximate analysis gives the proportion of
fixed carbon, volatile combustible, moisture, and ash. Since the
heating value of the volatile combustible is not determined, the
results are not as reliable as those of the previous method. It is
very often used, however, because it is easily made and affords a
rough comparison of various fuels.

In making the proximate analysis, a weighed sample of coal
is placed in a crucible of known weight and is kept at a tem-
perature a few degrees above the boiling point of water for an
hour. From the loss of weight, the moisture in the coal is cal-
culated. The sample is next heated in a hot flame a few minutes
with the lid on the crucible. This drives off the volatile matter,
and the difference in the new and previous weight is the amount
driven off. Now the sample is heated for two hours with the lid
off, all fixed carbon is burned out, and the weight is determined
by difference as before. The weight left is that of the ash. After
having found the percent by weight of the moisture, volatile
matter, fixed carbon, and ash, the approximate calorific value
may be found by means of the chart, shown in Fig. 3, which

ENGINES AND BOILERS

has been constructed from the values determined in a large num-
ber of accurate analyses. On this chart the heat-value in B.t.u.
per pound of combustible is plotted against the percent of fixed
carbon in the total combustible. It is assumed that the volatile
matter and the fixed carbon constitute the combustible, the ash
and the moisture being noncombustible. The method of getting
the heating value may be shown by examples.

EXAMPLE 1. Determine the calorific value in B.t.u. per pound of dry
coal having the following ultimate analysis: carbon = 75.46%, hydrogen =
3.34%, oxygen = 2.70%, nitrogen = .53%, sulphur = 2.54%, and ash = 15.43%.

SOLUTION. The B.t.u. per pound = 14600 X .7546 +62000 (.0334 - .0270/8)
+400t< .0254 = 12880 B.t.u. The oxygen calorimeter gave a value of 13000
B.t.u. per pound for this same sample.

Chart for determining Heat Value of Combustible with Different Percentages
of Fixed Carbon from Proximate Analysis

3+O0

^ / </ a/t

/3/QQ

\^ '*><*"

^ 14000

^ t+fVQ

V * r

tot

^ -14,00

14400

JO SS 6S 70 7S &0

Percent of ftxcd C arbor? m Tbta/
FIG. 3

EXAMPLE 2. Determine the calorific value of coal which has the follow-
ing proximate analysis: moisture =4. 7%, volatile matter = 24.6%, fixed
carbon =62.5%, and ash = 8.2%.

SOLUTION. The combustible being composed of the volatile matter and
fixed carbon comprises 24.6+62.5 = 87.1% of the weight of the coal. Of
this combustible the fixed carbon is 62.5/87.1 = 71.7%. From the chart
it is seen that for 71.7% the B.t.u. per pound of combustible is 15550 B.t.u.
As the combustible comprises only 87.1% of the total weight the heating
value will be .871X15550 = 13550 B.t.u. per pound of wet coal, or if reduced
to the dry basis, 13550/(1.00-. 047) = 14200 B.t.u. per pound.

FUEL 11

14. Combustion. By combustion is meant the rapid chem-
ical combination of oxygen with the carbon, hydrogen, and sul-
phur in fuel. Combustion is complete when the maximum amount
of oxygen is used in the combination. One atom of carbon will
combine with one atom of oxygen to form carbon monoxide (CO).
This is not complete combustion, however, for one atom of car-
bon will combine with two atoms of oxygen, forming carbon di-
oxide (CO2), if sufficient oxygen is present.

Since the atomic weight of oxygen is 16, and that of carbon is
12, it takes 32 pounds of oxygen to 12 pounds of carbon to form
carbon dioxide, i.e., one pound of carbon requires for its complete
oxidation 2.667 pounds of oxygen. Air, by volume, is composed
of about 21% oxygen and 79% nitrogen, and by weight, of 23.15%
oxygen and 76.85% nitrogen. So one pound of oxygen is con-
tained in 4.32 pounds of air. It therefore takes 2.667X4.32=
11.55 pounds of air for every pound of carbon burned. At ordi-
nary room temperatures one pound of air occupies about 13.4
cubic feet, so that it requires theoretically 13.4X11.55= 155 cubic
feet of air for the complete combustion of one pound of carbon.

The hydrogen in the hydrocarbons of the coal is also com-
bustible. A part of the sulphur present may be combustible,
but it is usually present in such small amounts that it will be
omitted in our present computation. As explained previously,
not all of the hydrogen content of the coal is combustible, since
part of it is already combined with oxygen in the form of water.
Therefore the available hydrogen may be expressed as (H O/8).
Since hydrogen combines with oxygen to form water, in the ratio
by weight of 1 to 8, it will require 8 pounds of oxygen to burn
each pound of hydrogen. Since one pound of oxygen is contained
in 4.32 pounds of air, it will take 8X4.32=34.6 pounds of air to
burn a pound of hydrogen. Hence the total weight of air re-
quired to burn a pound of coal to CO 2 and H 2 O is theoretically

11.55 C+34.6 (H-O/8),

where C, H and O have the same meaning as in 13.

Since the nitrogen of the air is inert, it is of no value to the
combustion. Since it passes up the stack at a higher tempera-
ture than that at which it entered the furnace, it carries away
heat. Any less air than the theoretically correct amount would
result in the formation of a mixture of carbon monoxide and

12 ENGINES AND BOILERS

carbon dioxide. The heat liberated by the formation of the carbon
monoxide is only 4450 B.t.u. per pound of carbon, while it is
nearly 14600 B.t.u. for the formation of carbon dioxide. Hence the
production of carbon monoxide in a furnace means a large loss
of heat. The presence of carbon monoxide in flue gas nearly
always indicates a large amount of unburned hydrocarbons and
hence an even greater loss of heat. If it were possible to so dis-
tribute the air that it all came in close contact with the fuel, and
also to give it time enough to combine thoroughly with the fuel,
the theoretical amount of air would be sufficient. Under actual
furnace conditions, however, it is found that 50% or more excess
of air is needed to give complete combustion of coal. A somewhat
smaller excess is needed when oil is used as a fuel, because there
is better distribution of the air. The greater the amount of air
passing through the furnace, the greater the amount of heat it
will carry along to the stack. Hence an unnecessary excess of
air is not desirable, and leads to lessened efficiency. The neces-
sary excess depends upon the conditions of draft and fire as well
as upon the kind of fuel and the type of furnace. It can only
be determined by actual test.

15. Composition of Flue Gas. As just explained, an excess
of air is needed in order to get complete combustion of the fuel.
If it were possible to get complete combustion without this excess,
our flue gas would be composed chiefly of nitrogen, carbon dioxide,
and water vapor. Due to the excess of air, there will be free
oxygen present in the flue gas. If there is an insufficient excess
of air there will also be carbon monoxide and probably some
hydrocarbons present. We have seen that the presence of carbon
monoxide indicates incomplete combustion and therefore low fur-
nace efficiency. On the other hand, a large excess of air, while
it may give complete combustion, gives poor furnace efficiency
because the air will carry a large amount of heat up the stack.
It is a matter of great importance that just the right excess of
air be admitted to the furnace. Since it is difficult to measure
directly the amount of air entering the furnace, an easier method
is used. This method consists in analyzing the flue gas to deter-
mine the amount of each of its constituents. From this analysis
we can easily compute the amount of air entering the furnace.
Knowing the composition of flue gas, we can regulate the amount

FUEL 13

of air entering the furnace so as to give the proper excess to insure
the best economy of operation.

16. Flue Gas Analysis. There are various types of apparatus
on the market for making the analysis of flue gas, most of which
are modifications of the apparatus designed by ORSAT. A com-
plete description of the Orsat apparatus will not be given here,
but the principle of its operation is as follows.

A sample of gas is taken from the rear of the furnace or between
the furnace and the stack. After being cooled to the room tem-
perature, it is carefully measured by volume at atmospheric
pressure. All measurements are taken at room temperature and
at atmospheric pressure. This known volume of our sample is
first passed a few times through a solution of caustic potash, which
absorbs the carbon dioxide. The volume is measured again, and
the difference between the new volume and the original volume
is the volume of the carbon dioxide absorbed. The same sample
is next passed several times through a solution of potassium pyro-
gallate, which absorbs the oxygen. The amount of oxygen is
determined by the loss in volume, as before. Next the sample
is passed several times through a solution of acid cuprous chloride
and the carbon monoxide removed, and its amount determined
as before. The amount of carbon monoxide is usually quite small.
The remainder of the sample is usually assumed to be nitrogen^

17. Heat Lost in Flue Gas. The weight of flue gas per pound
of fuel burned (assumed carbon and ash) may be computed from
the formula,

where

W = weight of flue gas per pound of fuel burned.
C = decimal part by weight of total carbon in fuel.
N = percentage by volume of nitrogen in flue gas.
CC>2 = percentage by volume of carbon dioxide in flue gas.
CO = percentage by volume of carbon monoxide in flue gas.

A = decimal part by weight of ash in fuel as fired.

Unless the ultimate analysis of the fuel is known, the weight

of carbon in the volatile matter will have to be estimated and

added to the weight of fixed carbon to give C in the preceding

formula. Marks has published a chart showing the approximate

ENGINES AND BOILERS

relation between the volatile carbon in the combustible and the
total volatile matter in it. With the aid of this chart (Fig. 4),
the value for C in the preceding formula can be approximated
from the proximate analysis. The specific heat of the flue gas
is usually taken as .24; and the heat lost per pound of fuel
burned is equal to the product of the specific heat of flue gas, the
weight of gas per pound of fuel, and the difference in temperature
between the leaving flue gas and the entering air.

Chart for determining the Carbon in the Volatile Matter

Marks

10 20 30 -9-0 SO t

Percent of Yo/otifa Afatt-rr jr? ffye Combust/hie

FIG. 4

EXAMPLE. How much heat is carried up the stack by the dry flue gases,
when the furnace is burning coal of the following proximate analysis? Mois-
ture = 3%, fixed carbon = 65%, volatile matter = 26%, and ash=6%. The
analysis of flue gases gives: CO 2 = 10%, O = 8%, and CO = .5%. The stack
temperature is 500 F. and the temperature of the air entering the furnace
is 80 F.

SOLUTION. From the chart of Figure 4, we find that the percent ofjvoh
tile carbon in the combustible is about 13.5, which corresponds
fuel. The total carbon is then 12.1+65 = 77.1%, and the weighiToT flue
gases per pound of coal is

3.032 X. 771

--6) = 19 - 1 Pounds.

The heat carried up the stack by the dry flue gases is then .24X19.1 (500
80) = 1925 B.t.u. for each pound of coal fired. In this problem, the heat-
value of a pound of coal found from the proximate analysis is 13860 B.t.u.
Hence the loss is 1925/13860 = 13.9% of the heat available.

FUEL 15

Both the fuel and the air contain moisture. This moisture also carries
heat up the stack, since it is a superheated vapor upon leaving the furnace.

18. Value of CO 2 for Best Efficiency. As has been stated,
the efficiency of the furnace will vary with the excess of air ad-
mitted. Since the percentage of CO 2 also varies with the excess
of air, we see that an indication of the efficiency .is given by the
CO 2 reading. Just what percentage of CO 2 corresponds to the
highest operating efficiency depends upon such factors as kind
and state of fuel, stack temperature, etc. After the proper
amount of C0 2 for best efficiency has been determined under
these conditions, the CO 2 reading will indicate whether or not
high efficiency is being obtained. Since the determination of the
CO 2 is a comparatively simple operation, it is an excellent way to
keep check on operating conditions. Some plants go so far as to
pay their firemen on the basis of the CO 2 record. In general,
high CO 2 means high efficiency, unless there is some abnormal
condition such as too much CO. The CO should be kept as near
zero as possible.

At the same temperature and pressure, CO 2 occupies the same
volume as the oxygen from which it was formed. The volume of
the oxygen in the air is 21%. Hence, if the products of combus-
tion are cooled down to the temperature of the entering air, the
CO 2 reading would be 21% for perfect combustion with no excess
of air, assuming the fuel to be carbon and ash. In practice, the
CO 2 runs 17% or lower. Even 15% is usually considered an in-
dication of very good efficiency.

19. CO 2 Recorders. Automatic devices are on the market
that will analyze and record the amount of CO 2 on a chart. An
analysis is made every few minutes, so that a complete record is
kept of the operating efficiency.

CHAPTER III

STEAM

20. Introduction. Definitions. A perfect gas may be at any
temperature under any pressure. For instance, air may be placed
under a certain pressure and have its temperature raised or low-
ered by the addition or subtraction of heat. On the other hand,
a saturated vapor, such as steam, can exist only at a certain
definite temperature for each particular pressure. Under ordinary
atmospheric pressure, saturated steam can exist only at a temper-
ature of about 212 F. Under an absolute pressure of 100 pounds
per square inch, saturated steam will be at a temperature of
327.86 F.

Let us consider a case in which a pound of water at 32 F. is
placed under a pressure of, say, 100 pounds per square inch.
The containing vessel is supposed to be so constructed that the
pressure remains constant, no matter what change of volume
takes place. Now suppose that the water is heated. There will
be little change in volume, but there will be a rise of temperature
of approximately one degree F. for each B.t.u. given to the water.
This will continue until we have added 298.5 B.t.u. The tem-
perature will then be 327.86 F. A further addition of heat up
to a limit will not cause any change of temperature, but will effect
a change in the physical condition of the water, turning it to
steam. We shall need to apply 887.6 B.t.u. to effect this change
completely. We have now added a total of 298.5+887.6 = 1186.1
B.t.u., and have converted the pound of water, originally at 32 F.,
into dry saturated steam at a temperature of 327.86 F., and
under an absolute pressure of 100 pounds per square inch.

Now that the water is all evaporated, if more heat be added
to this steam, the temperature will rise at the rate of nearly two
degrees per B.t.u. added. We now have superheated steam.

As stated previously, the volume of the water will change but
little until the boiling point is reached. The space occupied by
the saturated steam will be 4.432 cubic feet. This is many times
greater than the space formerly occupied by the water. Part
of the 887.6 B.t.u. that was used to evaporate the water was
evidently used to cause this change in volume under the pressure

STEAM

of 100 pounds per square inch. The remainder was used to make
the physical change in the water, to increase the kinetic energy
of its atoms. For pressures other than 100 pounds we would
have values different from those given above.

Figure 5 represents graphically the relation between the tem-
perature and the heat added to a pound of ice, starting at zero
degrees F. (with the assumption that the pressure is constant).
Upon the first addition of heat, the temperature of the ice will
rise until the melting point is reached. Further addition of heat

O* 32

Temperature

FIG. 5

causes the ice to melt. This occurs without a change of tempera-
ture. The part of the line representing the melting of the ice
is therefore vertical. When the ice is all melted, addition of more
heat causes the temperature of the water to rise. This will con-
tinue until the boiling point is reached. That part of the line
representing evaporation will be vertical since there is no change
of temperature during that period. When evaporation is com-
plete, the addition cf heat again causes a rise in temperature.

The amount of heat necessary to raise the temperature of the
water, the amount of heat required to change it to steam, and
also the volumes of steam formed under different pressures, have
been determined by numerous experiments and are published

18 ENGINES AND BOILERS

under the name of steam tables. We shall use in our work the
tables prepared by C. H. Peabody. 1 These are arranged in two
ways. In Table I, the various absolute pressures at which water
boils are given for each degree F. from 32 to 428. In Table II,
the various temperatures at which water boils are given for each
pound per square inch from 1 to 336. The values are arranged
in two tables not because they are different, but simply as a con-
venience in their use.

In Table I, the first column, headed t, gives the temperature
at which water boils. The second column, headed p, gives the
absolute pressure under which it must be in order that it boil
at the temperature given in the first column.

The third column, headed q, gives the heat of the liquid, which
is the number of B.t.u. necessary to change the temperature of one
pound of water from 32 F. to the temperature given in the first column.
This does not mean that there is no heat in the water at 32.
'The heat in the water below the freezing point is of no moment
to the steam engineer; hence it is chosen as the arbitrary starting
point.

Column four, headed r, gives the heat of vaporization, which is
the B.t.u. necessary to evaporate completely a pound of water at the
temperature and pressure given in the first and second columns.
This is sometimes called the latent heat of evaporation. The
sum of the heat of the liquid and the heat of vaporization is
called the total heat.

The fifth column, headed p, is that part of the heat of vaporiza-
tion that is used in energizing the atoms of the water to turn it to
a vapor; it is called the heat equivalent of internal work.

The sixth column, headed Apu, is the rest of the heat of vaporization,
or that port that is needed to do the work of increasing the volume,
under the pressure of column two; it is called the heat equivalent
of external work.

Columns seven and eight will not be discussed here. Column
nine, headed s, gives in cubic feet the specific volume, which is
the volume of one pound of dry saturated steam under the pressure
of column two.

Column ten gives the reciprocals of the values found in column
nine. It is the weight of one cubic foot of dry saturated steam under
the pressure of column two.

i C. H. PEABODY, Steam Tables.

STEAM

Steam generated in most boilers not equipped with a super-
heater is likely to carry with it, when leaving through the outlet
pipe, a small amount of water in a finely divided state or mist.
Steam containing this moisture is said to be wet steam. The

Chart Showing Specific Heat of Superheated Steam Values from
Knoblauch and Jakob

700

zoo

SO /OO /JO ZOO 250

Pressure In pounds per square /nch ffhso/ute
FIG. 6.

quality of wet steam is expressed in percent. If in a hundred
parts by weight of a mixture of steam and water, five parts by
weight are moisture, the quality of the mixture is said to be 95%
and the priming 5%.

As long as steam is in contact with water it will remain satu-
rated, and its temperature cannot be raised under constant pres-

20 ENGINES AND BOILERS

sure. If it is conducted away from the water and led to a super-
heater, its temperature will be raised by the addition of heat.
It is then superheated steam. The amount of heat necessary to
superheat depends upon the pressure and upon the degree of
superheat. The chart of Fig. 6 gives the specific heat of super-
heated steam for the ranges of pressure and temperature com-
monly found in practice. The specific heat of steam varies with
both temperature and pressure. The chart gives the average
values of specific heat as the steam is raised from the temperature
of saturation to the temperature of superheat.

EXAMPLE 1. How much heat is required to change a pound of water at
70 F. into dry saturated steam at a pressure of 120 pounds per square inch
absolute?

SOLUTION. On page 48 of Peabody's Steam Tables, we find that the heat
of the liquid, q, for 120 pounds pressure is 312.3 B.t.u. This amount of heat
would bring the temperature of the water from 32 F. to the boiling point.
As the temperature of the water to start with is 70 (p. 36), it already con-
tains 38.1 B.t.u. It is then necessary to add to it 312.3-38.1=274.2 B.t.u.
in order to bring it to the boiling point. To evaporate the water requires
the heat of vaporization, r, at 120 pounds (p. 48), which is 876.9. Hence
the total heat required to bring the water up to boiling and to evaporate it
is 274.2+876.9 = 1151.1 B.t.u.

EXAMPLE 2. If, in Example 1, the quality of the steam formed had been
97%, how much heat would it have required?

SOLUTION. The water must all be brought to the boiling point, which
takes the same amount of heat as in Example 1, 274.2 B.t.u. As the quality
is 97%, only .97X876.9 = 850.6 B.t.u. are needed to evaporate the water.
Hence the total amount of heat required is 274.2+850.6 = 1124.8 B.t.u.

EXAMPLE 3. Find the amount of heat necessary to generate the pound
of steam in Example 1, if it is superheated to a temperature of 475 F.

SOLUTION. To generate a pound of dry saturated steam under the con-
ditions of Example 1, requires 1151.1 B.t.u. The temperature correspond-
ing to 120 pounds pressure is 341.3 F. The superheat is then 475 -341.3 =
133.7. From the chart of Fig. 6, the specific heat of superheated steam is
.537. It will take .537X133.7 = 71.8 B.t.u. to superheat the steam. Hence
the total heat required is 1151.1+71.8 = 1222.9 B.t.u.

EXAMPLE 4. Find the volume of the pound of steam in Example 2.

SOLUTION. A pound of dry saturated steam at 120 pounds pressure occu-
pies 3.723 cubic feet. The quality being 97%, the volume occupied by the
steam is .97X3.723 = 3.611 cubic feet. A pound of water occupies .016
cubic feet, and, as 3% of the pound of wet steam is water, the volume of the
water is .03 X .016 = .0005 cubic feet. Hence the total volume is 3.61 1 + .0005 =
3.611 cubic feet.

STEAM

21. The Steam Calorimeter. In making tests of boilers or
engines it is necessary to know the quality of steam leaving the
one or entering the other. A steam calorimeter is used in mak-
ing this determination. Several types of calorimeters are in use.
If the quality of steam is high (between 94% and 100%), the
throttling type is usually used. When properly constructed this
calorimeter is sufficiently accurate for ordinary purposes.

Figure 7 shows a throttling calorimeter attached to a steam-
pipe A. If the steam is saturated, and the pressure is known
from a pressure gage H, its temperature may be determined from
the steam tables. If the steam in A is superheated, it is also
necessary to take its temperature by means of a thermometer,
and the heat contents may be calculated from the steam tables.
If it is wet, the quality must be known to find its heat contents.
The tube B in Fig. 7 is a sampling tube through which a sample
of steam is taken from the pipe A. This sample is throttled down

22 ENGINES AND BOILERS

in pressure at C from the pressure in A, say pi, to the pressure
in the calorimeter G, say p%. If the calorimeter is well covered,
but little heat is lost by radiation and the heat contents in one
pound of steam in A is the same as in the chamber G. The pres-
sure and temperature in G are measured by the gage F and the
thermometer E, and the heat contents are computed by means
of the steam tables.

Since the heat content is the same per pound in A as in G,
the quality in the former may be computed as follows. The total
heat in A equals q\-\-xri where x is the quality and qi and r\ are
the heat of liquid and the heat of vaporization in A, respectively.
If the calorimeter is working properly, the steam will be super-
heated in G, and its heat contents will be equal to

where q% and r% are the heat of the liquid and the heat of vaporiza-
tion in G, respectively, 3 is the temperature of steam in G as
measured by the thermometer E, and k is the temperature of
saturated steam at the pressure p%. The term .48 (3 2) is the
heat of superheat in G, since .48 is the specific heat of super-
heated steam at low pressure and temperature. Then we have

from which x may be found.

EXAMPLE. Find the quality of steam leaving a boiler when the pressure

is 165 pounds gage. The gage pressure in the throttling calorimeter is 3

founds, the temperature is 265 F., and the barometer reading is 29.6 inches.

SOLUTION. From the steam tables, q\ =345, n =851, <? 2 = 189, r 2 = 964, and

* 2 = 221. Then

345+z851 = 189+964+.48(265-221),

from which x = .975 or 97.5%.

CHAPTER IV
BOILERS

22. Introduction. The stored energy in fuels is utilized by
means of the steam engine as follows. The fuel is burned in a
furnace, resulting in a mixture of heated gases. These hot gases
pass over and along the surface of a boiler. A transfer of heat
takes place through those boiler surfaces that are exposed to
contact with the hot gases or to radiation from the incandescent
fuel bed on the one side and water or steam on the other side.
This boiler surface is called heating surface. This heat is con-
ducted through the shell of the boiler and is spent in heating and
evaporating the water contained in the boiler. The steam thus
formed is conducted from the boiler to the engine or turbine,
where it does work due to its pressure and to its tendency to
expand.

Practically all boilers have a considerable storage of heated
water and steam. This water and steam is under a high pres-
sure and would increase in volume hundreds of times if the pres-
sure were removed. A sudden release of this pressure causes
an explosion. Many lives have been lost and a great amount of
property destroyed by boiler explosions. Hence the consideration
of first importance in a boiler is its safety. Other considerations
are its first cost, its life, and the ease with which it may be cleaned
and repaired. In portable boilers and marine boilers, weight and
the space occupied are of great importance.

Since the purpose of the heating surface is to conduct heat
from the furnace to the water, it follows that the conduction
should be rapid and effective. To be efficient, the boiler must
extract a large proportion of the heat generated in the furnace.
The surface must be of such size and so arranged that time is
allowed to render this transfer as complete as possible.

Due to the erosion of some of the parts, or due to overheating
consequent on the formation of scale, a boiler originally fit for a
certain class of service may become so weakened that it is unsafe
for high pressures. Many states provide for an inspection of
boilers and equipment in order to prevent explosions. The vari-
ous boiler insurance companies also make periodic inspection of
insured boilers. As a result of this inspection, the inspector sets

24 ENGINES AND BOILERS

a limit to the pressure that the boiler may carry. The fireman
must be constantly on the alert to detect any signs of a develop-
ing weakness.

Much of the water used in boilers will deposit scale on the heat-
ing surface of the boiler. This scale greatly hinders the conduc-
tion of heat to the water and may even become thick enough
to allow the metal on which it settles to become overheated to
such an extent that it gives way and causes an explosion. If the
scale becomes too thick, it must be removed. The removal and
prevention of scale will be considered later.

The fire side of the heating surface in many boilers will collect
soot and fine ash. To maintain efficient steaming these surfaces
must be kept clean. The soot should be swept or blown from the
surface as fast as it collects.

In the construction of a steam-boiler, the following require-
ments are to be considered.

(1) Proper materials of uniform strength and reliability must
be employed, and the size of all parts must be so designed that
a sufficient factor of safety exists. The workmanship must be
good.

(2) There must be steam space and water capacity such that a
sudden change of load will not cause an undue drop in steam
pressure.

(3) There must be a sufficient water surface to allow for com-
plete separation of the steam from the water. Too small a sur-
face will result in foaming.

(4) A thorough circulation of the water must be provided, so
that a uniform temperature is maintained throughout the boiler.
Water is a very poor conductor of heat, and it is therefore essen-
tial that there be a continuous flow over the heating surface.

(5) Stresses due to temperature change must be eliminated.

(6) In so far as possible, all joints or seams must be protected
from the direct action of the flame.

(7) Access must be possible to all parts for cleaning and repair.

(8) A means for the discharge of mud or sludge that is left by
the evaporation of the water must be provided.

The modern steam boiler is the result of an evolution starting
with a vessel much resembling a closed kettle. The pressure in the
early boilers was nearly atmospheric ; hence the shape was not
influenced by the consideration of the strength of the boiler.

BOILERS 25

With the use of steam under pressure, boilers assumed a cylin-
drical or spherical shape, since these shapes are not distorted so
much by internal pressure. The simplest boiler is cylindrical,
with hemispherical ends.

For a given steam pressure, the thickness of metal required
varies directly as the diameter. Hence heavy plate must be used
for boilers of considerable size. Moreover, the ratio between the
heating surface and the volume decreases as the diameter in-
creases. Thus it is seen that the single cylinder is suitable for
small boilers only.

From this early and simple type of boiler the development
has proceeded along two distinct lines. First, in place of a single
large cylinder, several smaller ones were sometimes used, thereby
decreasing the weight of metal and increasing the rate of steam-
ing. Carrying this idea still further results in a large number of
very small cylinders or tubes filled with water and surrounded by
fire. This is the modern water-tube boiler.

The other way of increasing the heating surface of a cylindrical
boiler is to run smoke flues through it. The earliest types con-
tain one or two large flues. The modern type contains a large
number of small tubes through which the fire or the products of
combustion pass. This type is known as the fire-tube boiler.

23. Rated Horsepower. The rating of a boiler is usually
based on its heating surface. There is no standard for this rat-
ing. It is becoming general practice, however, to rate a water-
tube boiler on a basis of 10 square feet of heating surface per boiler
horsepower, and to rate a fire-tube boiler at 11 or 12 square feet
per boiler horsepower. Under average conditions of firing, care,
and draft, a boiler should develop good economy at its rated
horsepower. However, many boilers are able to carry great over-
load and still give excellent efficiency. Poor efficiency is due in
general more to overloading the furnace than to increasing the
evaporation from the boiler.

24. Heating Surface. It has been customary to consider as
heating surfaces those surfaces which have water on one side and
the products of combustion on the other side. No distinction
is made in the thickness of metal in different parts of the boiler,
or in the difference in temperature of the gases on their path along
the heating surface. However, great accuracy is seldom required

26 ENGINES AND BOILERS

in computing the heating surface of a boiler. In calculating
heating surface, the inside of fire tubes and the outside of water
tubes is used.

25. Rules for Finding the Heating Surface.

1. Horizontal Return-tubular boilers. Kent l gives the fol-
lowing rule:

Take the dimensions in inches. Multiply two-thirds of the
circumference of the shell by its length; multiply the sum of
the circumferences of all the tubes by their common length;
to the sum of these products add two-thirds the area of both
tube sheets; from this sum subtract twice the area of all the
tubes; divide the remainder by 144 to obtain the area in square
feet.

2. Vertical Tubular boilers. Kent 2 gives the following rule:
Multiply the circumference of the fire-box (in inches) by its

height above the grate; multiply the combined circumference
of all the tubes by their length, and to these two products add
the area of the lower tube sheet; from this sum subtract the
area of all the tubes, and divide by 144 : the quotient is the
number of square feet of heating surface.

3. General rule. The U. S. Bureau of Mines 3 gives the fol-
lowing rule:

A short approximate method for any boiler is to figure the
square feet of heating surface in the tubes and divide it by 0.85
for a return tubular boiler or by 0.95 for a water tube boiler.
In case the return tubular boiler has an arch over the top for
gas passage, giving the so-called third return, it is necessary
to add from 100 to 200 square feet to the result to obtain the
total heating surface.

In this last rule the heating surface in fire-tube boilers is figured
from the outside diameter of tubes.

26. Superheating Surface. In modern practice, steam is often
led off from the main steam space and taken through other heat-
ing coils. Since there is no water in contact with this steam, it

1 KENT, Mechanical Engineers' Handbook, 1916 Edition, p. 888.

a Ibid.

8 BUREAU OF MINES, U. S. DEPARTMENT OF THE INTERIOR, Bulletin No. 40, p. 9.

BOILERS

will be superheated. That surface which has this superheated
steam on one side and fire or hot gases on the other is called the
superheating surface.

27. Size of Boiler Tubes. The outside diameter is the nom-
inal diameter in boiler tubes. The following table given by the

5 team -pipe
Cffnnecfion

FIG. 8

National Tube Works gives the nominal or outside diameter, the
inside diameter, and the thickness, of standard lap-welded boiler
tubes.

SIZE IN INCHES OF STANDARD LAP-WELDED TUBES

External Diameter

1.0
.810
.095
3.25
3.010
.120

1.25

1.060
.095
3.5
3.260
.120

1.5
1.310
.095
3.75
3.510
.120

1.75
1.560
.095
4.0
3.732
.134

2.0
1.810
.095
4.5
4.232
.134

2.25
2.060
.095
5.0
4.704
.148

2.5

2.282
.109
6.0
5.670
.165

2.75
2.532
.109
7.0
6.670
.165

3.0
2.782
.109
8.0
7.675
.165

Internal Diameter
Standard Thickness
External Diameter
Internal Diameter
Standard Thickness

28 ENGINES AND BOILERS

28. Water-tube Boiler. Babcock and Wilcox Type. One of
the most common forms of water-tube boilers is the Babcock and
Wilcox type, shown in Fig. 8. Here the tubes are fastened into
headers, which in turn are connected by other tubes to a forged
steel cross-box that is riveted to the steam drum above, as shown
in Fig. 9. Headers are made of cast-iron for low-pressure work,
and of wrought steel for high-pressure work. They are of two
types, vertical and inclined. In the latter the tubes enter the
header at right angles. The tubes between the front and rear
headers are inclined so that a circulation of water is insured.
Opposite the end of each tube in the header is placed a hand-
hole to permit cleaning and access to the tube. These holes are
closed by means of a cap. This arrangement also allows ease in
cleaning the scale from the tubes, as the caps are easily removed
and a cleaner run through the tube. Clean-out doors are placed
in the setting opposite the headers to give easy access for cleaning.
The grates are so located, and fire-brick partitions or baffles
are so placed, that the hot gases usually pass across the tubes
^ three times on their way to the stack. The mud
| | drum is connected to the lower end of the back

header. This collects the sediment that is formed
during the evaporation of the water. This sediment
is blown off from time to time through the blow-off
pipe shown at the lower right hand of Fig. 8.

The feedwater is brought in at the front of the
steam drum, and is so delivered that its velocity
will aid in circulation of the water in the boiler.
The circulation is back through the steam drum,
down the tubes to the back header, where it is distributed to the
inclined tubes, up these to the front header, and thence to the
front of the steam drum. The steam is taken off through a dry-
pipe located at the top of the steam drum. If a superheater is
attached, as shown in Fig. 8, the steam is led from the dry-pipe at
the top of the steam drum into and through the superheater, and
then up to the steam pipe shown at the top of the steam drum.
Provision is made for flooding the superheater during the period
of getting up steam. This keeps the surface cool enough to
prevent its burning out. The boiler is carried by slings from
horizontal girders placed above it. The whole boiler is surrounded
by a smoke-tight brick setting.

BOILERS

A similar form, but using boiler plate headers instead of cast
or forged sectional headers, is called the Heine type.

29. Water-tube Boiler. Stirling Type. A common form of
water-tube boiler is- shown in Fig. 10. Here banks of tubes lead

feed water inlet.

Sfatmotrffef

FIG. 10

from the lower drum, or mud drum, to the three drums above.
The water is taken in at the rear upper drum and the steam is
drawn off at the top of the middle drum. Arched tubes connect
the steam space of the upper drums. Access to the inside of the
drums, for the purpose of expanding the tubes, for cleaning, and

ENGINES AND BOILERS

for inspection, is had by means of a manhole located at the end
of each drum. The sediment collects and is blown off at the
bottom of the mud drum through the blow off pipe and valve
shown at the lower left hand corner of Fig. 10. The baffling is

J/easnov/kf

connect/em

\3mo/ie out/rt

FIG. 11

so arranged that the products of combustion are forced to travel
practically the entire length of all the banks of tubes, up the first
bank, down the second and up the third. Clean-out doors for
the removal of soot are located at various points shown in Fig. 10.

30. Water-tube Boiler. Vertical Type. A Wickes vertical
water-tube boiler is shown in Fig. 11. In this type of boiler the
tubes are straight and are placed vertically. They are connected
to a mud drum below and to a steam drum above. The products

BOILERS 31

of combustion pass up along the front half of the tubes and down
the back half, leaving at the rear.

31. Fire-tube Boiler. Externally Fired, Return-tubular Type.
- In this country, where moderate pressures are wanted, the
common type of fire-tube boiler is the externally fired return-
tubular boiler. Figure 12 shows the construction of a boiler of
this type. It consists of a cylindrical shell, made of steel or
wrought-iron plates rolled into a cylindrical shape and riveted
together. The ends or heads are formed from flat circular plates
flanged around the outer edge and riveted to the cylindrical
shell. A large number of fire-tubes extend from one end-sheet
to the other. They occupy the lower two-thirds of the shell,
the top third being steam space. The flat plates or tube-sheets
tend to bulge outward with the internal pressure. To prevent
this, the part of the sheets above and below the level of the tubes
must be stayed. These stays are of two kinds, "through stays,"
and "diagonal stays." The former are steel rods that run the
entire length of the shell, pierce the tube-sheets and hold the
sheets in by means of nuts. The latter run from each tube-sheet
diagonally back to the shell. In certain types of large boilers,
some of the tubes are made heavier and are threaded on the ends.
These tubes are secured to the end-sheets by means of nuts on
the outside. Such tubes are called stay-tubes. In general, the
tubes act as stays, and that part of the ends occupied by the tubes
needs little if any extra staying. The tubes are expanded into
the sheet and their ends are beaded over.

The feed-pipe enters the front end of the boiler just below the
water line, and the steam leaves by the dry-pipe, which leads
out of the top of the shell. The mud is blown off through an
outlet at the bottom near the rear. If the feedwater is taken
from a pond or stream and contains much vegetable matter,
there should be a surface blow-off to skim the scum that forms
on top of the water. A manhole is located at the top near the
rear and a hand-hole in front beneath the tubes. The grate is
put beneath the front end of the shell. The products of com-
bustion pass over the bridge wall, back along the bottom of the
shell, enter the tubes from the rear, pass through them and out
of the uptake at the front of the boiler or through a flue over
the top of the boiler to an uptake at the rear.

ENGINES AND BOILERS

r-J

1 i!

i u t ,

h
n
1 1

L
r

r--jp

ri' ^ if

.-,- J J LH

1 i
i
i

-1-

r
P-.

"Hi r

. r a'J tfc

~ J

J f
i.

1
H

c ;

BOILERS

The shell is supported by brackets which are riveted to the
shell and rest on the brickwork of the setting. The rear bearing
is fitted with rollers to allow the boiler to expand without crack-
ing the setting.

32. Fire-tube Boiler. Internally Fired, Return-tubular Type.
Instead of having the fire outside the shell, as in the preced-
ing boiler, the fire is sometimes placed inside one or more large

Jftrahe out/ft

ooooooooc
oooooooo
ooooooooooc
ooooooooo
ooooooooooc
oooooo

FIG. 13a

fire-flues running the whole length of boiler. Figure 13 shows a
Springfield boiler of this type. These large fire-flues are of course
subjected to external pressure, and therefore have a tendency to
collapse. In order to prevent this, they must be stiffened or
stayed in some way. This is ordinarily done by rolling them in
corrugations. If the flue is braced in such a manner that the
metal at any one point is very thick, there is danger of its being
overheated, since the gases are hottest in this combustion flue.

The fire is at the front of the flue. The gases pass back through
the flue to the rear of the boiler, into the combustion chamber.
The heating surface is mostly composed of the comparatively
thin flue and tubes. The thick outer shell is not subjected to

ENGINES AND BOILERS

the high temperatures of the combustion chamber, as in the
externally fired boiler.

33. Fire-tube Boiler. Scotch Marine Type. While water-
tube boilers are used to some extent in marine service, the stand-
ard boiler is the Scotch marine type. This is very similar to the
internally fired boiler described in 32, which is also sometimes
called a Scotch marine boiler. It commonly has three combus-

Snsafji o/oen t*

^ On/pifJ^TT^^^' '

ta^X^VXK^VXVXV^V^ I Vi^S^>

FIG. 13b

tion flues, each of which has its own set of tubes. In marine
practice the combustion chamber at the rear of the flue and the
tubes is internal, there being a water-leg between the combustion
chamber and the rear head. There is a combustion chamber for
each flue and its set of tubes. Since the surfaces of the combus-
tion chamber are flat, they must be stayed.

These boilers are large in diameter; the outer shell must there-
fore be very thick. The longitudinal seams are usually triple
riveted, with two-strap butt joints. The Scotch marine boiler is
used to some extent on land, but as the space occupied is usually
an element of less importance here, the type is not common.

BOILERS

34. Fire-tube Boiler. Vertical Type. A boiler often used
for small or portable plants is the vertical fire-tube type (Fig. 14).
These boilers are internally fired, the fire being enclosed in the
lower part of the shell, and surrounded by an annular water-ring
or water-leg. The lower tube-
sheet is placed but a small

distance above the grate.
Therefore the space for com-
bustion is very limited. The
tubes are vertical and are ex-
panded into the lower and the
upper tube-sheets.

,/ connecfioas

35. Fire-tube Boiler. Loco-
motive Type. The type of
boiler used on locomotives,
and also often used on portable
plants, is shown in Fig. 15.
In this boiler, the fire-box is
at the rear end of the shell,
and its top and sides are water-
heating surfaces.

Since the sheets that form
the water-legs at the sides and
rear of the fire-box are flat, it
is necessary to stay them to

prevent distortion. A screw-stay is used for this purpose; it con-
sists of a threaded bolt screwed through the parallel plates. The
threads on the center part of these stay-bolts are removed so
that cracks will not start in the bolt at the root of the thread.
On what are called safety stays a small hole is drilled in from
the end so that a cracked bolt will leak steam and give warning.

The flat or arched sheet at the top of the fire-box is called the
crown-sheet. The crown-sheet is stayed in various ways, some-
times by radial stay-bolts which run between it and the outer
shell, or by sling stays, which are girders slung from the outer
shell. Fire-tubes extend from the tube-sheet at the front of the
fire-box to the tube-sheet at the front end of the boiler.

The tubes in locomotive boilers are smaller than in the types
previously described, and are placed as close together as good

FIG. 14

ENGINES AND BOILERS

BOILERS 37

circulation of the water will permit. By making the tubes small
and numerous, a large heating surface is obtained. Where a
superheater is used, as shown in Fig. 15, some of the tubes are
made larger and contain the superheating surface. This super-
heating surface is formed by tubes that extend into the fire-tubes
from the front end of the boiler and run to within a short dis-
tance of the fire-box.

The outer shell extends beyond the front tube-plate to form
a smoke-box. In this smoke-box vertical nozzles are located,
through which the exhaust steam from the engines escapes. This
induces a strong draft that allows a very rapid rate of combus-
tion. The rate of combustion often exceeds 100 pounds of coal
per square foot of grate surface per hour.

Since the steaming of this type of boiler is very rapid, the
steam is taken from a steam-dome located on the top of the
shell. This allows the steam to be taken at a distance from the
water surface, thereby insuring fairly dry steam. The throttle
valve for the engine is located in the dome.

In smaller locomotive boilers, the fire-box is set between the
rear drivers. In the larger sizes, this arrangement will not allow
a large enough grate area, and so the fire-box is extended later-
ally over low trailer wheels. Manholes and hand-holes are em-
ployed to give access for cleaning, as in other types of fire-tube
boilers.

36. Superheaters. During the past few years superheated
steam has come into quite general use, especially if it is to be
used in steam turbines. The amount of superheat used is gener-
ally not large, usually between 100 and 200 Fahrenheit. The
advantages gained by the use of superheated over saturated
steam will be considered later.

There are two types of superheater. One type is independ-
ently fired. The other is formed by the addition of some super-
heating surface to the main boiler. The latter is the more com-
mon form. In this type the steam is taken from the steam space
and led through superheating coils. Often provision is made for
the flooding of these coils during the period in which steam is
being raised, in order to prevent the coils from burning out. The
boiler shown in Fig. 8 has a common form of superheater at-
tached.

38 ENGINES AND BOILERS

37. Horsepower of Boilers. As explained previously, boilers
are generally rated by the manufacturer on the amount of their
heating surface. The rate at which a boiler is working, how-
ever, must be determined from a consideration of the amount of
steam that is being generated.

The amount of heat necessary to evaporate a given quantity
of water varies with the temperature of the feedwater, with the
pressure at which the steam is formed, and with the quality of
the steam produced. Hence it is desirable that there be a stand-
ard of temperature and pressure at which we can find the equiv-
alent amount of water evaporated, using the same amount of
heat as is used under the actual conditions of temperature and
pressure. The conditions of temperature and pressure set by
the A. S. M. E. 1 as a standard are 212 F. and 14.7 pounds per
square inch absolute. The equivalent evaporation is then the
amount of water that would be evaporated from and at 212 F.
if the same amount of heat were used in its evaporation as is
used in the evaporation under the actual working conditions.
From the steam tables, the B.t.u. required to evaporate one
pound of water from and at 212 F. is seen to be 969.9 B.t.u.
This is the unit of evaporation.

At the time when it first became necessary to rate boilers, a
good engine used about 30 pounds of steam per horsepower per
hour at a pressure of 70 pounds gage. The judges at the Cen-
tennial Exposition in 1876 awarded prizes using the following
unit as a standard. A one-horsepower boiler is one that will evapo-
rate 30 pounds of water per hour from feedwater at 100 F. into
steam at 70 pounds pressure by gage.

The A. S. M. E. has since adopted an equivalent standard and
defines a boiler horsepower to be the evaporation of 34.5 pounds
of water per hour at 212 F. into steam at 212 F. and at a pressure
of 14.7 pounds absolute pressure. As the heat of vaporization
at 212 F. or 14.7 pounds absolute is 969.7 B.t.u., it therefore
takes 969.7X34.5 B.t.u. per hour for one boiler horsepower.

To determine the horsepower at which a boiler is working, we
must therefore first find how many B.t.u. are used to generate
one pound of steam under the given conditions of steam pressure
and temperature of feedwater. The number of pounds of water
evaporated per hour multiplied by the number of B.t.u. to gen-

iA. S. M. E. POWER TEST CODE, Edition of 1915, Table 4, 21.

BOILERS 39

erate one pound of steam under the conditions will give the num-
ber of B.t.u. used by the boiler per hour. This product divided
by 969.7X34.5 will give us the horsepower of the boiler.

EXAMPLE. It is required to find the horsepower of a boiler working under
the following conditions:

Steam pressure = 115 pounds gage.
Temperature of feedwater = 65 F.
Quality of steam = 98% (i.e., 2% priming).
Water fed to boiler per hour = 3640 pounds.
SOLUTION. From the steam tables it is seen that
The heat of the liquid at 115 Ib. gage (129.7 Ib. abs.) = 318.4 B.t.u.
The heat of vaporization at 115 pounds gage =872.3

The heat of the liquid at 65 F. =33.1 B.t.u. The heat required to evapo-
rate one pound of water under the above conditions is 318. 4 + . 98X872. 3
33.1 = 1140 B.t.u., and the B.t.u. used per hour is 1140X3640 = 4150000.
Hence the horsepower of the boiler is 4150000/(34.5X969.7) = 124 h.p.

38. Factor of Evaporation. Since it takes 969.7 B.t.u. to
evaporate a pound of water from and at 212 F., and since it
takes more (1140 B.t.u. in the previous example) to evaporate
a pound of water under the actual conditions that exist in the
boiler, a certain ratio exists between these amounts. The factor
of evaporation is the ratio of the amount of heat required to evap-
orate a pound of water under actual conditions to the amount re-
quired to evaporate a pound from and at 212 F. In the previous
example, the factor of evaporation was 1140/969.7 = 1.176. If
the factor of evaporation is known, the equivalent evaporation is
found by multiplying the actual evaporation by this factor. 1

By this method, the heat used to raise the temperature of the
moisture in the steam from the temperature of the feed water to
that of the steam is not considered in computing the factor of
evaporation. In most cases the difference in results due to this
omission is very small.

39. Efficiency of Boilers. Usually speaking, the efficiency of
anything is the ratio of what is got out to what is put in ; output
and input being measured in like units. For boilers, the term
efficiency means the ratio of the number of B.t.u. in the steam
generated to the number of B.t.u. available in the coal fired. Boiler
efficiency is usually expressed in percent.

1 In the A. S. M. E. POWER TEST CODE the "Mean B.t.u." and steam tables by MARKS
and DAVIS are used, thereby giving 970.4 B.t.u. instead of 969.7 B.t.u. as used above for heat
required to evaporate a pound of water from and at 212 F. See POWER TEST CODE, Edition
of 1915, pp. 28 and 47.

40 ENGINES AND BOILERS

The fact that the combined efficiency of a boiler, furnace, and
grate is not 100% is due to several losses. These losses are due
to the following causes.

(1) A part of the fuel may drop through the grate and be lost
in the ash.

(2) Heat is lost up the stack. There are several sources of this
loss, and to them is due the greatest loss in efficiency. First,
unburned particles of solid fuel are often carried from the fur-
nace. The amount depends upon the draft and the kind of fuel.
In locomotives, with their high draft and with a fine fuel, this
loss may be considerable. Second, there is loss due to the un-
burned or partially burned hydrocarbons. Black smoke is caused
by the incomplete burning of some of the hydrocarbons. Third,
heat is carried away by the excess air and the inert nitrogen
which have been heated, and by the hot products of combustion.
Fourth, heat is required to evaporate and to superheat the
moisture in the fuel and in the air. Fifth, there may be loss
due to the burning of the carbon to carbon monoxide instead of
to carbon dioxide.

(3) Heat is lost by radiation from the furnace and from the
boiler.

It is very difficult to separate all these losses and the attempt
is seldom made. It must be remembered that what is often
called boiler efficiency is really the combined efficiency of grate,
furnace, and boiler.

Under the most favorable conditions, using coal as a fuel,
efficiencies of over 80% have been attained. Under ordinary
conditions of operation, efficiencies vary from 80% to less than
50%. The efficiency may sometimes exceed 80% when underfeed
stokers, described later, are used. When oil is used as a fuel,
higher efficiencies may be attained, due in part to the better mix-
ing of the air and the fuel.

EXAMPLE. It is required to find the combined efficiency of a boiler, fur-
nace, and grate, working under the following conditions:

Steam pressure = 127 pounds gage.

Superheat = 190 F.

Temperature of feedwater = 180 F.

Water fed to boiler per hour = 8750 pounds.

Coal fired per hour = 1 160 pounds.

B.t.u. per pound of coal as fired = 11540 B.t.u.

BOILERS 41

SOLUTION. The B.t.u. required to generate one pound of steam under
the above conditions is seen to be

325.4+866.8 - 148+.55 X 190 = 1148.7 B.t.u.

The total B.t.u. used in the generation of steam per hour then is 8750 X 1148.7 =
10051000 B.t.u. The total B.t.u. in coal fired per hour is 1160X11540 =
13386000 B.t.u. Hence the efficiency is 10051000/13386000 = .752 or 75.2%.

40. A. S. M. E. Boiler Test Code. 1 In reporting the results
of a steam-boiler test it is well to put them in the form prescribed
by the A. S. M. E. This form is as follows.

DATA AND RESULTS OF EVAPORATIVE TEST
CODE OF 1915

(1) Test of boiler located.

To determine

Test conducted by

DIMENSIONS

(2) Number and kind of boilers

(3) Kind of furnace

(4) Grate surface (width length ) sq. ft.

(a) Approximate width of air openings in grate in.

(6) Percentage of area of air openings to grate surface per cent

(5) Water heating surface sq. ft.

(6) Superheating surface sq. f t.

(7) Total heating surface sq. f t.

(a) Ratio of water heating surface to grate surface

(6) Ratio of total heating surface to grate surface

(d) Volume of combustion space between grate and heating surface

cu. ft.

(e) Distance from center of grate to nearest heating surface ft.

DATE, DURATION, ETC.

(8) Date

(9) Duration hr.

(10) Kind and size of coal

AVERAGE PRESSURES, TEMPERATURES, ETC.

(11) Steam pressure by gage Ib. per sq. in.

(a) Barometric pressure in. of mercury

(12) Temperature of steam, if superheated deg.

(a) Normal temperature of saturated steam deg.

(13) Temperature of feedwater entering boiler deg.

(a) Temperature of feedwater entering economizer deg.

(6) Increase of temperature of water due to economizer deg.

i A. S. M. E. POWER TEST CODE, Edition of 1915, p. 51.

42 ENGINES AND BOILERS

(14) Temperature of escaping gases leaving boiler deg.

(a) Temperature of gases leaving economizer deg.

(6) Decrease of temperature of gases due to economizer deg.

(15) Force of draft between damper and boiler in. of water

(a) Draft in main flue near boiler in. of water

(6) Draft in flue between economizer and chimney in. of water

(d) Draft or blast in ash-pit in. of water

(16) State of weather

(a) Temperature of external air deg.

(6) Temperature of air entering ash-pit deg.

QUALITY OF STEAM

(17) Percentage of moisture in steam or number of degrees of superheating

per cent or deg.

(18) Factor of correction for quality of steam

TOTAL QUANTITIES

(19) Total weight of coal as fired Ib.

C 20) Percentage of moisture in coal as fired per cent

(21) Total weight of dry coal (Item 19X

(22) Ash, clinkers, and refuse (dry)

(a) Withdrawn from furnace and ash-pit Ib.

(6) Withdrawn from tubes, flues and combustion chamber Ib.

(d} Total Ib.

(e) Weight of clinkers contained in total ash Ib.

(23) Total combustible burned (Item 21 Item 22d) Ib.

(24) Percentage of ash and refuse based on dry coal per cent

(25) Total weight of water fed to boiler Ib.

(26) Total water evaporated, corrected for quality of steam (Item 25 X Item

18) Ib.

(27) Factor of evaporation based on temperature of water entering boiler. . . .

(28) Total equivalent evaporation from and at 212 degrees (Item 26Xltem

27) Ib.

HOURLY QUANTITIES AND RATES

(29) Dry coal per hour Ib.

(30) Dry coal per square foot of grate surface per hour Ib.

(31) Water evaporated per hour, corrected for quality of steam Ib.

(32) Equivalent evaporation per hour from and at 212 Ib.

(33) Equivalent evaporation per hour from and at 212 and per square foot

of water heating surface Ib.

CAPACITY

(34) Evaporation per hour from and at 212 (Same as Item 32) Ib.

(a) Boiler horsepower developed (Item 34-5-34^) bl.-h.p.

BOILERS 43

(35) Rated capacity per hour, from and at 212 ........................ lb.

(a) Rated boiler horsepower .............................. bl.-h.p.

(36) Percentage of rated capacity developed ...................... per cent

ECONOMY

(37) Water fed per pound of coal as fired (Item 25-r-Item 19) ........... lb.

(38) Water evaporated per pound of dry coal (Item 26 -^ Item 21) ........ lb.

(39) Equivalent evaporation from and at 212 per pound of coal as fired

(Item 28 4- Item 19) .......................................... lb.

(40) Equivalent evaporation from and at 212 per pound of dry coal (Item

28^- Item 21) ............................................... lb.

(41) Equivalent evaporation from and at 212 per pound of combustible

(Item 28 -f- Item 23) .......................................... lb.

EFFICIENCY

(42) Calorific value of 1 pound of dry coal by calorimeter ............ B.t.u.

(a) Calorific value of 1 pound of dry coal by analysis ......... B.t.u.

(43) Calorific value of 1 pound of combustible by calorimeter ......... B.t.u.

(a) Calorific value of 1 pound of combustible by analysis ...... B.t.u.

(44) Efficiency of boiler, furnace and grate

/ Item40X970.4\

( 100X -TtoZlar-) ................ percent

(45) Efficiency based on combustible

/ in Item4lX970.4\

( 100X - Item 43 J ............... P6F C6nt

COST OF EVAPORATION

(46) Cost of coal per ton of ...... pounds delivered in boiler room .........

................ dollars

(47) Cost of coal required for evaporating 1000 pounds of water under ob-

served conditions ......................................... dollars

(48) Cost of coal required for evaporating 1000 pounds of water from and

at 212 .................................................... dollars

SMOKE DATA

(49) Percentage of smoke as observed ............................ per cent

(a) Weight of soot per hour obtained from smoke meter ..... per cent

FIRING DATA

(50) Kind of firing, whether spreading, alternate or coking ................

(a) Average thickness of fire .................................. in.

(6) Average intervals between firings for each furnace during time
when fires are in normal condition ...................... min.

. . min.

ENGINES AND BOILERS

(51) Analysis of dry gases by volume

(a) Carbon dioxide (CCh) per cent

(b) Oxygen (O) per cent

(d) Hydrogen and hydrocarbons per cent

(e) Nitrogen, by difference (N) per cent

(52) Proximate analysis of coal

As fired Dry coal Combustible

(a) Moisture

(6) Volatile matter

(d) Ash

100 per cent 100 per cent 100 per cent

(e) Sulphur, separately determined referred to dry coal per cent

(53) Ultimate analysis of dry coal

(a) Carbon (C) per cent

(b) Hydrogen (H) per cent

(d) Nitrogen (N) per cent

(e) Sulphur (S) per cent

(/) Ash per cent

100 per cent

(54) Analysis of ash and refuse, etc.

(a) Volatile matter per cent

(6) Carbon per cent

100 per cent
(d} Sulphur, separately determined per cent

(d) Fusing temperature of ash deg.

(55) Heat balance based on dry coal .

Dry Coal

B.t.u.

Percent

(a) Heat absorbed by the boiler (Item 40X970.4) . . .
(6) Loss due to evaporation of moisture in coal

by the burning of hydrogen

(d) Loss due to heat carried away in the dry flue gases

(e) Loss due to carbon monoxide

OLoss due to combustible in ash and refuse
i Loss due to heating moisture in air

(h) Loss due to unconsumed hydrogen and hydrocar-
bons, to radiation and unaccounted for . .

(i) Total calorific value of 1 pound of dry coal
(Item 42)

100

CHAPTER V
BOILER ACCESSORIES AND AUXILIARIES

41. Grates. Grates are used to support the fuel in a furnace.
Most grates are made of cast iron, which is cheap and less liable
than other convenient materials to be distorted or twisted under
the high temperatures to which it is subjected. The grate must
be strong enough to support the load placed upon it, and it must
be of such a form that sections can easily be replaced when broken
or burned out. It must have sufficient opening for the admis-
sion of air to the fuel. The openings or air spaces depend upon
the kind of fuel used. The combined area of the openings will
usually be from 30 to 50 percent of the total area.

The area of the grate depends upon the amount of coal to be
burned and the rate of combustion. Under natural or chimney
draft, from 10 to 25 pounds of coal can be burned per square
foot of grate surface per hour. Under forced draft, from 40 to
130 pounds of coal may be burned per hour. If hand firing is
employed, the grate must not be longer than the distance the
fireman can throw the coal accurately (six or seven feet) . Depend-
ing upon the fuel, draft, and economy of the boiler, the equivalent
evaporation per pound of coal will vary from 5 to 12.

Various forms of grates are used. For hand firing, plain grates
and shaking or dumping grates are used. The plain grate is
harder to keep clean than a dumping grate. Moreover, it is
necessary to keep the fire-doors open while the cleaning is in
process. The grates used in mechanical stokers are of various
types; some are stationary, others traveling and rocking. Occa-
sionally grates are water-cooled, to prevent their burning out.
Since this water is led to the boiler after becoming heated in the
grate, the boiler capacity is increased, but in most cases the extra
care and cost are prohibitive.

42. The Plain Grate. The grate bars shown in Figs. 16 and
16a are of the stationary type. These grates are cast in small
sections so that a section may be easily and quickly replaced
when it is burned out. The size of the openings in the bars is
governed by the size and kind of coal that is to be burned. If

ENGINES AND BOILERS

anthracite coal is used, the openings are small. If the coal is
bituminous, and if it cakes, the openings should be made large.

43. The Rocking Grate. A form of rocking grate is shown in
Fig. 17. The bars are supported on pivots, and are dumped or
rocked by means of a lever from the front of the furnace. Only the
largest clinker need be removed from the top, since the rocking
action of the bars breaks up most of the clinker that is formed.
In case a strong draft is used, as in the locomotive, this type of
grate is usually used in order to keep a clean fire, such as is required
with a high rate of combustion.

44. Mechanical Stokers. The first cost of a mechanical
stoker is greater than the equipment for hand firing, but it

Herat/on \ g

cfioi? at 4-3 '

FIG. 16

FIG. 16a

tr rr

requires less labor and attendance in its operation. A cheaper
grade of fuel can be used, a higher efficiency attained, and less
smoke is formed than is usual with hand firing. In a fair-sized
or large plant, it is usually better economy to use some form of
mechanical stoker. There are many forms of stokers in use.

45. The Chain Grate Stoker. Where a low-grade fuel is
used, as is often the case in the middle west and in the central
states, the chain grate (Fig. 18) is extensively used. The grate is
composed of a large number of short links, forming an endless
chain. This chain runs over front and rear sprockets. Power is
used to drive one of these sprockets, causing the whole chain
to revolve slowly at a speech which is regulated by a suitable
mechanism. The whole grate is mounted on wheels so that it can
be run out in the open for repairing and cleaning.

BOILER ACCESSORIES AND AUXILIARIES

The coal is fed to the front of the grate from a hopper which
extends across the entire width of the grate. At the rear of the
hopper there is a plate lined with firebrick that may be raised
or lowered, thus regulating the depth of fuel-bed. The volatile
matter in the fuel is distilled off as the coal first enters the fur-
nace. These volatile products pass back over the part of the
fire where the fixed carbon is burning, and are given a chance to
burn there. By the time the fuel-bed has reached the rear of
the furnace the combustion should be complete. The ash and
clinker are dropped off to the ash-pit at the rear.

STirry. ca- t "' o. : 'eg' .-ara>o o ^-^^L

'6'/l'6 ^0 A d^ "tffiuLllAJiH h a <iA' ' A'M rc'-'MV^AS^^

FIG. 18

The front part of the grate is overhung by a firebrick arch.
This allows sufficient time and temperature for complete com-
bustion before the gases strike the heating surface of the boiler.

Incomplete combustion is apt to occur with a chain grate if
the fire is forced very hard. Excess air is likely to leak in if the
fuel-bed becomes too thin. This causes a drop in the tempera-
ture of the combustion chamber and therefore poor combustion.

Under a light load, the fuel is often burned before it reaches
the rear of the grate. Air is likely to leak through the ash,
causing poor economy. As a remedy for this condition, a con-
trivance similar to a damper is sometimes placed under the rear
portion of the grate. This makes it possible to shut off the air
supply from this part of the grate.

ENGINES AND BOILERS

46. The Roney Stoker. The inclined grate stoker is one in
which the coal is fed from a hopper at the top, the coal burning

on its way down across the sloping grate. In the forms custom-
arily used, the grates are operated mechanically. There are two
classes of these grates, side-feed and front-feed.

BOILER ACCESSORIES AND AUXILIARIES 49

Figure 19 shows the Roney type of front-feed stoker. The
coal is fed into the hopper, usually by gravity from bins above.
A reciprocating pusher forces the coal from the hopper onto a
dead plate beneath the front of the arch, where the distillation
starts. From this plate, it is made to move downward by the
motion of the grates. By the time the fuel reaches the ash plate
at the bottom of the incline the combustion is complete. The
ash is dumped into the pit below from time to time by means of
a hand-lever that is operated from the front of the furnace. The
grates are rocked by means of an eccentric placed on a rotating
shaft running horizontally along the front of the whole battery
of boilers. The amount the grates are rocked, and the amount
of coal fed, are under the control of the fireman.

Since the distilled products are driven off at the front of the
firebrick arch, they have time, and are at such a temperature,
due to the fire below, that complete combustion takes place. In
some cases, steam and air are admitted under the front of the
arch to aid in the combustion. The length of arch varies with
the grade of fuel to be used, and with the kind of boiler under
which the stoker is installed. In some cases it covers the entire
grate, forming a Dutch oven which sits out in front of the rest
of the boiler setting.

47. The Underfed Furnace. In the types of stokers previ-
ously described, the volatile matter is distilled and burnt over
the bed of burning fixed carbon. As the feeding of fuel is uniform
the amount of gas given off at any one time is not so great as in
hand firing. Hence combustion has a chance at all times to be
more nearly complete. Another and radically different method
is to feed the green coal from beneath, blowing the volatilized
matter along with sufficient air up through the hot fuel-bed above,
where complete combustion takes place.

Figure 20 shows such a stoker. The coal is fed into a hopper
and is forced back under the fuel-bed into retorts, by means of a
ram or plunger. Air under pressure is forced in through tuyeres
at the distillation zone, and by the time the gases pass through
the hot fire of fixed carbon and reach the top of the fire, they
are completely burned. This type requires a forced draft. The
refuse is forced back and down onto the dump plates at the rear
of the wind-box, and is dropped from there into the ash-pit below
through an adjustable opening.

ENGINES AND BOILERS

BOILER ACCESSORIES AND AUXILIARIES 51

Since the temperature of the fire is very high, most of the ash
is fused. In order to prevent clogging it is sometimes necessary
to have a water-cooled bridge wall.

Due to the forced draft, a plant thus equipped is not subject
to the variations due to changing weather conditions. The rate
of combustion is regulated by the amount of coal fed and the
amount of air blown in, which are controlled together. The under-
feed type of stoker therefore is more flexible and will give higher
efficiency under conditions of forced load than those previously
described.

48. Smoke Prevention. Soft coal is generally considered to
be smoky. Nevertheless, it is possible to burn practically all
grades of bituminous coal with very little smoke. Only during
the past few years has there been any very determined effort to
rid ourselves of the smoke nuisance. Several of our larger cities
have ordinances which are enforced more or less rigidly against
the excessive emission of black smoke by power plants.

The soot, which is the solid and black part of the smoke, dis-
figures buildings and may even injure health by keeping out the
sunlight and by clogging up the respiratory organs. It is fre-
quently stated that a great amount of fuel goes to waste in the
soot of the smoke. While the soot may be a nuisance, yet it
represents in itself but little heat loss. Soot indicates, however,
incomplete combustion, which often means that there is a large
amount of unburned hydrocarbon and probably some CO that
should be burned to C02- Smoke prevention, in many cases,
results in an increased economy of the power plant. Thus we
often find instances of plants that have installed modern equip-
ment which prevents the formation of smoke, not so much with
the idea of eliminating the smoke as to obtain better efficiency
by means of proper combustion.

Let us consider in a general way the causes of smoke produc-
tion. Bituminous coal, upon being heated to moderate tempera-
tures (600 to 1000 F.), will have certain hydrocarbons driven
from it in a volatile state. This volatile matter will burn when
mixed with a sufficient amount of air and raised to a sufficiently
high temperature (1800 to 2000 F.), forming carbon dioxide
and water. If for any reason there is an insufficient supply of
oxygen, or if the hydrocarbons are not raised to a sufficiently high
temperature, incomplete combustion will follow, and black smoke

52 ENGINES AND BOILERS

may result. The fixed carbon left in the coal after the volatile
hydrocarbons have been driven off, with the addition of sufficient
air, will burn to carbon dioxide. If there is a lack of air, carbon
monoxide, or a mixture of carbon monoxide and carbon dioxide,
will result.

In conclusion, to burn bituminous coal smokelessly, a furnace
must have a sufficient supply of air to insure the complete com-
bustion of the volatile matter, and it must have a temperature
high enough to permit of this combustion. In the ordinary fur-
nace the time taken for the gases to pass from the grate to the
comparatively cool heating surface of the boiler, where they are
rapidly cooled, is quite small, perhaps less than a second. At
least, not enough time is allowed for the volatile matter to unite
properly with the oxygen of the air, and black smoke is the result.
If the path of the hot gases can be made longer, thus giving them
time to burn, a large reduction can be made in the amount of smoke.
It is sometimes possible to rearrange the baffling in the path of
the gases in such a way that this is accomplished easily. Another
way to lengthen the time of burning is to move the grates out
from under the boiler and place a long firebrick arch over the fire.

In ordinary hand firing, a large quantity of cold coal is thrown
on top of the fire, with the result that the fire is greatly cooled,
both to heat up the coal and also to cause distillation of the
volatile matter. This reduces the temperature to a point at
which the complete combustion of the volatile matter will not
take place, and smoke results. At the time the volatile matter
is given off an excess of air is needed to burn it. If this air is
let in over the top of the fire it will still further cool the fire,
which only adds to the trouble.

The use of the various over-feed stokers, such as the chain
grate, and those with the front or side feed, is a decided improve-
ment over hand firing because the fresh coal is added continu-
ously, and the air supply can be properly adjusted and main-
tained. In these types of stokers, the fresh coal, upon coming
to the furnace, passes through the distillation period in such a
position that the volatile products must pass over the fire of fixed
carbon and be burnt there.

In the underfeed type of stoker, the fresh coal is forced in from
the under side of the fire, and the distilled products along with
sufficient air to burn them, are forced to pass through the hot

BOILER ACCESSORIES AND AUXILIARIES 53

bed of burning carbon above, with the result that a sufficiently
high temperature is maintained to allow for their complete com-
bustion.

The purpose of the down-draft furnace is much the same as
that of the underfeed stoker. In this type, the zones of distill-
ation and of burning the fixed carbon are separated. The first
takes place on the upper grate and the second on the lower; the
distilled product passes over the hot fuel bed on the lower grate
and complete combustion occurs.

Many devices to prevent smoke are on the market. Most of
them consist of some form of steam jet that carries in and mixes
a sufficient amount of air with the volatilized hydrocarbons to
effect complete combustion. The steam itself has no power to
prevent smoke. It is used simply to carry the air and to mix
it with the volatile products. If the steam jet is left on too long
after the period of distillation, a loss greater than the gain
effected by the jets may result. Some makers use a dash-pot or
other arrangement that automatically shuts the steam off soon
after each firing.

49. Settings. The brickwork that surrounds a boiler is called
a setting. The outer side of this setting is built of common red
brick. The inner surfaces that are exposed to the high tempera-
ture of the flame are lined with firebrick. With the very high
temperatures that exist in modern furnaces, it is difficult to get
a grade of firebrick that will give satisfactory service. It is better
practice not to leave an air space between the outer wall and the
lining, since heat is transmitted through the air space, under the
high temperatures that exist in a furnace, faster than it would
be through the same thickness of brick.

In most furnaces, a firebrick arch is placed over the fire. This
arch forms a chamber in which the temperature is kept very high.
The length of this arch varies with the kind of fuel to be used
and with the type of boiler. Firebrick baffles are placed between
the tubes of water tube boilers in such a manner that the gases
are forced to pass the heating surface several times on their way
to the stack. The burnt gases pass from the setting to the stack
through a duct called the breeching.

Since there is a difference in pressure between the outside and
inside of the setting, it is important that there be no cracks for
the air to leak in or for the gases to leak out. Under natural

54 ENGINES AND BOILERS

draft, there is a leakage of air inward, which cools the boiler
and injures the draft. With a forced draft, the gases may leak
out into the boiler room.

50. Draft. Natural draft is obtained by means of a stack or
chimney. The gases as they leave the boiler are at a tempera-
ture of from 400 to 600 F. At this temperature they are much
lighter than the outside air. Since the column of gas in the stack
is lighter, it is forced up from the bottom by the heavier air
outside. The stack should be large enough so that but little
draft will be lost by the friction between the gas and the stack.
It should be insulated so that little heat is lost through the walls
of the stack.

There are three kinds of stacks in use, steel, brick, and con-
crete. The steel stack is cheaper and lighter, but it is expensive
in its upkeep, since it must be painted often to prevent the
corroding of the plates. In the better grades of steel stacks, a
firebrick lining is used at least a part of the way up to prevent
conduction of heat to the outside. Brick stacks are sometimes
built of hard common brick, but of late years more often of
special radial brick. They are sometimes lined with firebrick,
as in the steel stack. Reinforced concrete stacks have come into
use during the past few years. When properly put up, they give
good service.

Brick and concrete stacks must be heavy enough to resist the
overturning effort of the wind. Steel stacks are either anchored
and designed to withstand the bending action due to the wind,
or else they are supported by guys.

Where forced draft is used, the stack need be only high enough
to discharge its smoke above the surrounding buildings. Forced
draft is obtained by means of fans or blowers which force the
air into the ash-pit or wind-box and thence through the fire.
In locomotives, the forced draft is obtained by means of nozzles
through which the exhaust steam from the engines is discharged
into the chimney. A draft caused by this method is sometimes
called induced draft.

The amount of draft is measured in inches of water. Under
natural draft it will increase in going from the ash-pit to the stack
and at the base of the stack it will be from 0.5 to 1.5 inches.
Under forced draft the pressure in the ash-pit is greatest, and will
vary from 1 to 5 inches.

BOILER ACCESSORIES AND AUXILIARIES 55

51. Dampers. A damper should be interposed between each
furnace and the stack. The efficient operation of the furnace
necessitates careful attention to the damper. Automatic damper
regulators are in use, but for the best results they should be sup-
plemented by intelligent manual control.

52. Safety Devices. There are in general three causes of
explosions of properly designed boilers: a weakened part, high
pressure, and low water. The first is due to the corroding or
wearing away of some part of the structure, or to a local over-
heating due to an accumulation of sediment or scale. It may
be due to an undetected flaw in the materials entering into the
makeup of the boiler, or it may be due to carelessness or poor
workmanship during construction.

The second cause, high pressure, is due to a pressure much in
excess of that for which the boiler was designed. This may be
due to a faulty safety valve, or to the ignorance of a fireman.

The third cause, low water, allows some of the parts to get
overheated and therefore much weakened. It may exist unknown
to the fireman on account of foaming, which is liable to cause an
untrue indication of the water level in the gage glass, or on ac-
count of some stoppage in the connection to the glass.

To safeguard against accidents due to a weak part, it is neces-
sary to have a thorough inspection both of materials entering into
construction of the boiler and of the workmanship during con-
struction. There also should be frequent inspection after the
boiler is put into service. The common test for strength is hydro-
static. Before being put into service a boiler should have water
pumped into it, and a pressure should be reached much in excess
of the working pressure.

53. The Pressure Gage. The pressure that exists in a boiler
is measured by a steam gage. In this country, the dial of the
steam gage is graduated to read in pounds per square inch. The
gage almost universally used is known as the Bourdon gage.
Figure 21 shows its internal construction. The pressure is ad-
mitted to a curved flattened tube which is closed at its free end.
This internal pressure tends to make the curved tube straighten
out. The free end is connected to the needle by means of levers,
a rack, and a pinion. Any movement of the free end causes the
hand or needle to turn, and the pressure causing the movement
is indicated on the properly graduated dial. The flattened tube

ENGINES AND BOILERS

is made of brass or steel. Since a change in the temperature of
the tube would cause an error in the reading of the gage, it should
be connected to the boiler or steam-pipe by means of a siphon
so that steam may never enter the gage. On locomotives, where
the gage is subject to continual and severe jarring, two stiff er
tubes are used in place of the one shown in the figure. The better
grades of gages have a light hairspring to take up the backlash
in the levers and gears.

A gage similar to the one shown in Fig. 21 is often used to
indicate vacuum. For a vacuum, the tube is bent still more and

the levers are so arranged that
the motion of the needle is re-
versed from that of the one shown
in Fig. 21. The dials of vacuum
gages are commonly graduated to
read in inches of mercury. Where
pressures are had that may fluctu-
ate from a vacuum to a positive
pressure, gages are used that will
indicate either the vacuum or the
pressure above the atmosphere.
Gages should be tested from time
to time to see that they give the
correct pressure reading.
54. The Safety Valve. The purpose of a safety valve on a
boiler is to prevent an undue or dangerous pressure. It is im-
possible in an ordinary furnace to regulate the combustion quickly
enough to correspond to sudden changes of the amount of steam
used. For instance, a boiler may be furnishing its maximum
amount of steam, when for some reason the engine is shut down
without warning, and therefore before the fire can be deadened.
As a result, the rapid formation of steam will continue long
enough to cause an excessive boiler pressure if the safety valve
does not give relief. Hence the safety valve should have such a
capacity that it is capable of discharging all the steam that the
boiler can generate without allowing the pressure to become
dangerous. Furthermore the safety valve must be absolutely
reliable in its action, and it should be so constructed and placed
on the boiler that it cannot be put out of action through careless-
ness or ignorance. No stop valve should be placed between it

FIG. 21

BOILER ACCESSORIES AND AUXILIARIES

and the boiler. Many explosions have been caused by the failure
of the safety valve to operate. As far as the writer knows, how-
ever, none have occurred that were due entirely to excessive
pressure when the valve was in action. Several types of safety
valve have been used in the past, but with the pressure ordi-
narily carried in this country, the use of the pop type has become
almost universal, and will be the only one described here.

Figure 22 shows in section a pop safety valve. Most safety valves
are made with a 45 seat. The valve is held on the seat by a
helical spring. When the steam pressure becomes sufficient to
overcome the force of the spring, the valve is raised enough to al-
low some steam to escape. This steam passes into a huddling
chamber. The area upon which the steam now acts is slightly
greater than before the valve opened, with the result that the spring
is compressed suddenly to a greater extent than if the steam acted
only upon the original area. Escape of steam will continue until
the pressure has dropped to a few pounds less than that at which
it opened. When the valve stops blowing, it seats firmly. Since
the pressure is less than that at which it opened, it will remain
shut until the steam pressure
again reaches the popping point.
The valve should be constructed
and set so that the difference be-
tween the popping pressure and
the closing pressure or Slowdown
is not too large, in order to pre-
vent shock to the boiler and an
excessive loss of steam during
ordinary operation.

A lever is provided at the top
of the casing for locking the
valve open. The compression in
the spring may be adjusted by
screwing the top cap up or down.
In most valves the spring is en-
cased so as to protect it from
the escaping steam, and to pre-
vent back pressure in the dis-
charge pipe from acting on the
top of the valve. (See Fig. 22.) FIG. 22

58 ENGINES AND BOILERS

55. Safety-valve Capacity. The amount of steam that a
safety valve discharges depends upon the steam pressure and upon
the effective opening. The latter varies with the lift of the valve.
Not all valves of the same diameter have the same lift; hence
they differ in capacity. There has been considerable agitation
recently to have all pop valves put upon a uniform rating. Tests
have been made to determine the discharge and the lift under
various conditions of pressure. These tests show that the com-
monly used empirical formula given by Napier is substantially
correct when applied to the safety valve.

56. Napier's Formula. Napier's formula for steam issuing
from an orifice into the atmosphere is

W- A ' P

in which W is the weight in pounds of steam issuing per second,
A is the area of the orifice in square inches, and P is the abso-
lute pressure in pounds per square inch in front of the orifice.

Applying this formula to the safety valve with a 45 seat, it
is seen that the area oi opening, A, equals approximately the
product of irD and the lift times the sine of 45, where D is the
diameter of the valve in inches. If the lift is known, the dis-
charge may be calculated. Some valve makers use the assump-
tion that the lift is 1/30 of the diameter, and rate their valves
accordingly. Under this assumption it is seen that

7rD 2 X.707xP

W =

70X30

and the weight of steam discharged per hour is 3600 W = 3.81 PD 2 .
As previously explained, the safety valve must be large enough
to discharge the maximum amount of steam the boiler is capable
of generating. We can compute this maximum amount from the
heating surface of the boiler, allowing an evaporation of from six
to ten pounds of water per square foot of heating surface per
hour, or we may compute it from the grate area, assuming a
boiler efficiency and a rate of combustion consistent with the
draft and with the kind of fuel used. The former method is con-
sidered better. After conducting numerous tests P. G. DARLING 1

i TRANS. A. S. M. E., vol. 31 (1909), p. 109.

BOILER ACCESSORIES AND AUXILIARIES 59

advocated to the A. S. M. E. formulas for pop safety valves
derived according to the following method:

for stationary boilers, D= 0.068 ^j

-L/Z

for locomotive boilers, D= 0.055 f^>

Lir

in which D is the diameter of the valve in inches, H is the heat-
ing surface of the boiler in square feet, L is the lift of the valve
in inches, and P is the absolute boiler pressure in pounds per
square inch. It is noticed that smaller valves are required for
locomotives, because the maximum draft can be secured only
when the steam is being drawn from the boiler by the engine.

57. Other Safety-valve Formulas. Various cities and states
have their own rules governing the sizes of safety valves, a few
of which are given below.

CITY OF CHICAGO. One square inch of pop-valve area (7rD 2 /4)
for every three square feet of boiler grate area.

CITY OF PHILADELPHIA. For pop valves, A = 22.5XCr/(p 8.62),
in which A is the area of the valve in square inches (not the effec-
tive opening for the escape of steam), G, the grate area in square
feet, and p the gage boiler pressure.

U. S. SUPERVISING INSPECTORS. A = .2Q74:XWH/P, in which A
is the area of the valve as in the previous formula, WH is the
number of pounds of water evaporated by the boiler per hour,
and P is the absolute boiler pressure.

Safety valves are not made in sizes over 5 or 6 inches in diameter.
In large boilers it is therefore necessary to use more than one.

EXAMPLE. What should be the size of pop safety valve on a boiler with
1500 square feet heating surface if the pressure carried is 130 pounds gage?

SOLUTION. Assuming a maximum rate of evaporation of 8 pounds of
water per square foot of heating surface per hour, we get 8X1500 = 12000
pounds of steam to be discharged per hour through the valve. The weight
discharged per second is 12000/3600 = 3.33 pounds. From Napier's formula,
W = AP/70, we see that 3.33 =AX(130+15)/70, whence A, the area of
opening, in square inches, is 1.61 Assuming a lift of valve equal to 1/30 of
the diameter, the area of the opening will be approximately .7077rD 2 /30.
Then .7077rZ) 2 /30 = 1.61, or D = 4.67 inches. Hence a 5-inch valve should be
used.

58. The Water Glass or Gage Glass. In order that the
amount of water in the boiler may be known, a water glass is
attached. The lower end of the water glass is attached to the

ENGINES AND BOILERS

water space, and the upper to the steam space. Since there is
danger of the glass becoming stopped in these connections, and
the water level thereby being falsely indicated, or of the glass

being broken, three gage-cocks
or try-cocks are placed on the
boiler or water column. The
top cock is placed above, and
the bottom one below, the
normal water level. By open-
ing these cocks in succession,
one may determine whether
or not the gage glass is giving
the correct level.

59. High and Low Water
Alarm. High- water and low-
water alarms are sometimes
used to attract the attention
of the fireman when the water
falls below or rises above the
safe level. The alarm is
operated by means of a float
in the water column. When
this float rises too high or falls
too low it will open a valve,
and allow the escaping steam to blow a whistle. (See Fig. 23.)

60. Fusible Plug. Another safety device used to detect low
water is the fusible plug. This plug (Fig. 24) has a tin core
that will melt when the water
level falls below it. These
plugs are placed in the crown
sheet of an internally fired
boiler in the rear head a little
above the top tubes in the
return-tubular boiler and in the bottom of the steam drum of a
water-tube boiler. They should be kept free from scale on the
inside and from soot on the outside. None of the above safety
devices are absolutely certain in their action. Under conditions
of very rapid steaming or with feedwater that foams, the water
level in the boiler may be below that in the water column.

FIG. 23

//7S/t/f or arfsjure
iat/Se ry/*e

Outft'efe ft/pes
0ttts/e/e or //

FIG. 24

BOILER ACCESSORIES AND AUXILIARIES 61

61. Boiler Feedwater Treatment. The impurities in water
that are responsible for most of the scale formation are the car-
bonates and sulphates of calcium and magnesium. If muddy
water is used, the mud may be deposited on the heating surface
and aid in scale formation.

The carbonates of lime and magnesium are soluble in water
containing carbon dioxide. These carbonates cause what is known
as temporary hardness. . Upon heating to 212 F. the carbon
dioxide is driven off, and the carbonates are precipitated. If
these are the only impurities in the feedwater, an open feedwater
heater will remove most of the scale-forming material. Where a
heater is not used the carbonates may be precipitated by the
addition of a solution of slacked lime. The lime combines with
the carbon dioxide to form the insoluble monocarbonate of lime.

The sulphates of lime and magnesium are not precipitated at a
temperature of 212, but are precipitated at a temperature such
as exists in the boiler. They cause what is known as permanent
hardness. The addition of carbonate or hydrate of soda (or a
mixture of the two) will cause precipitation. The carbonate of
soda decomposes the sulphates and forms insoluble carbonates of
lime and magnesium, which precipitate, leaving neutral soda and
sodium sulphate in solution. If carbon dioxide is present, the
soluble bicarbonate of lime is formed, which may be precipitated
by heating or by the addition of lime as explained previously. In
most purification processes both the lime and soda are used.

If organic matter, from sewage or from some other source, is
present in the water, it may be removed by filtration. Before
passing the filter a coagulant, such as alum, is often used. Organic
matter in feedwater is often the cause of foaming.

62. Scale Prevention and Removal. Many substances have
been used to prevent the formation of scale. Some of these
probably do as much damage to the boiler as would the scale.
Aside from the treatment to remove the scale-forming material,
the best substance seems to be graphite. When it is injected
into the boiler, it is said to help in the prevention of scale formation.

Where no precaution is taken to prevent the scale from form-
ing, it is necessary to clean it from the tubes periodically. This
is usually done by means of a cutter or hammer that is driven
by a small air, steam, or water turbine.

ENGINES AND BOILERS

Figure 25 shows one make of cleaner that is applied to a fire
tube. Figure 26 shows the form that is applied to a water tube.

FIG. 25

Jca/a

tfrr/zeafaf
jecf/o/J

t/rfer fi/bc

FIG. 26

63. Oil Separators. In plants where all or part of the steam
is condensed and used again as boiler feed, the oil that was used
to lubricate the engine will find its way to the boiler. This does
not apply to steam turbines, as oil is not usually used internally
with them. This oil forms a very hard scale that it is almost im-

possible to remove. To pre-
vent this, oil separators are
used to remove the oil, either
from the exhaust steam or
from the water after it is con-
densed. The removal before
condensation is preferable,
since the oil does not have to
be contended with in the con-
denser or feedwater heater.
Figure 27 shows an Austin oil
separator. Since the separator
is quite large, the steam passes
through it with a small veloc-
ity and deposits the oil on the
surface of the corrugated ver-
tical baffle plate shown in plan
and section in the figure.
With high vacua, a spray of
water keeps the surface moist,
which aids in the separation
FIG. 27 of the oil.

BOILER ACCESSORIES AND AUXILIARIES

64. Boiler Feed-pumps. The feedwater is usually forced into
a boiler by means of a pump. Figure 28 shows a common type
of boiler feed-pump. This pump is direct acting, the steam piston
and water piston or plunger being fastened to the same piston rod.
Since the steam and the water in a boiler are both under the
same pressure and since pipes and fittings offer a resistance to
the flow of both water and steam, it is seen that it is necessary to

FIG. 28

make the steam piston of the feed-pump considerably larger in
diameter than the water plunger.

Steam is admitted by the steam valve alternately to the two
ends of the steam cylinder. At the same time that the valve
is admitting steam to one end of the cylinder it is allowing it to
exhaust from the other, thus giving the piston a reciprocating
motion. The water piston sucks up water from the suction pipe
in one end of the water cylinder while forcing it into the delivery
pipe on the other.

The suction pipe leads either to the hot well or to the cold well
from which the feedwater is taken. The lower end of the suction
pipe should be provided with a strainer to prevent any large
pieces of solid matter from getting into and clogging the valves.
A foot valve should be provided also to keep the suction pipe

64 ENGINES AND BOILERS

full of water when the pump is not running, thus eliminating the
priming of the pump every time it is started.

There are two sets of valves at each end of the water cylinder.
On the suction stroke, the suction valves are lifted and the water
is sucked in back of the piston. During the forcing stroke, these
valves are closed and the water is forced out through the upper
set and into the delivery pipe. A light spring is employed to
help seat the valve. Most valves are faced with a composition
disc which may be replaced when it becomes worn.

The flow from a reciprocating pump is not steady; to insure
a more uniform rate of flow of the water an air chamber is placed
on the delivery line close to the pump. This is kept partly filled
with air, which acts as a cushion. A check valve is placed in the
feedline between the pump and the boiler. This prevents the
water from the boiler escaping back through leaky valves when
the pump is not in full operation. There should also be a stop
valve in the feedline.

There are two types of reciprocating steam pump, one in which
there is only a single steam cylinder and a single water cylinder,
and the other in which there are two steam cylinders and two
water cylinders placed side by side. The latter type is called a
duplex pump. In this type the valve for one steam cylinder is
operated by the movement of the piston of the other steam
cylinder.

These boiler feed-pumps take steam the full length of the stroke,
not allowing it to expand in the cylinder, and they are not eco-
nomical in the use of steam. (See Chapter VI on the steam en-
gine.) However, only a small proportion of the steam generated
by the boiler is needed to run the pump. To secure better
economy, feed-pumps are occasionally driven by power taken
from the main engine. The supply of water from these pumps
is not easily regulated. They are made to pump more water
than is normally required, the excess being passed back to the
suction through a relief valve. In large electric power plants,
triplex pumps, driven by electric motors are often used for boiler
feeding. Of late years centrifugal and turbine pumps have been
employed for boiler feeding.

An automatic regulator is sometimes installed with the pump
so that the pump will furnish just the proper amount of water
to keep the boiler water-level constant.

BOILER ACCESSORIES AND AUXILIARIES

65. The Injector. On portable boilers and in small plants,
the water is often forced into the boiler by means of an injector
or inspirator. This is usual also on locomotives. The principle
upon which the injector works is illustrated by Fig. 29. Steam
from the boiler enters the injector through a steam nozzle, a, in
which it expands and some of its heat energy is transformed into
kinetic energy. The steam leaves the nozzle with a high velocity
and enters a small combining tube, b. The water inlet leads to a
chamber which is located between the nozzle and the combining
tube. As the steam flows from the nozzle to the combining tube
it tends to form a par-
tial vacuum in the
water chamber and
thus sucks up and car-
ries the water along
with it. The steam
mixes with the water in
the combining tube and
is condensed. This mix-

ture of condensed steam
and water has a high
velocity and therefore
a considerable amount
of kinetic energy; its
pressure, however, is "
atmospheric or less.
This mixture passes
from the combining
tube to the delivery FIG. 29

tube, c, which has an

increasing diameter. The mixture therefore loses a large amount
of its velocity and kinetic energy. What it loses in kinetic energy
it gains in pressure energy, so that by the time it leaves the in-
jector it has gained enough pressure to force open the check
valve leading to the boiler. An overflow is located at the end
of the combining tube, so that when steam is first turned on,
it escapes through the overflow. The overflow is fitted with a
valve which automatically closes when the pressure inside the
combining tube falls below the pressure of the atmosphere, thus pre-
venting air from coming into the injector and impeding its action.

66 ENGINES AND BOILERS

Unless specially constructed, an injector cannot lift water a
very great height. Moreover, since the injector must condense
the steam in order to work at all, it is necessary that the water
be cold. Considered as a pump, the efficiency of the injector is
very low, because the greater part of the energy of the steam
goes to heat the water. If it is used to feed a boiler, the heat
spent in raising the temperature of the feedwater is not lost, as
it goes back into the boiler. Hence it is efficient for this purpose.
The injector is light, occupies but little space, and is cheaper
than a pump, but it is not so dependable.

66. Boiler Feeding by Return Trap. The condensation from
various parts of the plant is sometimes returned to the boiler
by what is known as a return trap. This trap is located above
the level of the boiler and the water runs into the boiler under
the influence of gravity and the pressure of live steam. These
traps are quite economical in the use of steam and they may be
used to supply all the feedwater. They are not nearly so reliable
as the pump or injector, however, and are therefore but little
used to furnish the entire feedwater supply.

67. The Steam Line. Steam pipe is made of wrought iron or
of steel. The nominal diameter corresponds approximately with
the inside diameter. Sizes of standard pipe vary, by the y%"
from y 8 " to y 2 ", by the M" from %* to IJ^", by the %' from
IY 2 " to 5", and by the 1" from 5" to 15".

It has been customary to allow an average velocity of steam
in the line of from 4000 to 6000 feet per minute. In modern
turbine plants, however, where the flow is uniform, and especially
where superheated steam is used, velocities much in excess of
these values are used. If a velocity of wet steam much greater
than that just mentioned is used, the drop in pressure due to skin
friction will be excessive. On the other hand, if a velocity much
less is allowed, too large and expensive a pipe will be required.

If the volume and velocity of steam to be carried by the pipe
line are known, the diameter is easily determined. The volume
carried per unit of time equals the product of the area of the
cross-section of the pipe and the velocity. If the size and speed
of the engine to be supplied are known, we may compute the vol-
ume of steam needed. At maximum it may be assumed that the
engine takes steam during the full length of the stroke. When

BOILER ACCESSORIES AND AUXILIARIES

more than one boiler is used it is customary to discharge the
steam into a common pipe called a header. In such a case
each boiler should be provided with a non-return stop-valve be-
tween the boiler and the header. This non-return stop-valve
acts as a check valve in case the direction of steam flow should
be reversed, which would happen in case a tube blew out or some
other similar accident occurred.

The piping must be provided with a sufficient number of
hangers to prevent breaking due to its own weight. The line
should slope downward in the direction the steam is to flow in
order that the condensation may be carried along with the steam.
If this precaution is not followed condensed steam will collect
in the pipe and may be carried in slugs by the steam in amounts
large enough to injure and cause leakage or even breakage of
the fittings. Provision should be made at the low points of the
line to remove condensation. A pipe should be run down from
the low point and the

Va/re seaf
fa/re

water collecting in this
may be blown out from
time to time by open-
ing a valve by hand.
A trap may be installed
that will remove it
automatically.

68. The Steam Trap.

- Several types of
traps are in use. In the
more common kinds,
the valve is operated

by means of either a

* T
float, the unequal ex-

pansion of two differ-
ent metals with chang-
ing temperature, press-
ure of collected water
on a flexible diaphragm,
or the weight of a bucket as it fills with water. The latter kind
is illustrated in Fig. 30. In this type the buoyancy of the
bucket keeps the valve closed until enough water flows over the

FIG. 30

ENGINES AND BOILERS

edge and collects in the bucket to sink it. The sinking of the
bucket opens the valve and the water collected in the bucket is
forced out through the valve by the steam pressure inside the trap.
The bucket now being lightened, it again rises, closing the valve.

In many traps, the valve is operated by a float. The water
collects in a float chamber and raises the buoyant float until the
valve is opened. The water then escapes until the float is lowered
enough to allow the valve to seat. An air valve is located at the
top of the trap to allow the air to escape if enough should be caught
there to interfere with the operation of the trap.

Another form of trap is one in which the valve is operated by
the unequal expansion of two metals. When the trap is cold the
valve is open and the water is allowed to escape. As soon as
the steam flows through, however, the parts are heated and ex-

FIG. 31

FIG. 32

pand unequally, closing the valve. Water then collects again,
and as the parts cool, the valve will again open and the opera-
tion will be repeated. When large amounts of water are to be
handled, dumping traps may be used. The discharged water from
the trap is led to the drain or is piped back to the hot well.

69. Expansion Joints. Since the pipe is laid cold, it will ex-
pand when steam is turned into it and its temperature becomes
that of the steam. The expansion amounts to 2.5 inches per
hundred feet of pipe with ordinary steam temperatures, and may
be greater when the steam is of very high pressure and is super-
heated. The piping must be so arranged that this expansion may
take place without injury to the pipe. If the pipe is not laid
straight but contains elbows, it may bend enough so that no dan-
gerous stresses will be induced. If there is a considerable run of
straight pipe, however, expansion joints must be provided.

There are several types of expansion joints in use. A very com-
mon kind for use with low-pressure steam is the slip- joint. In

BOILER ACCESSORIES AND AUXILIARIES

this, provision is made for the slippage of one part of the joint
on the other. The joint is kept steam tight by means of a stuf-
fing box. Figure 31 shows this type. Goosenecks and expan-
sion loops (Fig. 32) are used when the steam pressure is high.

70. Steam Separators. Unless superheat is used, steam leav-
ing the boiler will always contain some moisture. If the steam-
pipe is very long, some condensation also takes place. Due to
these causes, the steam is liable to reach the engine quite wet.
It is desirable both for safety and for economy to have the steam
as dry as possible when it enters the engine. To remove the
moisture from the steam, a separator is placed in the line just
before it reaches the engine. The steam is

given a sudden change in direction upon enter-
ing the separator. The moisture resists this
change to a greater extent than does the steam.
In the type shown in Fig. 33 the steam is first
deflected downward and then upward, and as
the moisture cannot change its direction of
motion as rapidly as the steam, it is caught
and collected in the bottom of the separator.

In some makes the steam is given a whirl-
ing motion and the water, being denser than
the steam, is forced to the outside of the
separator, where it is collected. Another type,
which is similar to the oil separator of Fig. 27,
is that in which a corrugated baffle plate is
interposed in the path of the steam. The steam passes around
the baffle while the moisture is caught by it and runs down the
corrugations to the bottom of the separator, where it is collected.

A separator should remove most of the moisture, but it should
not offer too great a resistance to the passage of the steam, since
this would cause a drop in pressure. The moisture, after being
collected, is trapped off and discharged to the drain or returned
to the hot well. Often the separator is made large and acts as a
steam receiver. This reduces the pulsation in the steam line
when the steam is used by a reciprocating engine.

71. Steam-pipe Covering. To prevent radiation of heat from
the steam-pipe and the consequent condensation, a covering is
applied to the pipe. The covering is made from materials that

FIG. 33

ENGINES AND BOILERS

are poor conductors of heat. A finely-divided, dead air space is
one of the best non-conductors of heat. In most coverings the
object is to get as much finely-divided dead air space as possible.

7 r <c y-

FIG. 34

The most common kinds of covering are made from asbestos
or a mixture of asbestos and carbonate of magnesia. The mag-
nesia used in pipe covering contains a great number of very small
air cells, and therefore makes an excellent insulator. When the
magnesia is used it is usually moulded into hollow cylinders to-
gether with enough asbestos fiber to give it strength. Another
form of covering much used is made of several thicknesses of
corrugated asbestos paper formed into hollow cylinders by wind-
ing on a mandrel.

BOILER ACCESSORIES AND AUXILIARIES

72. Feedwater Heaters. In most plants, all or some of the
steam is exhausted at atmospheric pressure. If this steam is
exhausted to the air all of its heat is wasted. Some of this heat
may be used to heat the boiler feedwater by running the exhaust
steam through a feedwater heater and extracting its heat of
vaporization. There are two types of

heater, the open and the closed.

I^the open feed^watexjieater the
steam comes m direct contact with
the feedwater, which is made to flow
over shallow pans, thus exposing a
large area to the steam. The tem-
perature of the water is thereby
brought near to the boiling point. If
the water is hard, a large part of the
scale-forming materials will be de-
posited on the pans, which may be
easily removed and cleaned. Figure
34 shows a common form of open
heater. A skimmer is provided to
remove the oil that comes in with the
exhaust steam, and there is also a filter
to purify the feedwater. The pur-
pose of the open heater is thus seen to
be twofold: to utilize the heat that
would otherwise be wasted and to
purify the water. In the closed type

(Fig. 35), the steam and the feedwater do not come into direct
contact. The steam is led through tubes around which the feed-
water is forced to flow. If the water is very hard, the tubes are
liable to collect scale, which hinders the operation of the heater.

73. Economizers. In the ordinary steam plant, the flue gases
pass up the stack at a temperature of about 500 F. This tem-
perature usually will be higher than that of the steam and water
in the boiler, since the latter get their heat from the gases. More-
over, the higher the steam pressure and its temperature, the
hotter will be the flue gas. The tendency during the past few
years has been to use higher pressures, which means a greater loss
of heat up the stack than with low pressures.

FIG. 35

72 ENGINES AND BOILERS

In order to utilize a part of this heat that otherwise would be
wasted, economizers are sometimes installed between the boiler
and the stack. The economizer is simply an added heating surface
in the form of water tubes about which the products of combus-
tion pass on their way to the stack. At best the feedwater will
be at a temperature of only 212 as it enters the boiler, if it has
been heated with exhaust steam at atmospheric pressure. Con-
siderable heat can be added before it reaches the boiling point
when under high pressure. This heat is added in the econo-
mizer. The boiler feedwater is first pumped to the economizer,
where it is heated to near the boiling point corresponding to
boiler pressure, and it then passes on to the boiler.

In a common type of economizer, the heating surface is com-
posed of vertical tubes through which the water flows and around
which the hot gases pass. These tubes are kept clean from soot
by scrapers that are continually moved up and down the
tubes, by means of a small engine or electric motor. As the
economizer depends for its action upon the extraction of heat
from the burnt gases, it follows that the gases will be much cooled,
and if natural draft be employed, they may be cooled enough to re-
duce the draft to such an extent that the efficiency of the whole
plant may be lowered. If forced draft is used this objection does
not hold to so great an extent. In any case, the economizer offers
some resistance to the gases, with a consequent lowering of the
draft. Whether or not an economizer will effect enough of a
saving to pay for itself must be determined in each individual
case. Economizers are often sold under a guarantee to add a
certain percentage to the efficiency of the whole plant.

74. Condensers. After steam has passed through an engine or
turbine, it is often led to a condenser, in which a pressure con-
siderably below that of the atmosphere is maintained. The pro-
cess has several advantages which will be studied in more detail
later. In general, the decreased back-pressure adds to the effi-
ciency and to the capacity of the engine or turbine to an extent
more than sufficient to pay for the additional cost of the condenser,
provided plenty of cool water is available for cooling the con-
denser so as to condense the steam. Moreover, with a surface
condenser, the condensed steam is led back to the boiler and is
thus kept free from scale-forming materials. This last factor is of

BOILER ACCESSORIES AND AUXILIARIES

great importance where the available feedwater is poor. Boilers
can be operated far beyond their rated capacity if they are kept
free from scale. Hence the saving in boilers and in their upkeep
may also go a long way toward paying for the condenser.

There are two types of condensers, one in which the cooling or
circulating water is kept separate from the steam, as in the closed
feedwater heater, and one in which the water (called injection
water) is mixed with the steam, as in the open feedwater heater.
The former is called a surface condenser, and the latter a jet
condenser.

75. The Surface Condenser. Figure 36 shows the construc-
tion of a surface condenser. The exhaust steam enters at the

5urface Condenser?

FIG. 36

top of the shell, passes around the tubes, and, after being
condensed, is pumped out from the bottom of the shell. The
tubes are usually made of thin brass or of special metal, and extend
between two tube sheets. At the outer side of these tube
sheets and within the tubes is the water space. The circulating
water is pumped in at the opening shown in Fig. 36, flows through
the lower half of the tubes to the other end of the condenser,
and then flows back through the top tubes and out. This arrange-
ment is called a two-pass condenser. In some smaller types, the
water enters at one end and flows out at the other; these are
called one-pass condensers.

Occasionally three passes are made, but this type is not gen-
eral. In the design of a condenser, care must be taken that the
steam, upon entering, is directed over the entire surface of the
tubes and that no air pockets may be formed. The condensed
steam should leave that part of the condenser where the circu-
lating water is coldest. Packed joints are used between the
tubes and the sheet.

ENGINES AND BOILERS

Since water will absorb and dissolve air when they come into
contact, some air will be taken into the boilers with the
feedwater and will pass over with the steam to the engine. Air
also may leak into the steam through the stuffing boxes of the
engine or turbine when run condensing. This air would soon
clog the condenser and prevent condensation of the steam if it
were not removed. The air is pumped out either with the con-
densed steam by means of a wet-air pump or else separately by
means of a dry-air pump.

The circulating water is pumped through the condenser by the
circulating pump. To maintain a high vacuum, the circulating
water must be at a low temperature when it leaves the condenser.
This means that each pound of circulating water can absorb only
a few B.t.u. Each pound of the steam that condenses gives up
to the water something like a thousand B.t.u. It therefore is
evident that a large volume of circulating water must be used.
As the pressure to be pumped against is small, a centrifugal pump
is commonly employed for forcing the circulating water through
the condenser. Air pumps are made both of the reciprocating
type and of the rotary type. The former are more common. The
design of the air pump is a rather difficult problem, since the air is
very rare at a high vacuum, so that if the pump
has much clearance it will fail to maintain this
vacuum.

76. The Jet Condenser. As stated before,
the steam and water mix in the jet condenser.
This mixture of condensed steam, injection
water, and the air contained in the steam and
water, may be pumped out by a wet-air pump.
However, the water is often allowed to run
out of the condenser by gravity through a ver-
tical pipe thirty feet or more in length that has
its lower end submerged. The air is pumped out
by a dry-air pump. This arrangement is com-
monly called a siphon or barometric condenser.
Figure 37 shows in section this latter type of jet condenser.
The injection water enters at A and runs over the edges of trays,
thus exposing a large surface to the steam which enters at B.
The air pump sucks the air out at the top through the pipe C.

FIG. 37

BOILER ACCESSORIES AND AUXILIARIES

In this type, the flow of the steam, until it is condensed, is with
the air and opposite to that of the water. Such a condenser
therefore is called a counter-current condenser.

Another type of jet condenser in which the air pump is dis-
pensed with is shown in Fig. 38. The air is carried along with the
condensed steam and the injection water. This is due to the high
velocity of the water as it passes out of the con-
stricted opening A. This is called the injector
or ejector type. It usually is furnished with a
barometric tube, as in the preceding type.

The jet condenser is more compact and less
expensive than the surface condenser. If it is
well made and equipped with a good air pump,
it will give a very high vacuum, but it mixes
fresh water with the condensed steam, and this
may cause scale if the mixture is used for the
boiler feed.

77. Cooling of Circulating Water. Since a
large amount of cold water is costly in some lo-
cations, it is sometimes necessary to cool the circulating water so
that it may be used over and over. This may be done by run-
ning it into a pond where the natural evaporation from the sur-
face will cool it. If the space for a large pond is not available,
the evaporation may be increased by the use of spray nozzles
which break the water up into a fine spray, thereby exposing a
large surface for evaporation. Another method for the rapid
cooling of the circulating or injection water is the use of cooling
towers. The water is allowed to trickle down over a lattice or
other surface, thus exposing a large surface for evaporation. Air
is either passed up through the tower by natural draft, or blown
through by means of fans. Where large bodies of cold water,
such as rivers or lakes, are available, the circulating water is
drawn from them and is thrown away after being used. If the
water is taken from lakes or rivers, it is often necessary to pass
it through a screen to remove such material as weeds, drift, and
fish. These screens require cleaning periodically and are some-
times made in the form of an endless chain, so that while some
sections are being cleansed others may be in use.

CHAPTER VI
THE STEAM ENGINE

78. History. In 1698 THOMAS S A VERY produced the first
steam engine that proved to be commercially successful. It was
used for pumping water. The engine consisted of two egg-shaped
vessels, each of which connected with the supply of water to be
pumped and to a steam boiler. In its operation steam was ad-
mitted to one of the vessels, and when it was full, connection
with the boiler was cut off. Cold water was then sprayed on the
outer surface of the vessel, causing the steam inside to condense
and to form a partial vacuum. This vacuum opened a valve in
the pipe leading to the well and sucked water into the vessel.
Steam was then turned on again, and the pressure forced the
water out of the vessel through the delivery pipe. When the
water had all been forced out, the process described above was
repeated. The engine was operated so that while one cylinder
was forcing, the other was sucking water. This engine is the
same in principle as the modern pulsometer.

Since the steam in the Savery engine came in direct contact
with the water during the forcing stroke, there was much loss
of steam by useless condensation. DENIS PAPIN, in 1705, made
an improvement on the Savery engine by making the steam vessel
of cylindrical shape and separating the steam and water by a
floating piston, thereby preventing a part of the unnecessary con-
densation.

About 1711, there came into use a machine that was known
as the Newcomen engine. THOMAS NEWCOMEN, with the aid of
JOHN GALLEY, and with certain ideas from Papin, made his engine
with a vertical cylinder into which a piston was fitted from the upper
end. The cylinder was placed directly above the boiler and con-
nected with it. Steam was admitted to the cylinder by the open-
ing of a valve placed between the boiler and the cylinder. The
piston was connected to a pump through a walking beam, one
end of the walking beam connecting to the pump rod and the
other to the piston rod. The beam was so counterbalanced that
it took but little steam pressure to force the piston up. Steam
was generated at about atmospheric pressure, and as the piston

THE STEAM ENGINE 77

moved up the steam valve was opened and steam filled the space
beneath it. When at the top of its stroke, water was sprayed
into the cylinder, condensing the steam and forming a partial
vacuum. This vacuum under the piston allowed the atmos-
pheric pressure from above to force the piston down.

While the Newcomen engine was an improvement over the
Savery engine, it was very wasteful of steam because the cylinder
was cooled by the spray of water on each downward stroke. Much
condensation of steam occurred in heating up the cylinder walls
on each upward stroke. While repairing one of these Newcomen
engines, JAMES WATT conceived ways in which it might be im-
proved. Patents covering these ideas were granted him in 1769.
Watt's chief aim was to keep the cylinder walls as hot as the
incoming steam at all times and thereby prevent the initial con-
densation that rendered the older engine so inefficient. This high
cylinder temperature was to be maintained, first, by condensing
the steam in a vessel away from the cylinder, and second, by a
steam jacket placed around the cylinder walls. Lagging was also
to be placed around the outside of the cylinder to keep down the
heat lost to the outside air. In the previous engines, the piston
was kept tight by a kind of water seal on top of the piston. Watt
used fibrous packing and tallow to keep the piston tight and
saved heat that previously was lost to the water above the piston.
In the operation of his engine Watt found it necessary to remove
the air from the condenser and so he equipped the condensers
with air pumps.

While it is seen that Watt is not the inventor of the steam
engine, yet it must be admitted that he did more to advance
its development than any other one man. Up to the time of
Watt, the steam engine was used almost exclusively for pumping
water in collieries, but he applied it to the driving of other forms
of machinery. After many hardships and discouragements, Watt
at last was able to produce his engine in large numbers. The engine
became increasingly popular, and we may say that the era of our
present industrial development started at the time of James Watt.
In applications for his patents Watt advocated the use of high-
pressure steam from which work could be obtained by using it
expansively, but in the actual construction of his engines he never
used pressures much above that of the atmosphere.

Since the time of Watt, various improvements have been made

78 ENGINES AND BOILERS

in the steam engine. The mechanical construction has been bet-
tered, the valve mechanism improved, compounding adopted, and
the steam pressures greatly increased. While its thermal effi-
ciency may not be as high as some forms of internal-combustion
engines, the steam engine is very reliable.

79. The Plain Slide-valve Engine. Figure 39 shows in ver-
tical and horizontal sections the parts of a simple steam engine.
Its action is as follows : Steam comes from the boiler through the
steam-pipe, and after passing the throttle valve, enters the steam
chest. The valve, driven by an eccentric on the shaft, moves
backward and forward on the valve seat, uncovering alternately
the two steam ports. When a steam port is uncovered by the
valve, the steam flows through the port into the cylinder and by
its pressure moves the piston in the cylinder. In Fig. 39, the
left steam port is shown partly open and the steam is then push-
ing the piston to the right. At the same time that the steam is
forcing the piston to the right, the valve has uncovered the right
port, so that the steam on the right of the piston may escape to
the exhaust pipe.

This motion of the piston is transmitted by the piston rod to
the cross-head and from this through the wrist pin and the con-
necting rod to the crank. The reciprocating motion of the pis-
ton is transformed by the connecting rod and crank into rotary
motion of the shaft. The power generated in the cylinder usually
is taken from the shaft by a belt on the flywheel or by an electric
generator coupled to the shaft. The valve is made to move so
that when" steam is being admitted to one end of the cylinder it
is being exhausted from the other. As the incoming steam is at
a much higher pressure than the exhaust, there is a resultant
force pushing the piston in the direction of the outgoing steam.

That end of the cylinder farthest from the crank is called the
head-end^ and the end nearest the crank the crank-end. With
the piston at the extreme left of its travel, the crank will be in
a direct line between the cylinder and the shaft. While the force
on the crank pin may be large, there is no turning effort. The
crank is then said to be on head-end dead center. With the pis-
ton at the extreme right of its travel the crank is on crank-
end dead center. When the engine is running, if the crank rises,
as the piston leaves the head-end dead center (i.e., if in Fig. 39

THE STEAM ENGINE

the crank moves in a clockwise direction), the engine is said to
be running over. If the crank moves in the opposite direction,
the engine is said to be running under.

FIG. 39

The stroke of the engine is the distance the piston travels in
half a revolution. It is equal to twice the length of the crank.
On the head-end stroke, the piston moves from head-end to
crank-end dead center, and the reverse motion takes place on
the crank-end stroke.

80 ENGINES AND BOILERS

80. Parts of the Steam Engine. CYLINDER. Steam-engine
cylinders are made of cast iron and the bore is carefully machined.
With proper lubrication, the surface exposed to wear acquires a
high polish and the metal is worn away very slowly. The ports
are cored in the casting and are not finished except along the
edges at the valve seat. At each end of the cylinder the diameter
is made slightly larger. This enlarged part is called the counter-
bore. It should extend far enough so that the piston ring comes
to its edge, in order that a shoulder may not be worn in the cylin-
der wall. The counterbore also serves a purpose when the cylinder
has to be rebored, since the boring machine may be set on the
counterbore, which will not be worn away, and thus the alignment
need not be lost. On some larger engines used in marine service,
a thin inner shell or liner is placed in the cylinder, so that it may
be replaced without changing the cylinder when it becomes worn.

The cylinder head is bolted to the cylinder, and the joint is
made steam-tight by means of a gasket or a ground joint. In the
smallest engines the cylinder is cast as a part of the frame, but
ordinarily it is cast separately and bolted to the frame.

In large engines where high efficiency is desired the cylinder
may have a steam jacket. The heads may or may not be jack-
eted. In any case the cylinder is covered by some non-conductor
of heat, called lagging. With proper lagging very little heat is
lost by radiation.

Unless the exhaust valves are so placed that any water in the
cylinder can drain to them, it is necessary to tap the bottom of
the counterbore at each end of horizontal cylinders and place a
drain cock there. These cocks are opened when the engine is
warming up so that the condensed steam will not be caught when
the piston comes to the end of the stroke. As water is incom-
pressible, its presence in too large a quantity would cause the
breaking or straining of some part. Sometimes automatic relief
valves are placed at the ends of the cylinders to take care of any
water that may get into the cylinders. Each end of the cylinder
is tapped for the connection of an indicator.

VALVES. The subject of valves will be taken up in detail in
Chapter VIII.

PISTON. Pistons for stationary engines are made of cast iron.
The piston is turned to a slightly smaller diameter than the

THE STEAM ENGINE 81

cylinder, and leakage of steam past it is prevented by the use of
piston rings which fit into grooves cut around the piston. In the
small sizes the rings are made in one piece slightly larger than
the diameter of the cylinder. A piece is then cut out of each
ring, and they are snapped into the groove in the piston.
Because of their elasticity, they spring out and make contact with
the cylinder walls. When worn they may be replaced by new
rings. On larger pistons the rings are built up in sections and
are pushed out against the cylinder wall by springs placed under
them.

On small engines the pistons are cast in one piece, and usually
are hollow, to make them as light as possible. In the larger
sizes, they are built up of two or more pieces. In vertical engines
and locomotives, the pistons sometimes are made dished to a
slight extent. The dishing may be to add strength, to shorten
the length of the engine a small amount, or to facilitate drainage.

PISTON ROD. The piston rod connects the piston to the cross-
head. It is made of steel, and the connection must be such that
it will be always tight. If a little play is allowed, a bad knock
develops that rapidly grows worse. In horizontal engines, a tail-
rod sometimes extends from the piston out through the head-end
cylinder head, and its outer end is carried by a slipper on a guide.
This arrangement allows the weight of the piston to be carried
by the slipper and the cross-head and lessens the wear on the
piston and the cylinder.

STUFFING Box. The joint between the piston rod and crank-
end cylinder head is made tight by a stuffing box. The packing
used on low-pressure engines in this box may be fibrous, but with
high steam pressure a metallic packing is commonly used. A good
packing should keep the joint steam-tight and at the same time
give but little friction on the rod. Wear in the cylinder, improper
adjustment of the cross-head, poor alignment, or a pitted or scored
rod, may cause excessive wear on the packing. It is then diffi-
cult to keep it steam-tight.

CROSS-HEAD. The cross-head with the cross-head pin or wrist
pin forms the connection between the piston rod and the con-
necting rod. The cross-head is made to move in a straight line
by guides on the frame of the engine. The two-guide type shown
in Fig. 39 is the most common, although four-guide and one-

82 ENGINES AND BOILERS

guide types are used occasionally. The slipper type also is often
used; in this the cross-head takes the form of a slipper which
slides on the flat surface of one guide. In all types the wear
between the cross-head and guides is taken up by the adjust-
ment of the wedge-shaped slippers or by the use of shims. In a
few steam engines and most gas engines a trunk piston is used.
In this type the piston itself acts as cross-head and carries the
wrist pin. With this latter arrangement, the engine is single
acting, i.e. the steam acts only on the head of the cylinder and on
the face of the piston.

CONNECTING ROD. The connecting rod connects the wrist pin
to the crank pin. It is alternately in compression and tension,
and usually is made of steel. On high-speed engines, considerable
bending stress may be developed in the connecting rod on account
of its fling, and hence its cross-section is usually rectangular or
I-section. On slow-speed engines this bending stress is small,
and the rod is made circular in section. Brass or other metal
bearing-pieces called brasses are used at the bearing points on the
pins. These brasses are often babbited. As wear occurs ad-
justment must be made to take it up. If the rod is shortened
in taking up the wear, the end on which such adjustment is
made is said to have an open stub-end. If the rod is length-
ened by an adjustment at one end, that end is called a closed
stub-end. In Fig. 39, the end at the crank pin is of the marine
stub type, which is used on all center-crank engines. In this
last type the wear is taken up by removing liners, and in the
former types usually by means of wedges v/hich move the brasses.

CRANK. The crank pin is made of steel, and it may be a part
of the same forging as the crank and shaft, or it may be set into
crank discs which are keyed to the shaft. When the pin is placed
between two crank discs, as in Fig. 39, we have a center-crank
engine; when it overhangs one crank disc, we have a side-crank
engine.

COUNTERBALANCE. The crank pin and half of the connecting
rod usually are considered as rotating parts and must be coun-
terbalanced to make the engine run smoothly. Furthermore the
piston, piston rod, cross-head, and half of the connecting rod have
a reciprocating motion. It takes a large force to start and stop
them on each stroke. Unless they are counterbalanced the whole

THE STEAM ENGINE 83

engine will vibrate on the foundation. While it is impossible to
counterbalance exactly both the rotating and the reciprocating
parts at the same time, yet it can be partially done. The proper
sized counterbalance or counterweight (Fig. 39), sometimes made
of lead but usually of iron, is put on to give smooth running.

SHAFT. Engine shafts usually are made of steel. As explained
above, they are either forged to make the crank and crank pin
integral parts of the shaft, or else the crank is keyed to the shaft.
In addition to the key, a shrunk fit sometimes is used, or the
crank disc may be pressed on by hydraulic pressure.

BEARINGS. With a side-crank engine there is one main bear-
ing, and with a center-crank engine, there are two. The weight
of the flywheel may be carried partly by an outer bearing called
an outboard bearing. There will be wear on the main bearing
in a vertical direction on account of the weight of the flywheel
and of the rotating parts, and there will be wear in a horizontal
direction on account of the thrust from the piston. Many main
bearings are made up of four parts, the cap, the bottom part
which takes the vertical wear, and two side-pieces which take
the horizontal wear. The latter are called quarter-boxes. These
parts may be adjusted separately.

FLYWHEEL. The turning moment on the crank varies at differ-
ent parts of the stroke. At dead center it is zero. In order to
keep the shaft turning at approximately the same speed at all
times in the revolution, a flywheel is put on the shaft. This acts
to store up and give out energy at the proper times, thereby keep-
ing the angular velocity approximately uniform. The flywheel may
carry the belt or there may be a separate belt wheel in addition
to the flywheel, in which case the latter is sometimes called a
balance wheel. The flywheel commonly is made of cast iron.
In the smaller sizes it is cast in one piece, but in the larger sizes,
it is cast in sections. Great care must be taken in its manufac-
ture, since a crack may cause a disastrous accident.

ECCENTRIC. An eccentric is placed on the shaft or governor arm
to drive the valve. It is encircled by the eccentric strap, which
is connected to the valve by means of the eccentric rod and
valve stem. This is really a substitute for a crank and connect-
ing rod and gives to the valve a motion similar to that of the
piston.

84 ENGINES AND BOILERS

FRAME. The frame is made of cast iron in stationary engines,
and the better engines have the heavier frames. The greater the
weight of frame the more smoothly the engine will run, other
things being equal. On some small high-speed engines there is a
cast-iron sub-base placed between the frame and the foundation.
Frames are given different names according to their shape and
the cylinder arrangement.

FOUNDATION. Foundations usually are made of brick or con-
crete. The latter is now the more common. The frame is fas-
tened to the foundation by anchor bolts. The foundation should
be quite massive and should rest on soil that is firm enough to
carry the weight of the engine and foundation without settling.

81. Piston Displacement. The volume the piston displaces
in moving from one dead center to the other is called the piston
displacement. It is commonly expressed in cubic feet. The head-
end piston displacement is equal to the length of stroke in feet
times the area of the piston in square feet. The crank-end piston
displacement is the area of the piston minus the area of the cross-
section of the piston rod, times the stroke. .

The size of an engine is given in inches, the diameter of the
cylinder bore first and the length of the stroke second, e.g. an
18 // X24' / engine is one whose cylinder is 18" in internal diameter
and whose stroke is 24". The size of a compound engine is given
by the diameter of the high-pressure cylinder, the diameter of the
low-pressure cylinder, and the stroke, as 10"X18' / X24' / . If the
volume is calculated in cubic inches, remember to change tc cubic
feet in giving the piston displacement.

82. Clearance. When the piston is at the extreme end of its
travel, there will be some volume back of it, because it is neces-
sary to have a little space in which to take up the wear on the
connecting rod brasses and to allow for unequal expansion of
parts as the engine heats up, and because of the space in the
ports. This volume, the larger part of which is often the volume
of the ports, is called clearance. Clearance is expressed as a per-
centage of the piston displacement. Thus, when we say that the
head-end clearance of an engine is 4.5 percent, we mean the
volume back of the piston when the engine is on head-end dead
center is .045 times its head-end piston displacement.

To determine the clearance of an engine, first place the engine

THE STEAM ENGINE 85

on the dead center of the end for which the clearance is to be
measured. Second, disconnect the valve from the eccentric
rod, move it so that the port for that end is closed, and block it
up in that position. It is necessary to disconnect the valve be-
cause the port is usually open a small amount when the engine
is on dead center. Third, pour water into the opening for
attaching the indicator cock until the clearance space is full of
water. If the valve is placed at the top of the cylinder, as is
the case with horizontal Corliss engines, remove the valve and
pour the water into the port. Having recorded the amount of
water poured in, and having made correction for leakage, the
volume of w r ater necessary to fill the clearance space is computed.
This volume divided by the piston displacement for that end gives
the clearance. A rough method of computing clearance from the
indicator diagram is given in 89.

83. Steam Back of Piston during Stroke. The weight of the
dry steam back of the piston may be computed for any percent
of the stroke if we know the clearance, the size of the engine,
and the steam pressure at that percent of the stroke. Add the
percent of the stroke to the percent of clearance. Multiply this
by the piston displacement, and divide by 100. The result is
the volume of steam back of the piston at the given percent of
stroke. By means of the indicator, the pressure may be determined
for the same position. The density of steam may be read from
the steam tables for that particular pressure. The product of
this density and the volume back of the piston gives the weight
of the steam there.

EXAMPLE. What is the weight of dry steam back of the piston at 27%
of the crank-end stroke of a 14"X16" engine? The engine has a 2" rod and
the crank-end clearance is 6.3%. The crank-end indicator card is at hand.

SOLUTION. Measure the pressure from the card at 27% of the stroke.
Suppose this pressure is 115 pounds gage, and the atmospheric pressure is
14.5 pounds per square inch. The absolute pressure is then 115 + 14.5 = 129.5
pounds per square inch. The density of dry saturated steam at this pres-
sure is, from the steam tables, .2887 The piston displacement is seen to be
16(7r7 2 TT)/ 1728 = 1.398 and the volume back of the piston is therefore equal
to (.27 + .063) XI. 398 = .466 cubic feet. The weight of dry steam is then
.2888 X .466 = . 1 342 pound.

84. The Indicator and Its Purposes. The steam-engine indi-
cator was first used by JAMES WATT, who invented it. Since his
time it has been perfected and is now very extensively used.

86 ENGINES AND BOILERS

The indicator records on paper a line showing the relation between
the pressure in the cylinder and the movement of the piston.
The diagram or card produced is of value in setting the valves, in
computing the horsepower developed in the cylinder, and in mak-
ing analyses of the operation of the engine. A description of the
mechanism of the indicator will not be given here. When we
speak of the scale of spring of an indicator, we do not mean the
actual scale of the spring used in the indicator, but rather the
relation between the movement of the pencil on the indicator dia-
gram and the pressure that causes it. That is, a 60-pound indicator
spring is one that gives a pencil movement of one inch for each
60 pounds per square inch increment of pressure in the cylinder.
The heavy lines of Fig. 40 show a representative diagram as
it comes from the indicator. It is seen that the diagram is a
closed irregular curve with a straight line underneath. The
Cuf . off straight line is the atmospheric-

pressure line, which is used for
reference in measuring pres-
._ ____ + ^ e/eeje sures. The ordinates of the
points on the curve show to
* some scale the pressures in the

jp IGt 40 cylinder, and the abscissas the

movement of the piston. Start-

ing at the upper left-hand corner of Fig. 40, we have the pressure
back of the piston when it is on dead center. As the piston
moves forward, the pressure changes as shown by the upper curved
line of the diagram. It is seen that the pressure drops but little
for the first part of the stroke, and later more rapidly, until at the
end of the stroke it is nearly down to that of the atmosphere.
On the backward stroke, the pressure remains nearly constant
until near the end, when it rapidly rises to the initial point.

85. Events of Stroke. There are four events of the stroke.
1. Admission occurs when the valve uncovers the port and allows
the steam to enter the cylinder. 2. Cut-off takes place when
the valve closes the port and prevents any more steam from
entering. 3. At release, the valve uncovers the port and allows
the steam to escape to the exhaust. 4. Compression occurs when
the valve again closes the port and prevents any more steam
from leaving the cylinder for the remainder of the stroke.

THE STEAM ENGINE 87

It is seen that the steam enters the cylinder from admission to
cut-off, and that the steam thus let in expands and does work
on the piston from cut-off to release. From release to compression,
the used steam is being exhausted from the cylinder. The steam
that is caught when the exhaust port closes is compressed into the
clearance space during the time from compression to admission.

86. Location of Events on Diagram. After a little practice
the events may be quite accurately located on the ordinary in-
dicator diagram or card. In Fig. 40 it is seen that the upper
line drops down and that there is a point of inflection in this
curve. This is the point of cut-off. This point of inflection is
easily detected by drawing a continuation of the two curves as
shown by the dotted lines at the point of cut-off. Release occurs
at the next point of inflection; it may be located in a manner
similar to that in which we located the cut-off. At compression
there is no point of inflection; therefore its proper location is diffi-
cult. The exhaust valve ordinarily closes rather slowly, and the
passageway for steam being small when it is nearly shut, the pres-
sure may start to rise even before the valve is completely closed.

A common error is that of taking the point of compression
too low on the compression curve. Admission occurs where the
compression curve stops and the straight line starts.

To determine the percentage of stroke at the different events,
draw the two end-ordinates. The distance between these, / in
Fig. 40. represents the length of stroke to some scale. The dis-
tances from the left end-ordinate to the events shown by a, b, e,
and d, divided by the distance /, give the ratios of the stroke
at these events, and the percentages of stroke are those ratios
multiplied by 100.

87. Equation of Expansion and Compression Curves. There
is a definite relation between pressure and volume during expan-
sion and compression. The equation pv n = p\v\ n = p 2 V2 n expresses
this relation, where p is the absolute pressure on these curves,
v is the volume back of the piston (including the clearance space),
and n is some constant exponent for each individual curve. An
analysis of many cards shows that n is sometimes a little less
than 1 and sometimes a little larger than 1. For rough calcu-
lations, it may be considered equal to 1. With this assumption,
the equation of the expansion and compression curves is pv=C.

ENGINES AND BOILERS

This is the equation of the equilateral hyperbola. Certain geo-
metric facts about this curve are of importance to us. In Fig. 41,
let us choose any two points on an equilateral hyperbola: E,
whose coordinates are (pi, vi), and H, whose coordinates are
(P2, ^2)- We shall show that the line of the diagonal CA of the
rectangle EC HA drawn through these points, passes through the
origin. In the triangles OAB and OCD, OB = Vi, BA=p 2 ,
OD=vz, and DC =pi. Since their sides are parallel, these two
triangles are similar. Hence we have

yo/ume t/j Ct/jb/c

which satisfies the equation of the curve.

It follows that we can construct the entire curve if one point E

on it is given. We can locate other points on it in the following

manner. Through the point E
draw horizontal and vertical
lines. Choose some point C on
the horizontal line, and draw
the line OC through the origin.
Drop a vertical line from C,
and draw a horizontal line

jj x / x " ' ' ' f%r through A, the point where the

line OC cuts the vertical line
through E. The intersection H
of the vertical line through C

and the horizontal line through A is a point on the curve. Other

points may be found in the same manner.

Another geometric fact of value is that the length of FE is

equal to HG on a line drawn through the points E and H. This

may be proved readily from Fig. 41. The area under the curve

from E to H, i.e. BEHD, may be determined by integration.

The increment of this area has dimensions p and dv, and dA

is equal to pdv. Hence the total area is

A = \ dA = \ pdv,
but since p\v\ = p%V2 = pv, we have

and

THE STEAM ENGINE

88. Hypothetical Indicator Diagram. Diagrams are some-
times constructed on the hypotheses that the expansion and com-
pression curves are equilateral hyperbolas, that the pressure dur-
ing the admission of steam from the end of the stroke to cut-off
is constant, and that the back pressure is constant up to the
point of compression. Such a diagram will be called the hypo-
thetical diagram. This diagram is also sometimes called the
theoretical, or ideal, or conventional diagram.

The construction of such a diagram, shown in Fig. 42, is car-
ried out as follows. First choose a suitable length / for the dia-
gram, and draw in the atmospheric-pressure line. Next choose

a scale of pressures, ^ ^

and draw the volume ~~~ e ' - -

axis at a distance c, the
atmospheric pressure,
below the atmospheric
line. Draw the pres-
sure axis at a distance
from the point F of the
diagram equal to the
ratio of clearance times
the length /of the dia-
gram. Measure up from
the atmospheric-pres-
sure line the initial steam pressure, and draw the steam-admis-
sion line FA. The length of FA is equal to the ratio of cut-off
times the length /. From the point A, construct an equilateral
hyperbola as in 87. The length e is equal to the ratio of release
times /. From B, the point of release, draw a line to the end of
the diagram at C. The distance h from C to the atmospheric-
pressure line represents the back pressure. From C to D draw the
back-pressure line parallel to the atmospheric line. From D,
the point of compression, construct an equilateral hyperbola to
the point E, whose distance a from the end of the diagram is the
admission distance. Connect E and F by a straight line.

The actual diagram may vary considerably from the diagram
just constructed. The speed of the engine, the throttling of
steam in the ports and by the valve, the condensation in the
cylinder, etc., affect the form of the actual diagram, which may
resemble the dotted diagram FGHCIJF in Fig. 42. If the en-

in tsto/c feef
FIG. 42

ENGINES AND BOILERS

FIG. 43

gine exhausts into a condenser in which a vacuum is maintained,
the back-pressure line will fall below the atmospheric-pressure line.

89. Determination of Clearance from Card. Since it is often
impossible to measure the clearance of an engine when a test is
being made, a rather rough approximation sometimes is used.
Suppose the indicator diagram is as shown in Fig. 43. Two
points A and B are chosen on the compression curve. On these

two points a rectangle is
drawn, and the diagonal is
extended until it cuts the vol-
ume axis, which is drawn
below the atmospheric line at
a distance equal to the bar-
ometric pressure. The point
of intersection of this diagonal
and the volume axis establishes the origin, and the pressure axis
may be drawn. The distance c divided by the length of card
gives the ratio of clearance. The pressure axis may also be
located by drawing the other diagonal through A and B and
laying off DA equal to BE. If the piston rings are not tight
in the cylinder, or if the valve leaks steam, this method will not
give even approximately correct results.

90. Determination of the Mean Effective Pressure. It has
been mentioned that the indicator diagram shows the pressure
at all points of the stroke.

Since the total pressure on the
piston times the distance the
piston moves is the work done
by the steam, we see that the
indicator card is a work dia-

FIG. 44

gram.

In order to calculate the work
done in the cylinder we must
know the average effective pressure, or the mean effective pres-
sure (m. e. p.) on the piston. In Fig. 44 when the piston is at G
on the forward stroke, the steam pressure is a. When the piston
is at G on the backward stroke, the steam pressure is 6. During
the forward stroke the steam is working on the piston, but on
the backward stroke the piston is doing work on the steam.

THE STEAM ENGINE

Hence the effective pressure for the two strokes at G is a b,
which is shown by the dotted ordinate G H.

To get the average, or mean, of these effective pressures for the
whole card, we may proceed as follows. Draw in the end-ordi-
nates of the diagram. By means of a scale or the edge of a
ruled piece of paper, divide the length of the card into a number
of divisions of equal length. The number of divisions should be
more than eight, and need not be more than fifteen. Through
these division points, draw in ordinates as shown by the full
lines. The diagram is thereby cut into a number of strips of
equal width. The average height of each strip is about equal
to the dotted ordinate located midway between the solid lines,

' *3vm / Jotted orSin'*t*s'

FIG. 45

i.e. the average height of the strip CEFD is GH. While this
may not be true in every case, the fact that some parts of the
diagram are concave while other parts are convex will tend to
neutralize the error when the dotted ordinates are averaged.
Next add the lengths of the dotted ordinates. This is best done
by laying a strip of paper on the diagram and marking the dif-
ferent ordinates directly on the edge of the strip. Figure 45
shows the strip of paper with the dotted ordinates added graph-
ically. The average of the ordinates is this sum divided by their
number. This average height of card, multiplied by the scale
of spring, gives the mean effective pressure (m. e. p.).

If the m. e. p. of many cards is to be found, it may be quicker
to use a polar planimeter to get the area of the card. This area
divided by the length gives the mean height. This mean height
multiplied by the scale of spring gives the m. e. p. This method
is usually no more accurate than the former, when the former
is done with reasonable care.

92 ENGINES AND BOILERS

91. Indicated Horsepower. The mean effective pressure
(m. e. p.) in pounds per square inch times the area of the piston
in square inches gives the total average effective force exerted
by the steam on the piston during the forward and backward
stroke, or for one complete revolution of the crank. The work
done by this force is equal to the force times the distance the
piston moves. It must be remembered that the effective force
on the piston was obtained by taking the difference of pressures
during the forward and backward strokes. With this in mind,
we see that the work done on the piston per revolution is equal to
the m. e. p. times the area of the piston, times the length of the stroke.
If the length of stroke is expressed in feet, the result will be in
foot-pounds.

If N denotes the revolutions per minute (r. p. m.), L the length
of stroke in feet, P the m. e. p. in pounds per square inch, and A
the area of piston in square inches, the foot-pounds of work done
per minute is equal to PL A N. The horsepower of the engine is
then PLAAV33000. Since this result is obtained by means of
the indicator, it is called the indicated horsepower (i. hp.).

If the engine is double-acting, the i. hp. for the crank-end is
found in a similar manner, taking the m. e. p. from the crank-
end card and using the area of the piston on the crank-end, which
will be less than that for head-end because the piston rod occu-
pies some of the area of the piston. For very rough work, the
average of the m. e. p. for the two ends is sometimes taken and
the area of the piston rod is neglected; then we have, approxi-
mately,

total indicated horsepower = QQQQQ *

EXAMPLE. The head-end m. e. p. is 42.6 and the crank-end m. e. p. is
45.1 pounds per square inch in a 12" X 18" engine running at 220 r. p. m. The
diameter of the piston rod is two inches. What is the indicated horsepower?

SOLUTION. The area of a 12" circle is 113.1 square inches and of a 2"
circle is 3.1 square inches. The area of the head-end of the piston is then
113.1 square inches and of the crank-end 113.13.1 = 110.0 square inches.
The stroke is 18", or 1.5'; hence we have

and

. . ,., 42.6X1.5X113.1X220 .__,
head-end i. hp. = - 33QQO - =48.2 hp.,

45.1X1.5X110.0X220
crank-end i. hp.= 33000 =49.7 hp.

whence the total i. hp. is 48.2+49.7 = 97.9 hp.

THE STEAM ENGINE

92. Brake Horsepower. It is usually not difficult to take
cards from an engine and to compute the i. hp. from them. This
does not give the horsepower that the engine is actually deliv-
ering, of course, but that which is developed in the cylinder.
In testing an engine that is in actual use, it is often impossible
to measure its actual output without considerable trouble and
expense. Under such conditions one must be content with the
i. hp. If the engine is not too large and conditions will permit,
the actual power delivered by the engine is often measured. If
it is direct-connected to an electric generator and the losses in the
generator are known, this is fairly simple. If not, some form of
dynamometer may be used. The most common means of meas-
uring the delivered horsepower is by a brake on the flywheel.

FIG. 46

If a number of tests are made, the Prony brake is generally used.
For an occasional test a rope brake may answer the purpose.

Figure 46 shows a common form of Prony brake. Two or more
bands of strap iron with wooden blocks fastened to them are
put around the circumference of the flywheel or the belt wheel.
The tension in the band is regulated by means of a hand-wheel
shown at the top of the figure. Two arms are fastened to the
band which at the right end in the figure carry a knife-edge that
rests on a block placed on a platform scales. The flywheel
rotates in the direction of the arrow, and the friction between
the blocks and the wheel causes a pressure on the scales.

When the engine is not running, there will be some pressure
on the scales due to the weight of the brake-arm. This may be
determined by weighing the brake before putting it on the wheel
and balancing it so that its center of gravity is determined. If

94 ENGINES AND BOILERS

the weight of the brake is w and its center of gravity is located
at a distance b from the knife-edge and at a distance a from the
center of the wheel, we have, taking moments about the center
of the wheel,

wa = component of weight on scales X (a +6).
From this we can compute the effect of the weight of the brake
arm on the scales. This component of weight must be subtracted
from the pressure on the scales when the engine is running. If we
denote by P the net pressure on the scales, which is due to the
friction between the blocks and the wheel with the engine running,
we may compute the power being absorbed by the brake as fol-
lows. The work absorbed by the brake per revolution is equal
to PX2?rr, in which r is the horizontal distance from the center
of the wheel to the knife-edge and is known as the radius of the
brake-arm. If the engine is running N revolutions per minute, the
work absorbed per minute is 'ZnrPN, and the horsepower absorbed
is 27rrPAV33000. This is called the brake horsepower (b. hp.).
The radius of the brake wheel R does not enter into the compu-
tation of the b. hp. To prevent the burning of the wooden blocks,
the wheel is kept cool by putting water inside the rim which
evaporates and carries off the heat.

EXAMPLE. During the test of an engine, the r. p. m. was 220. The pres-
sure of the brake arm on the scales was 318 pounds (Fig. 46). The weight of
the entire brake is 118 pounds, and its center of gravity (c. of g.) is 6.05
feet from the knife-edge. The radius of the brake-arm, r, is 7.16 feet. Find
the brake horsepower.

SOLUTION. If the c. of g. of the brake is 6.05 feet from the knife-edge, it
is 7.166.05 = 1.11 feet from the center of the wheel when the c of g. is in
line between the knife-edge and the center of the wheel. That part of the
weight of the brake supported by the scales is 118X1.11/7.16 = 18.4 pounds.
The net pressure on the scales then equals 318 18.4 = 299.6 pounds. The
brake horsepower=27rrPAY33000 = 27rX7.16X299.6X220/33000 = 89.7 hp.

93. Mechanical Efficiency. The ratio of the brake horse-
power to the indicated horsepower is called the mechanical effi-
ciency. It is usually expressed in the form of a percentage.
The difference between the indicated horsepower and the brake
horsepower is the frictional horsepower.

EXAMPLE. If the i. hp. of an engine is 97.9, and the b. hp. is 89.7 hp.
find the mechanical efficiency and the frictional horsepower.

SOLUTION. The mechanical efficiency = 89.7/97.9 = .916 = 91.6%. The
frictional horsepower = 97.9 - 89.7 = 8.2 hp. Hence the frictional horsepower
= 8.2/97.9 = 8.4% of the indicated horsepower.

THE STEAM ENGINE 95

94. Thermal Efficiency. In general the efficiency of a ma-
chine is the ratio of the out-put to the in-put. In the steam
engine heat is put in and mechanical work is taken out. The
thermal efficiency of the steam engine is the ratio of the work got
out to the work equivalent of the heat put in. The thermal effi-
ciency may be based on either the i. hp. or the b. hp. The latter
is called the overall efficiency, and is equal to the thermal efficiency
based on i. hp. times the mechanical efficiency.

A common but approximate way of stating the efficiency of a
steam engine is to give the weight of dry steam consumed per
hour per horsepower. This may be based on either the i. hp.
or the b. hp. It is not an exact way of stating the efficiency
because steam may contain different amounts of heat, depending
on pressures and superheat. Efficiency may also be expressed
in terms of B.t.u. per minute per horsepower.

In computing the B.t.u. given to the engine in a unit of time,
proceed as follows. The weight of dry steam is found by deduct-
ing the weight of moisture in the steam from the amount of
wet- steam used. The heat in this moisture is not charged to
the engine, since it is not possible for the engine to extract work
from it. From the steam tables find the total heat in a pound
of dry steam, or the total heat in a pound of superheated steam
if superheated steam is used. From this total heat per pound,
subtract the heat of the liquid at the pressure of the exhaust.
The reason for subtracting the heat of the liquid at the pressure
of the exhaust is that although the engine has used the steam,
the heat of the liquid can be saved by feeding the condensed
steam back to the boiler, which is often done. Whether it is
done or not, it is not fair to charge this heat to the engine. Mul-
tiply the amount of heat in a pound of dry steam that is charged
to the engine by the weight of dry steam used in unit time. The
result gives the B.t.u. upon which the efficiency is computed.

EXAMPLE. An engine during a test developed 97.9 i. hp. and 89.7 b. hp.
The engine used 3060 pounds of 97% quality steam per hour. The steam
pressure was 125 pounds gage, and the engine exhausted to the atmosphere.
The barometer reading was 29.3 inches. Find the thermal efficiency of the
engine.

SOLUTION. The weight of dry steam used per hour is 3060 X. 97 = 2970
pounds. 125 pounds gage pressure = 125+ 14.4 = 139. 4 pounds per square
inch absolute. At this pressure, the total heat in a pound of steam is 318. 2 -j-
872.3 = 1190.5 B.t.u. The heat of the liquid at the atmospheric pressure is

96 ENGINES AND BOILERS

179.3, therefore the heat to be charged to the engine per pound of dry steam
used is 1190.5-179.3 = 1011.2 B.t.u. The dry steam used per hour per i. hp.
is 2970/97.9 = 30.35 pounds. The work delivered per hour per horsepower =
33000X60 foot-pounds. The foot-pounds of energy equivalent to the B.t.u.
supplied per hour per horsepower is 778X1011.2X30.35. Therefore, since
the efficiency is the out-put divided by the in-put,

This is based on the i. hp. The dry steam used per hour per b. hp. is
2970/89.7 = 33.1 pounds; hence the thermal efficiency based on the b. hp. is

33000X60
778X1011.2X33.1

In this solution the factor 33000X60/778=2545 will always
occur in the equation for thermal efficiency. Hence the formula
may be written in the form

4i i a 2545

thermal efficiency = - . p . , , , -- - - >

nX B.t.u. per pound of dry steam

where n is the number of pounds of dry steam used per hour per
horsepower.

The thermal efficiency of a steam engine will seldom, if ever,
exceed 25 per cent. This may seem to be a very low value, but
it is impossible for the engine to use a very large part of the
heat supplied due to the fact that the exhaust steam carries
with it its heat of vaporization. On account of condensation of
steam in the cylinder and other causes of heat loss, the efficiency
of a reciprocating engine seldom approaches the efficiency of an
ideally perfect engine working under the same range of pressure,
but a well-designed steam turbine of large size may do so.

95. Cylinder Condensation. The largest single loss in the
average engine is due to what is known as initial condensation.
Since the cylinder walls are made of iron, which is a good con-
ductor of heat, they naturally absorb heat from any hotter
body or substance placed in contact with them and they give
up heat to a cooler body. The steam comes into the cyl-
inder at a relatively high pressure and temperature. Both the
pressure and the temperature drop in the cylinder, and the steam
leaves at a relatively low pressure and temperature. Since the
cylinder walls are exposed first to hot, and then to cool steam,
their temperature will never be as great as that of the incoming
steam, nor as low, during operation, as that of the outgoing steam.

When the steam first enters the cylinder and strikes the cooler
walls, a part of its heat will be absorbed by the walls. This

THE STEAM ENGINE 97

causes a partial condensation of the steam. Since the engine
operates by virtue of the steam pressure and volume, it is readily
seen that a shrinkage in volume causes a loss of work, and a lower-
ing of efficiency. By the time the steam leaves the cylinder, it
is cooler than the cylinder, and it takes back some of the heat
it gave to the walls, but at too late a time to avoid the loss in
efficiency. Depending upon the type of engine and the condi-
tions of operation, the condensation may continue until release
occurs, or re-evaporation may start during the expansion of the
steam between cut-off and release. By computing from the in-
dicator diagram the weights of steam at cut-off and at release,
we find a net condensation during expansion if the weight at
release is less than at cut-off, and a net re-evaporation if the
weight at release is greater than at cut-off. The computation of
condensation or re-evaporation during expansion is of little value
since most of the re-evaporation occurs after release and before
the steam leaves the exhaust ports.

96. Steam Accounted for by the Indicator Diagram. The

A. S. M. E. code for testing steam engines calls for the computa-
tion of the steam accounted for by the indicator diagram at points
near the cut-off and release. Mark the points of cut-off and
release and a point on the
compression curve where we
are sure the exhaust valve is
closed, as in Fig. 47. Find the
ratio of stroke at these points.
The volume back of the piston
at cut-off is the ratio of stroke
at cut-off plus the ratio of p IG 47

clearance, shown by a, times

the piston displacement. Scaling the pressure at cut-off from the
diagram, we may compute the weight of dry steam back of the
piston at cut-off by means of steam tables.

Not all of the steam back of the piston at cut-off entered on
that one stroke from admission to cut-off, since some of it was
in the cylinder during compression. The amount that was ad-
mitted is the weight at cut-off minus the weight caught at com-
pression. The weight of steam compressed may be computed
in a manner similar to that at cut-off. The weight of steam
per hour accounted for by the indicator diagram is then equal to

ENGINES AND BOILERS

FIG. 48

FIG. 49

FIG. 50

FIG. 51

FIG. 52

FIG. 53

FIG. 54

FIG. 55

FIG. 56

FIG. 57

THE STEAM ENGINE 99

(Wc.o.-TF C omp.)XWx60/i. hp., where W c . . and TF comp . are the
weights of steam back of the piston at cut-off and compression,
respectively. The weight is calculated separately for the head-end
and crank-end of the cylinder, and the two values added to give
the total for the engine. The weight accounted for at release is
computed in the same way, but the two results usually differ slightly
on account of the net condensation or the re-evaporation during ex-
pansion from cut-off to release. The weight of steam accounted for
by the indicator diagram will be considerably less than the actual
amount used by the engine, because of the initial condensation
of steam when it first enters the cylinder.

97. Valve-setting from the Indicator Diagram. It has been
mentioned previously that one of the uses of the indicator is to
assist in the setting of the valves While the subject of valve-
setting will not be discussed thoroughly here, faulty setting may
be recognized from the appearance of the diagram. Figure 48
shows the effect when the admission valve opens too soon.

Figure 49 shows the results of late admission. Due to the
tardiness of the valve opening, the steam is throttled, and the
pressure for a large part of the stroke is lowered considerably.

Figure 50 shows the effect of early compression. The steam
caught when the exhaust valve closes is compressed to a pressure
above that in the steam chest, the admission valve is lifted off
its seat, and some of the steam escapes into the steam chest.

Figure 51 shows almost no compression. This would cause no
harm in a very slow-speed engine, but with higher speeds the
steam caught at compression acts as a cushion and makes for
smooth running.

Figure 52 shows too early a release, and Fig. 53, too late a
release. Both cause a loss in the area of the diagram.

Figures 54 and 55 show unequal cut-off in the two ends of the
cylinder. The crank-end is doing a much larger proportion of
the work. The work done by the two ends should be about equal.

Figure 56 shows improper lubrication of the indicator piston
or the binding of some part. A wavelike motion of the curve is
sometimes noticed when the diagram is taken from a high-speed
engine, due to the vibration of the indicator spring, but it differs
materially from Fig. 56. In Fig. 57, the indicator drum is strik-
ing the stop on account of improper adjustment of the length of
the cord connecting the indicator and the reducing mechanism.

CHAPTER VII
COMMON TYPES OF STEAM ENGINES

98. Slide-valve Engine. Where simplicity and reliability
are of more importance than high efficiency, the slide-valve en-
gine is used. The simplest type of slide-valve was shown in Fig.
39, and the principles of its operation were explained to some
extent in the previous chapter. There are many varieties of
slide-valves, the more common of which will be described later.
Most of the smaller stationary engines in use are equipped with
the slide-valve, and all American locomotive engines are of this
type.

While the plain slide-valve is very simple, it has certain defects.
One of these is the impossibility of obtaining the proper steam
distribution at all loads, i.e. of making the events of stroke occur
at the proper place to give the highest efficiency at light load
and also at heavy load. Various modifications and improvements
have been made on the slide-valve to remedy this defect, the
chief one of which is to place a second slide-valve on top of the
main valve, and to control cut-off by a rider.

Another defect of the plain slide-valve is the slowness with
which the steam ports open up and close at some loads, which
cause what is known as wire-drawing. This is simply a throttling
of the steam by the valve as it enters the cylinder. This throt-
tling usually causes a lowering of the efficiency of the engine.

99. The Corliss Engine. By far the most common type of
high-grade reciprocating stationary steam engine in this country
is the Corliss engine. The name comes from its inventor and
first producer, GEORGE CORLISS, an engineer and engine builder of
Providence, R. I. There are two distinguishing features of this
engine. The first of these is the oscillating cylindrical valve.
The second is the means for disengaging the valve from the mech-
anism that drives it, and the quick closing of the valve after its
disengagement. To understand the Corliss valve mechanism
thoroughly, it is necessary to make a rather thorough analysis of
its motion. We shall not do this in this chapter. The general
principle of operation of the gear is fairly simple, however.

Figure 58 shows a typical Corliss engine cylinder. The right
end is cut in section to show the construction of the valves and

100

COMMON TYPES OF STEAM ENGINES

101

their locations relative to the cylinder. The left end shows an
ordinary form of the mechanism that moves the steam valves and
the exhaust valves. An eccentric on the shaft is connected to
the hook rod which operates the valves through an eccentric rod
and rocker arm. This gives the hook rod a horizontal recipro-
cating motion that is nearly harmonic. The hook rod is attached

- 3 fa am fff>e

FIG. 58

to a wrist plate which is pivoted to the cylinder at C. The wrist
plate is thereby given an oscillating motion.

Four rods are attached to the wrist plate. The two upper
rods are the steam rods, which transmit the motion to the two
steam valves, and the lower or exhaust rods drive the two exhaust
valves. The steam rods are attached to bell cranks or double
arms that are pivoted on the valve spindle but are not attached
to it. It is thus possible for the wrist plate and the bell crank

102 ENGINES AND BOILERS

to move without affecting the valve in any way. The cylindrical
steam valve has a spindle which extends out of the steam chest.
To the outer end of this spindle there is keyed a steam arm.
Any motion of this arm causes the valve to move. A block is
attached to the back side of the steam arm, and a hook which is
carried by the upper arm of the bell crank catches over it. As
the steam rod moves to the right, the steam arm is picked up
and the valve is turned. After having lifted the steam arm a
certain distance, the hook is made to disengage -with the block
and the steam arm is released.

The steam arm is connected to the piston of a dash-pot by a
dash-rod. As the steam arm is raised, a partial vacuum is
formed in the dash-pot. When the steam arm is released from
the hook, it is suddenly pulled downward by the vacuum in the
dash-pot. As the steam arm is lifted, the valve opens and admits
steam to the cylinder. When it is pulled down, the valve is
closed suddenly, giving a quick cut-off. The time at which the
hook is made to release the steam arm is controlled by a cam
whose position is regulated by the governor. This cam engages
with the tail of the hook and causes the disengagement. At light
loads, the trip occurs soon and an early cut-off is given, and the
cut-off is retarded as the load of the engine is increased.

Figure 59 shows the trip mechanism on a larger scale. DAC
is the bell crank. As the point D moves backward and forward
in a nearly horizontal direction, the point C moves up and down
in a nearly vertical direction. The pin C carries the steam hook.
The tail of the hook engages with the knock-off cam, and its jaw
engages with the block attached to the steam arm at B. For
any one load the knock-off cam is stationary, and as C goes up,
the tail of the hook is pushed away from A by the cam, which
causes the latch to disengage with the block B. When there is
a heavy load on the engine, the governor rod moves to the left,
which raises the knock-off cam and makes the trip come later,
giving a long cut-off. There is also a safety cam, shown in Fig.
59. If the governor fails to rotate, the safety cam comes into
contact with the tail of the hook and prevents the picking up of
the steam arm and therefore causes a failure to admit steam to
the cylinder.

There is no disengagement between the exhaust arm and the
exhaust valve, so that the events of release and compression occur

COMMON TYPES OF STEAM ENGINES

103

at the same ratio of the stroke for all loads. Both the steam
valve and the exhaust valve of Fig. 58 are double-ported, which
gives twice the opening for the passage of steam with the same
valve movement as with the single-ported type.
From the description just given it is seen that cut-off is inde-

To forer/ior

FIG. 59

pendent of the other events, and that the steam valve closes
quickly, thereby preventing wire-drawing at closing. If a care-
ful analysis of the motion is made, it will be seen that the steam
valve opens the port nearly as widely at light loads as at full
loads. The force necessary to operate the valve mechanism is
not large and the work done in moving the valves is a very small
part of the total output of the engine.

100. The Four-valve Engine. With a single slide-valve,
the changing of one event necessitates the changing of all the
others. To avoid this difficulty, engines are often made that

104

ENGINES AND BOILERS

have four valves, a steam valve for each end of the cylinder, and
an exhaust valve on each end. The exhaust valves are driven
by a fixed eccentric, so that the release and compression are the
same at all loads. The steam valves are controlled by the gov-
ernor; hence cut-off and admission will vary for different loads.

f FIG. 60

Figures 60 and 61 show one style of four- valve engine. This par-
ticular engine has oscillating cylindrical valves similar to those
shown in Fig. 58 for the Corliss engine. Some makers, probably
to make use of the enviable reputation of the Corliss engine, call
this type a non-releasing Corliss. This engine lacks, however,
the distinct advantage of the trip found in the true Corliss type.

FIG. 61

101. The Compound Engine. When all the expansion of
steam takes place in one cylinder, we have what is known as a
simple engine. If the steam passes through two successive cyl-
inders, the engine is said to be a compound engine. If there are
three successive cylinders, it is called a triple-expansion engine.

COMMON TYPES OF STEAM ENGINES 105

If there are four successive cylinders, it is called a quadruple-
expansion engine, etc. An engine may have two cylinders and
not be compound, i.e. it may be a twin-cylinder engine, in which
half the steam passes through one cylinder and half through the
other. Likewise a compound engine may have three cylinders,
all the steam passing through one high-pressure cylinder, and then
dividing, half passing through each of the two low-pressure
cylinders.

The purpose of compounding is to reduce the initial conden-
sation. It does not necessarily follow that there is a greater
ratio of expansion of steam in a compound engine than in a
simple engine, since that depends upon the point of cut-off. We
have seen that initial condensation is caused by the range in
temperature within the cylinder. The temperature range is less
in each cylinder of a compound engine than in the single cylinder
of a simple engine of the same capacity. Since the amount of
condensation does not vary directly as the total temperature
range, there may be considerably less total condensation if the
steam is passed through two successive cylinders than if all the
expansion occurred in one cylinder.

Several years ago the idea of compounding was very popular
and was carried to the extreme. Many triple-expansion engines,
and some quadruple-expansion engines, were built. Experience
proved, however, that there was a practical limit to which the
idea might be carried. Now stationary engines are seldom built
with more than two pressure-stages, except in direct -acting pumps.
In the marine service, the triple-expansion engine is still popular,
partly for the reason that it is desirable to have three cranks on
the same shaft to give a greater uniformity of torque on the pro-
peller shaft, and partly on account of the uniformity of load on
marine engines. Of the many types of compound engines that
have been built, only two are in common use in land service at
present. We shall proceed to consider these.

102. The Tandem-compound Engine. In the tandem-com-
pound engine, the pistons of the two cylinders are placed on the
same piston rod, as shown in Fig. 62. The cylinder to the left
is the high-pressure cylinder, and the one to the right is the low-
pressure cylinder. The steam ports of the high-pressure cylinder
are at a and 6; the exhaust ports of the low-pressure cylinder are

106

ENGINES AND BOILERS

at c and d. The pistons, in Fig. 62, are shown moving to the
right. Steam is entering the high-pressure cylinder through a,
and leaving it through /. The exhaust steam from the crank end
of the high-pressure cylinder passes to the head end of the low-
pressure cylinder either directly or through a stationary vessel
called a receiver, i.e. the back pressure on the piston A is the
forward pressure on the piston B. On the return stroke, steam

FIG. 62

enters the port b and the exhaust from the high-pressure cylinder
leaves through e and enters the low-pressure cylinder at h either
directly or through the receiver. With the tandem arrangement
only one cross-head, connecting rod, crank, and frame are needed.
In locomotive work, the Baldwin or Vauclain compound engine
is sometimes seen. In this engine, the cylinders are placed side
by side, and both piston rods attach to the same cross-head.
The method of steam distribution is similar to that of the tan-
dem type. Figures 63 and 64 show the high-pressure and low-

FIG. 63

FIG. 64

FIG. 65

pressure indicator diagrams, as taken from a Baldwin compound
engine. The high-pressure card comes from the crank end of the
high-pressure cylinder and the low-pressure card from the head
end of the low-pressure cylinder. These cards were taken with
springs of different scales. Figure 65 shows the same diagrams when
drawn to the same scale of pressure. It is noticed that the back-
pressure line of the high-pressure card parallels the admission
line of the low-pressure card from the left end up to the point
of cut-off in the low-pressure cylinder. The reason for this is

COMMON TYPES OF STEAM ENGINES

107

obvious, since the exhaust from the high-pressure cylinder passes
directly to the low-pressure cylinder. When the admission
valve of the low-pressure cylinder closes, compression must
necessarily start in the high-pressure cylinder. Since it is neces-
sary, with such a high pressure at exhaust as exists in the high-
pressure cylinder, to have the compression occur late, it follows
that cut-off must come very late in the low-pressure cylinder.
This is not an ideal condition, but it is necessary if no receiver
is placed between the two cylinders. If a receiver were placed
between the two cylinders so that it could act as a reservoir into
which to discharge, and from which to draw steam, it would not
be necessary to have the preceding relation between compression
and cut-off.

103. The Cross-compound Engine. In the cross-compound
engine (Fig. 66), each cylinder has its own cross-head, connect-

FIG. 66

ing rod, crank, and frame. The cranks are usually spaced 90 apart.
A by-pass is arranged so that the engine can be started even if
the high-pressure cylinder stops on dead center, by admitting
steam directly to the low-pressure cylinder.

Each cylinder has its own valve mechanism, and the exhaust
from the high-pressure cylinder passes into a receiver from which
the low-pressure cylinder takes its steam. This arrangement per-
mits a better steam distribution than that used in the tandem

108

ENGINES AND BOILERS

type without a receiver. If the receiver is quite large, the back
pressure in the high-pressure cylinder during exhaust will be
nearly constant. Figures 67 and 68 show the indicator diagrams
from a cross-compound engine. It should be noticed that the
engine exhausts into a condenser.

104. Cylinder Ratio. The cylinder ratio of a compound engine
is the ratio between the piston displacements of the low-pressure and
the high-pressure cylinders. While it is not essential that the length
of stroke be the same for both high-pressure and low-pressure

cylinders of a cross-
compound engine, they
are made so. The cyl-
inder ratio is then the
ratio of the squares of
the diameters of the
low-pressure and high-
pressure cylinders.

105. The Combined
Indicator Diagram.

The combined diagram
is constructed by plot-
ting both cards to the
same scale of pressure
and volume. Usually
we do not change the
low-pressure diagram
but change the scale of
the high-pressure card
conform to it. Figure
69 shows the combina-
tion of diagrams of Figs.
67 and 68. The low-
pressure diagram is
identical with Fig. 68, while the length of the high-pressure dia-
gram equals the length of the low-pressure diagram divided by
the cylinder ratio. The high-pressure diagram is placed to the
right of the pressure axis its clearance distance, i.e. its distance
from the axis equals the ratio of the high-pressure clearance
times the new length of the high-pressure diagram.

Condenser Pressure.

FIG. 69

COMMON TYPES OF STEAM ENGINES 109

106. Diagram Factor. The definition of the diagram factor
as given in the 1915 edition of the A. S. M. E. Power Test Code is
as follows :

The diagram factor is the proportion borne by the mean
effective pressure measured from the actual diagram to that
of a hypothetical diagram which represents the maximum power
obtainable from the steam accounted for by the actual diagram
at the point of cut-off; assuming first, that the engine has no
clearance; second, that there are no losses through wire-
drawing the steam either during admission or release; third,
that the expansion line is a hyperbolic curve; and, fourth,
that the initial pressure is that of the boiler, and the back
pressure that of the atmosphere for a non-condensing engine,
and of the condenser for a condensing engine.
To determine the steam accounted for by the actual diagram
at the point of cut-off, draw hyperbolic curves through the point
of compression P and the point of cut-off (Fig. 70) until they

FIG. 70

cut the boiler-pressure line at R and S. The length of RS
is the length of the admission line for the hypothetical diagram,
FA in Fig. 42, drawn to proper scale. The hypothetical diagram
is drawn as in Fig. 42 except that the boiler pressure is taken as the
initial pressure, release comes at the end of the stroke, the back
pressure is the atmospheric pressure (condensing pressure in a con-
densing engine), there is no compression, and there is no clear-
ance. The hypothetical diagram for the combined diagrams of
Fig. 69 is shown dotted. Since we assume there is no clearance,
the length of the hypothetical diagram is equal to that of the
low-pressure card. The distance RS at boiler pressure is deter-
mined from the high-pressure diagram, as in Fig. 70. From S
to T construct a hyperbolic curve, using the origin 0', and not 0.
Release is at the end of the stroke and the back-pressure line is
at the condenser pressure.

*-*'*' /}b
i<$ CX^Trar

110 ENGINES AND BOILERS

In Fig. 69, the mean effective pressure of the combined dia-
grams and of the hypothetical diagram are in the same ratio as
the areas of the combined and hypothetical diagrams, because
they are of the same length. To find the diagram factor of the
combined cards, divide their area by the area of the hypothetical
diagram.

107. Ratio of Expansion. The A. S. M. E. Power Test
Code gives the following rule:

To find the percentage of cut-off, or what may best be termed
the commercial cut-off, the following rule should be observed:

Through the point of maximum pressure during admission
draw a line parallel to the atmospheric line. Through a point
on the expansion line where the cut-off is complete, draw a
hyperbolic curve. The intersection of these two lines is the
point of commercial cut-off, and the proportion of cut-off is
found by dividing the length measured up to this point by

e total length>

To find the ratio of expansion, divide the volume correspond-

bg to the piston displacement, including clearance, by the
olume of the steam at the commercial cut-off, including clear-

ance.

In a multiple-expansion engine the ratio of expansion is found

by dividing the volume of a low-pf essure cylinder, including

clearance, by the volume of the high-pressure cylinder at the

commercial cut-off, including clearance.

108. The Unaflow Engine. The unaflow, or uniflow, engine
is shown diagrammatically in Fig. 71. There is an admission
valve at each end of the cylinder. The exhaust steam escapes
through a port located around the circumference of the cylinder
midway between the two ends. The piston, which is longer than
in most engines, itself uncovers the exhaust port at about 90
per cent of the stroke. Compression must start when the piston
is at the same place on the return stroke. Under non-condensing
conditions this would give a very excessive compression pressure;
hence the engine normally is run condensing, under which con-
ditions the compression pressure is moderate. The thermal
efficiency is about the same as that of a compound engine. The
gain in efficiency over the ordinary double flow engine is due
to the reduction of initial condensation. The condensation

COMMON TYPES OF STEAM ENGINES

111

is reduced with the unaflow principle because the ends of
the cylinder are kept hotter than the central portion. High-
pressure steam never comes in contact with the central part of
the cylinder and the flow of steam is from the ends toward the
middle. The exhaust steam passing out through the central port
does not cool the walls as much as it would if it flowed back to

the ends of the cylinder upon leaving. In actual engines, pro-
vision must be made for relieving the excessive compression pres-
sure, should the vacuum break. This is done by a relief valve
that adds to the clearance or allows the compressed steam to
re-enter the steam chest, or by adding an auxiliary exhaust port
nearer the end of the cylinder, which is opened automatically
when the vacuum fails.

CHAPTER VIII
VALVES

109. Introduction. From our previous study of the steam
engine we have learned that the purpose of the valve is to admit
steam to the cylinder, and to release steam from it. The time
at which the events occur must be such that the engine is capable
of doing the work required, and that it may have as high an
efficiency as possible under the conditions of operation. An en-
gine may run with the valves improperly set or designed, but
more steam will be used than if the valves functioned properly.

110. The D Slide-valve. The engine of Fig. 39 has what
is commonly known as a D slide-valve. The valve slides back
and forth on its seat, alternately opening and closing the ports.
An eccentric on the shaft drives the valve. The eccentric rod
is usually quite long in comparison with the throw of the eccen-
tric, so that the valve may be considered to have the same motion
as the horizontal component of the eccentric.

In Fig. 72a, the valve is shown in mid-position; consequently
the eccentric will either be directly above or directly below the
center of the shaft. In mid-position, the valve laps over the
edges of the port; the amount it extends over on the steam side
is called the steam lap, and on the exhaust side, the exhaust lap.

At the right side of Fig. 72a is shown the relative position of
the eccentric and the crank. The eccentric leads the crank by
angle 8. This angle 8 will be the same at all times during the
revolution. In Fig. 726, the crank is on head-end dead center,
and the valve is uncovering the head-end port a small amount.
The amount the port is open when the crank is on dead center
is called the lead. It is measured in inches. The valve in Fig.
726 is to the right of its mid-position by an amount equal to the
steam lap plus the lead. In moving from the position shown
in Fig. 72a to that in Fig. 726, the eccentric has moved a hori-
zontal distance equal to the steam lap plus the lead, and has turned
through a certain angle which is called the angle of advance.
It is evident that the eccentric is ahead of the crank by an angle
of 90 plus the angle of advance. It is customary to speak of
the angle of advance and -not of the whole angle 6.

Figure 72c shows the valve at head-end admission. The valve

112

VALVES

113

FIG. 72

114 ENGINES AND BOILERS

is on the point of opening the head-end port for the admission
of steam, and is traveling to the right. In the admission posi-
tion it is to the right of its mid-position by a distance equal to
the steam lap.

Likewise the eccentric will be to the right of its mid-position
by a horizontal distance equal to the steam lap. The crank
is back of the eccentric by the angle 6 and is seen to be
approaching its head-end dead-center position. If there is
any lead, the crank will never be quite up to its dead-center
position at admission.

At head-end cut-off, Fig. 72d, the valve is to the right of its
mid-position by a distance equal to the steam lap and its direc-
tion of motion is to the left. The position is the same as for
admission, but it is going in the opposite direction. The eccen-
tric is now below the center line of the shaft.

Figures 72e and 72/ show the relative positions at head-end
release and compression. At both of these events, the valve is
at the left of mid-position by an amount equal to the exhaust-
lap distance. It is moving to the left at release and to the right
at compression.

In all the six diagrams of Fig. 72, the positions of the piston
and its direction of motion are shown. The cylinder section in
each diagram is in a horizontal plane and therefore is at an angle
of 90 from the diagram showing crank and eccentric positions

For an understanding of the slide-valve and the analyses to follow,
it is essential that the student have a precise conception of the rel-
ative positions of the valve on the seat, the eccentric and the crank
relative to the center positions, and the position of the piston in the
cylinder.

111. Relative Motion of Crank and Piston. Since the con-
necting rod is not very long compared with the crank arm, we
cannot assume that the horizontal movement of the crank is the
same as the piston movement. This is clearly seen from the
diagram in Fig. 73. As the crank moves from A to C, the cross-
head moves from E to D. That part of the stroke completed
by the cross-head is a'. It is evident that a' is considerably
larger than the horizontal movement of the crank in going from
A to C.

In our analysis of valve motions, it is not customary to draw

VALVES 115

in the cross-head D to find the proportion of the stroke at dif-
ferent crank positions, but the following scheme is used. The
horizontal diameter of the crank circle AB is extended to the
left. With the length of connecting rod DC as a radius at the
desired scale, and with D as a center, strike the arc shown by
the dotted line CG. This gives the distance AG = a, on the diame-
ter of the crank circle, that is equal to ED = a', the movement of

FlG. 73

the piston or cross-head from E to D. In the valve analysis it
is not necessary to draw the crank circle to any particular scale
if we keep the proper ratio between the lengths of the crank and
the connecting rod. This ratio usually is expressed as R/L.
No matter what the scale of the crank circle may be, the ratio of
a to the length of stroke will be constant.

112. Valve Diagrams. Many diagrams have been used to
show graphically the relation of the movement of the valve to
the movement of the piston, or of the relative movements of
eccentric and crank. Only those most commonly used in this
country will be explained here, i.e. the valve ellipse, the Bil-
gram diagram, and the Zeuner diagram.

113. The Valve Ellipse. In this diagram the system of rec-
tangular coordinates is used. The valve displacement is plotted
vertically and the piston displacements are plotted horizontally.
On the left of Fig. 74 is shown a crank and eccentric. The eccen-
tric is ahead of the crank by 90 plus a.

With the crank at C, the piston is at a distance x from the cen-
ter of the stroke. At the same time, the valve is at a distance y
from its mid-position. If we plot x against y, we get a point G.

116

ENGINES AND BOILERS

The coordinate axes are the horizontal and vertical diameters
of the crank circle.

On the right of Fig. 74 this same operation is carried out for
twelve crank positions with their corresponding eccentric posi-
tions. The crank positions are denoted by Ci, 2, C%, etc., and
the corresponding eccentric positions by E\, E^ E%, etc. Plotting
the displacements, we get the points 1, 2, 3, etc. Connecting the
points thus found by a smooth curve, we get what is known as
a valve ellipse. It is evident from Fig. 74 that this is not a true

fro re/ ' ~ - -^

FIG. 74

ellipse. It would have been but for the distortion due to the
short length of the connecting rod.

The upper half of the ellipse ABD represents the valve move-
ment to the right of its mid-position. The lower half, DFA,
represents the valve movement to the left of its mid-position.
From A to B it gives the movement from mid-position to ex-
treme right; from B to D, from the extreme right to mid-posi-
tion; from D to F from mid-position to extreme left; and F to
A, from the extreme left to mid-position.

The valve ellipse of Fig. 74 is reproduced in Fig. 75. Four
horizontal lines are drawn through the ellipse. The head-end
steam lap is the distance from the top line to the horizontal axis.
The crank-end steam lap is the distance from the bottom line
to the axis. The head-end exhaust-lap line is drawn below the
axis and the crank-end exhaust-lap line above. When the valve
has moved to the right a distance equal to the head-end steam
lap, head-end admission takes place. Admission is shown by the

VALVES

117

point H on the ellipse, and the crank position corresponding to
H is determined by projecting vertically from H to the axis.

As the valve moves from its extreme right position back to
mid-position, head-end cut-off takes place. This is shown by the
point / on the ellipse. The crank position corresponding to I is
found by projecting down from I to the axis and striking an
arc upward from this point to the crank circle. The radius of
the arc is the length of the connecting rod. In like manner, the

fxg/w

Ce.mof.ft.0.

FIG. 75

crank position at head-end release is found from the intersection
of the head-end exhaust-lap line with the ellipse at J.

The head-end compression point is at M , where the exhaust-
lap line cuts the ellipse. The crank-end events are determined
from the points K, L, N, and P. The vertical distance from the
top point of the ellipse to the head-end steam-lap line is the head-
end maximum port-opening, and the head-end lead is given by
the distance from the extreme left point of the ellipse to the
head-end steam-lap line. The crank-end maximum port-opening
and lead are found in a similar manner.

In actual use, the ellipse is rather burdensome because it takes
considerable time to construct it. It is evident also that the
crank position at admission is not easily determined with accuracy.
The ellipse is little used except in locomotive work.

114. The Bilgram Diagram. On the left of Fig. 76, the
crank and the eccentric are shown by C and E, respectively.
The displacement of the valve from mid-position is y. The dis-
tance y is laid off perpendicular to the crank, and a line is drawn
parallel to the crank 'at a distance y from it.

118 ENGINES AND BOILERS

At the right of Fig. 76, this has been done for twelve crank
positions. It is seen that these lines all pass through two points
P and P', and that a line drawn from P to P' passes through the
center of the crank circle and makes an angle a with the hori-
zontal. Hence the perpendicular distance from the points P or
P' to the crank at any position is the distance that the valve
is from mid-position. The points P and P' are called the con-
struction points in the Bilgram diagram.

Figure 77 shows the application of the Bilgram diagram.
About the point P draw two circles, one whose radius is equal
to the head-end steam lap, and the other with a radius equal

FIG. 76

to the head-end exhaust lap. Draw the crank-end lap circles
with center at P'. The crank positions tangent to these lap
circles give the positions at the different events. Head-end ad-
mission occurs when the valve is at a distance from its mid-
position equal to the head-end steam lap, and when the valve
is going away from its mid-position. The head-end admission
position shown in Fig. 77 fulfills these conditions. At that
time the crank is at a distance from the point P equal to the
head-end steam lap, and further motion moves it farther from P.
The crank position at head-end cut-off is tangent to the head-end
steam-lap circle on the other side, i.e., the valve is then at a dis-
tance equal to the steam lap from mid-position and further mo-
tion brings the valve nearer mid-position.

The crank positions for the other events are shown in Fig. 77.
Reasoning similar to the preceding will show them to be correct.
It is customary to draw the head-end lap circles about P, and the

VALVES

119

crank-end lap circles about P f , although there is no inherent
reason for so doing.

Half of the valve-travel minus the steam lap equals the maxi-
mum port opening if the valve has no over-travel, i.e. if it does
not move beyond the far edge of the port. Therefore the maxi-
mum port-opening is as shown in the figure; It is remembered
that the port is open a distance equal to the lead when the
crank is on dead center. In other words the valve is then at a
distance equal to the steam lap plus the lead from its mid-position.
Therefore the perpendicular distance of P from the horizontal

c,e. t a

FIG. 77

axis is equal to the steam lap plus the lead. The distance of
the steam-lap circle from the axis is then s the lead.

115. The Zeuner Diagram. On the left of Fig. 78, the valve
displacement y is laid off radially on the crank from the center
outward. This gives a point G on the crank. This has been done
for twelve crank positions on the right of Fig. 78 and the points
connected by a smooth curve. The points fall on the circum-
ferences of two equal circles, the diameter of each of which is
one-half the valve-travel. The line which forms the diameters
of these two circles makes an angle a with the vertical. The

120

ENGINES AND BOILERS

circle whose center is at A shows the movement of the valve to
the right of its mid-position, and is called the right valve-circle.
The other circle is called the left valve circle; it shows the move-
ment of the valve to the left of the mid-position. When the
crank is drawn in any position, the displacement of the valve
is given by the distance from the center of the crank circle to
the intersection of the crank with the valve circle.

The application of this diagram is shown in Fig. 79. With
the crank on head-end dead center, the eccentric is at E, at an
angle a to the right of the vertical. The diameters of valve
circles, P'P', are at an angle a on the other side of the vertical.

FIG. 78

The extremity P of the diameter of the valve circle is called the
construction point in the Zeuner diagram. The lap circles are
drawn as shown. The head-end steam-lap circle intersects the
right valve-circle at the point T. The crank position through
T is then head-end admission, because with the crank in this
position the valve is at a distance equal to the steam lap to the
right of mid-position. The crank position at cut-off is drawn
through the point K, where the steam-lap circle intersects the
right valve-circle. Head-end release and compression occur when
the head-end exhaust-lap circle intersects the left valve-circle.

It may be proved by means of the similar triangles OWE and
PRO or by actual construction that a line drawn from A to B is
tangent to the steam-lap circle at W . This line is perpendicular
to the diameter of the valve circles. In like manner a line drawn
from I to F is tangent to the head-end exhaust-lap circle, and is
perpendicular to the diameter of the valve circles. The same
thing is true of the lines HG and JD for the crank end. It is

VALVES

121

often better to draw the lines A B, IF, HG, and JD than to
determine the crank positions at the events by the intersections
of the valve circle and the lap circle.

ON is equal to the steam lap plus the lead, because the valve
circle cuts the crank on dead center at N. But ONP is a right
triangle, since it is inscribed in a semicircle. If QS is drawn
parallel to AW, the triangle OSQ is a right triangle, and it is
similar and equal to the triangle, ONP. Therefore OS is equal

Right w/re ctrck

be. c.o.

fivnfr circle
Eccentric c/n,/e

4t. comp
t.rel.

ce. co.-

c/rc/e

FIG. 79

to the steam lap plus the lead, and WS is equal to the lead. If
we then draw a circle about Q as a center, tangent to the line
AW, its radius is the lead. This is called the head-end lead
circle. The crank-end lead circle is drawn about X tangent to
HG. The head-end maximum port opening equals PW, which is
one-half the valve-travel minus the steam-lap. The line PK is
perpendicular to the crank at cut-off, because PKO is a right
angle since it is inscribed in a semicircle.

The application of these valve diagrams to practical problems
will show their value. Space will not be taken here to give the
solutions of the various common problems in which these dia-
grams are used. The Bilgram and Zeuner diagrams are both

122

ENGINES AND BOILERS

adapted to problems of valve setting, but the Bilgram diagram is
the more convenient for use in designing valve gears.

116. Types of Slide-valves. The simple D slide-valve has
been discussed and its action explained. This type of valve is
much used, but it has certain defects which have been overcome
in other types. One of the defects of the simple D valve is the
large force necessary to move it when high steam pressure is used.
The steam pressure on the back of the valve presses it against
the seat. This pressure times the coefficient of friction between
the valve and the seat is the force that must be exerted to move
the valve. The work done in operating the valve is the force
times the distance the valve is moved. If either the force or the
distance is decreased, the work necessary to operate the valve will
be lessened.

117. Valve with Pressure Plate. The pressure on the back
of the valve may be removed by putting a pressure plate above
it, somewhat in the manner shown in Fig. 80. A steam-tight fit

between the valve and the
plate is made by strips set into
s | ots j n t h e va i v e. These are
pressed up against the plate
by springs from beneath, and
they act in much the same way
as rings on a piston. If any
steam leaks by the strips, it
may escape to the exhaust
through a vent in the valve.
This scheme enables us to remove as much of the pressure from
the back of the valve as we desire, but some pressure downward
is desirable in order to keep the valve firmly seated. Many flat
slide-valves have pressure plates. Aside from removing the pres-
sure, the plate does not affect the valve in any way.

118. The Piston Valve. Instead of a flat valve such as we
have considered, a piston valve is used extensively. Figure 81
shows a form of this type where the valve is cylindrical and slides
in a cylindrical chamber. It is readily seen that it is perfectly
balanced, since the steam causes no thrust either endwise or
on the seat. Piston valves are very easy to operate, but are
liable to leak steam when they become worn. Many of them

FIG. 80

VALVES

123

have rings similar to piston rings to keep this leakage of steam
to a minimum. If rings are used it is necessary to bridge the
ports. A broken ring is liable to cause severe damage and care
must be exercised to keep them in good condition.

In Fig. 81, the steam is led to the inside of the valve, so that
the steam lap is on the other side of the port from the valves
previously considered. A valve so constructed is said to be an

FIG. 81

indirect valve, or to have inside admission. Most piston valves
have inside admission, but this is not a necessity. It is easier
to keep the stuffing-box tight against exhaust steam than against
high-pressure steam. With the inside admission arrangement,
also, the live steam has less surface exposed to radiation. With
an inside admission valve, the eccentric follows the crank by an
angle of 90 a. The valve diagrams studied will have the same
form as before, but what was right-hand is now left-hand, i.e.
in the Zeuner diagram, for instance, what was formerly the right
valve-circle is now the left, but otherwise there is no change in
the diagram and no difference
is made in the solution of a
problem.

119. Double-ported Valves.
As has just been mentioned,
the work required to move the
valve is the product of the
force required to move it and
the distance it is moved. The

distance may be cut in half by making the valve double-ported.
Figure 82 shows a so-called double-ported valve. It is seen that
for a certain valve movement twice the area for the passage of
steam is given in this type compared with a simple slide-valve.
There are many different forms of double-ported valves, but they
are much the same in principle as that shown in Fig. 82.

FIG. 82

124 ENGINES AND BOILERS

120. The Gridiron Valve. If the idea of a double-ported
valve is carried a step farther, we may get a very large aggregate
opening for the passage of steam with but a small movement
of the valve. Figure 83 shows a gridiron valve. In many valves
of this type there are a large number of openings, whereas Fig. 83
shows only three. A valve of this character can have no exhaust
functions, and separate exhaust valves must be provided. If one
exhaust valve takes care of both ends of the cylinder, the engine
is called a two-valve engine; if there is a separate exhaust valve

FIG. 83

for each end, it is called a three-valve engine. When gridiron
valves are used, it is more common to have a steam valve and
an exhaust valve for each end of the cylinder. The engine
then has four valves.

If a governor is so constructed that it regulates the amount
of steam admitted to the cylinder by changing cut-off, it is evi-
dent from the valve diagrams that the events of release, com-
pression, and admission are changed when cut-off is changed if
a single valve is used. Under some conditions this is a serious
defect, and it makes the use of separate steam and exhaust valves
desirable.

121. The Riding Cut-off Valve. To utilize the expansive
force of the steam in an engine, it is necessary to have an early
cut-off. With a single valve, early cut-off will necessitate either
early release or early admission. With an early cut-off and with
release near the end of the stroke, compression is bound to occur
too soon for satisfactory operation under non-condensing con-
ditions. The student only needs to draw a valve diagram to con-
vince himself of this fact. To use an early cut-off, and to have
at the same time reasonable percents of release and compression,

VALVES

125

a riding cut-off valve is often used. There are several forms of
riding cut-off valves, but we shall describe only one of them.

Figure 84 shows the Myers riding cut-off valve. A main valve
slides on the seat in the same manner as an ordinary D valve.
The steam lap is made small so that the proper relation exists
between the events of admission, release, and compression. If
the main valve were acting alone, cut-off would occur very late.
To give an early cut-off, a rider valve which controls only the
event of cut-off is placed on the back of the main valve. The
working edge of the rider valve effects cut-off when it matches
with the edges of the main valve at B and at D.

In Fig. 84 both valves are shown in their mid-position. This
would not occur normallv unless one of the valves were discon-

FIG. 84

nected from its eccentric, but the figure is drawn in this manner
to give a clearer idea of the laps. Each valve is driven by its
own eccentric. The rider valve is made in two parts and the
relative position of the two parts may be changed by revolving
the valve stem. One part of the rider valve is secured to the
valve stem by a right-hand thread and the other part by a left-
hand thread. The hand wheel to the left is arranged so that by
turning it the valve stem is rotated and the parts of the valve
are brought nearer together or moved farther apart, thereby
effecting a change in cut-off. With the parts farther apart, cut-
off occurs earlier. When the valves are both in mid-position,
as shown in Fig. 84, it is easily seen that the rider valve has
negative steam lap or steam clearance.

To determine the crank position at cut-off from the valve dia-
grams, it is necessary to consider the relative motion of the two
valves. Figure 85 is the Zeuner analysis for the rider valve.
The crank circle and the two eccentric circles are shown by the
light lines. The point PI is the extremity of the diameter of the

126

ENGINES AND BOILERS

right valve-circle for the main valve, and a\ is the angle of ad-
vance for the main eccentric. The point P% is the extremity of
the diameter of the right valve-circle, and a 2 is the angle of advance
for the rider-valve eccentric. If a number of crank positions be
chosen and the displacement of the rider valve relative to the
main valve be laid off on the crank from the center radially
outward, a number of points will be established. Connecting

Relative
ffiqht l/e/ve Ctrt/e

C.e. f?/cfer fa/re
Steam top (neyafiri)

ff/JfrVafoe
c.e a/t-o/f.

FIG. 85

these points by a smooth curve, we find that it is composed of
the two circles shown by heavy lines in Fig. 85. The point PS
is the extremity of the diameter of the right relative valve-circle
and 3 is the angle which that diameter makes with the vertical.

It is seen that the diameter of the right relative valve-circle,
OPs, is equal in amount and parallel to the dotted line PiP2.
Knowing this fact, the solution of a rider-valve problem is quite
simple, for it is not necessary to plot the points to determine
the relative valve-circle.

Since the steam lap for the rider valve is negative, the intersec-
tion of that lap circle with the left valve-circle gives the crank
position at head-end cut-off. The cut-off positions of the crank
are shown by the heavy lines in Fig. 85. The crank positions
for admission, release, and compression are determined from a
valve diagram for the main valve in exactly the same manner as
previously explained for the D valve.

122. Effect of Rocker Arm on Location of Eccentric. In

the previous discussion of slide-valves, it has been supposed that

VALVES

127

the eccentric rod is attached directly to the valve rod, and that
the movement of the valve is the same as the horizontal move-
ment of the eccentric. Quite often a rocker arm is interposed
between the eccentric and the valve rods, in which case it may
be necessary to modify our previous assumption.

In Fig. 86, three arrangements of rocker arms are shown. In
I, the eccentric rod and the valve rod are both connected to the
same pin at B, and our assumption is not changed.

In II, the arm B AC reverses the motion of the eccentric. In
this case, if a direct valve is used, the eccentric must be placed
on the shaft at 180 from the position it would have had without
the rocker arm, i.e. if a direct valve is used, the eccentric must
follow the crank by an angle of 90 a.

If an indirect valve is used with a reversing rocker, the eccen-

Eccentr/e /foe/

Eccentric flotf

n
FIG. 86

trie is placed 90+ a ahead of the crank. This modification does
not effect the work in making the valve analysis.

In case III, the two arms of the rocker AB and AC are not
of the same length. The travel of the valve is the diameter of
the eccentric circle times AB/AC.

123. Oscillating Valves. One of the most common valves is
cylindrical and oscillates or rocks on a cylindrical seat. A spindle
fastened to the valve extends out of the steam chest and carries
an arm that is moved back and forth by the eccentric. The
Corliss valve shown in Fig. 58 and the valves of Figs. 60 and 61
are of this type. It would be possible to make an oscillating
valve control both the admission and exhaust events, but this is
seldom done. Where the oscillating valve is used, four valves
usually are employed. Because of the distortion of motion due
to the valve arm, it is not possible to show by the valve diagrams

128

ENGINES AND BOILERS

the exact motion of the valve. However, the relation of horizontal
motion of the eccentric still holds, and so the valve diagrams are
of value in the analysis of these valves.

124. Poppet Valves. While they are not common for steam
engines in this country, most gasoline engines are equipped with
poppet valves. This is a lifting valve and there is no sliding of
the valve on the seat. Figure 87 shows this type of valve. These
valves do not have to be lubricated and would seem to be well
adapted to conditions where highly superheated steam is used,
since one of the troubles met with in the use of superheated steam
is the difficulty of proper lubrication of the valve. Where poppet

valves are used on steam engines,
they usually are operated either by a
cam or by an eccentric from a lay
shaft that is parallel to the axis of
the cylinder. The lay shaft usually
is driven by mitre gears from the
main shaft.

125. Reversing. The direction of
rotation of an engine may be changed
FIG. 87 by shifting the eccentric to the proper

position. Occasionally an ordinary

engine must be reversed. Unless the eccentric is keyed to the
shaft, this is not usually a difficult task. With a direct valve,
the eccentric leads the crank by an angle of 90+ a. This is true
irrespective of the direction of rotation. To reverse, then, move
the eccentric in the direction the engine has been running through
an angle of 180 2 a. The same rule applies with an indirect or
inside-admission valve.

With certain classes of engines, such as are used for locomotives,
in marine work, etc., reversing is a common occurrence. Some
handier and quicker means must then be provided than that men-
tioned above. The devices used for this purpose are called re-
versing gears. There are a very great many types of reversing
gears in use, but space will not permit the discussion of more
than the most common types.

126. The Stephenson Link. One of the most widely used
reversing gears is the Stephenson link gear, which is much used
for small locomotives. In this arrangement, there are two ec-

VALVES

129

Gentries placed on the shaft at an angle of 180 2 a apart. The
forward eccentric controls the forward motion, and the back-
ward eccentric controls the reverse motion. In the diagram of
Fig. 88, the crank is shown on head-end dead center, and the
valve is indirect, so that the eccentrics will be at an angle of 90 a

FIG. 88

from the crank. In Fig. 88, FE is the forward eccentric, and
BE the backward eccentric. The two eccentric rods connect to
the eyes of a link at H and I. This link may be raised or low-
ered by the bell-crank BAD and the link DJ. When the link
is down, as shown in Fig. 89, the forward eccentric entirely con-
trols the motion of the valve. With the link all the way up, the
backward eccentric controls the valve. In Fig. 88, the link is

FIG. 89

shown in mid-position with both eccentrics controlling an equal
amount. With the crank on head-end dead center as in Fig. 88,
the valve is as far to the left as it can get, the port is open a dis-
tance equal to the lead, and the valve is at a distance equal to
the steam lap plus the lead from its mid-position. With the

130

ENGINES AND BOILERS

valve opening only lead distance, the engine will not get enough
steam to run. At mid-gear, half the travel of the valve is equal
to the steam lap plus the lead.

With the link in some position between those shown in Figs. 88
and 89, both eccentrics will control the motion of the valve, but
the forward one will predominate, and the engine will run for-
ward. The cut-off will now be earlier than it would be if the
link was all the way down in full gear. It is evident that the
Stephenson gear may be used to change the cut-off as well as to
reverse.

It is possible to give the same motion to the valve
with the link in intermediate position, by a simple equivalent
eccentric. The method of determining this equivalent eccentric
is not difficult but it will not be discussed here.

127. The Walschaert Valve Gear. Most large locomotives
in this country are equipped with the Walschaert gear, or some

FIG. 90

similar reversing gear. In this type of gear, shown in Fig. 90,
the motion of the valve is derived partly from the cross-head,
and partly from a return crank or eccentric. That part of the
motion coming from the cross-head is constant, while that de-
rived from the eccentric is varied for different conditions of load
and direction of rotation. Since the eccentric is placed outside
the driver, it is commonly called a return crank. The bar CE
is fastened to the outer end of the crank pin C, and the eccentric
pin E moves in the dotted circle (Fig. 90), about as a center.

VALVES 131

The angle between the crank and the eccentric is 90. The hori-
zontal motion of the eccentric is transmitted through the eccen-
tric rod EM to the lower point of a link. The link is pivoted
at its center G to the frame of the engine, so that the point M
oscillates about the point G. In this link is fitted a block which
can be raised or lowered in the link by the bell crank DAB.
As shown in the dotted position, the block is at its lowest posi-
tion in the link, and the engine is running forward, taking steam
during the largest possible part of stroke. In the full line posi-
tion, the block is at the center of the link, and the engine will
not get enough steam to drive it. With the block in mid-position
in the link, the eccentric will give no motion whatever to the
valve. Under this condition the motion of the valve comes
entirely from the cross-head.

The lever HIJ is called the combination lever, because it com-
bines the motion from the eccentric with the motion from the
cross-head. The ratio I H/JH is fixed by the condition that

I H __ 2(steam lap plus lead)
J H length of stroke

With the block in the center of the link at G, H has no horizontal
motion, but the horizontal motion of / is equal to

length of stroke X I H n , , , , , ,.

T-PT - = 2 (steam lap plus lead).

J ti

As the motion of the valve at mid-gear is 2 (steam lap plus lead),
it is seen that the valve will open only a distance equal to the
lead on each end, and the engine will not get enough steam to
run. When the block is dropped to the dotted position, the point
H does have a horizontal motion which comes from the eccen-
tric, and with the engine running in the direction shown, the
horizontal motion of H will add to the port opening. To reverse,
the block is raised above the center of the link.

By changing the position of the block in the link we may get
not only a reversal of direction of rotation but also a change in
cut-off. As in the Stephenson gear, it is possible to find an equiv-
alent eccentric which would give the motion actually obtained
from the mechanism.

132 ENGINES AND BOILERS

128. The Joy Valve Gear. The Joy gear is a so-called radial
gear. Unlike those previously described, it has no eccentric.
Figure 91 shows diagrammatically the principle of its operation.
A point such as D on the connecting rod FC will move in the
path of an ellipse, which is shown dotted. A bar BED is pinned
to the connecting rod at D. The other end of the bar is con-
nected to the frame of the engine by the link AB, A being a point
on the engine frame. It is evident that a point E on this
bar will have a combination of the elliptic motion of D
and the nearly vertical motion of B. The path of E is shown
dotted. A bar EGH is connected to BED by the pin E. The
point G is in a block which is at liberty to slide along a curved

FIG. 91

link. At any particular cut-off this link is held stationary, and
the block G slides up and down in it. It is thus seen that the
point H gets a combination of the motions of the point E and
of the point G. The horizontal component of the motion of H
is transmitted through the link I H to the valve. By rotating
the link about its center J, a reversal in the direction of rotation
of the shaft may be obtained. As in the other gears, cut-off
may be varied as well as the engine reversed.

Under conditions of light load an early cut-off may be used
and sufficient power obtained if the steam is not throttled between
the boiler and the engine. It is the practice on locomotive engines
to vary the cut-off to suit the load under normal running condi-
tions. The power generated by the engine might also be regu-
lated by throttling the steam, but it has been found that a higher
efficiency is obtained at light loads by using an early cut-off

VALVES 133

than by using a late cut-off and a low steam pressure, especially
where valve gear is used which increases compression when cut-
off is shortened, as in the Stephenson gear. In the reversing
gears commonly used, an early cut-off is accompanied by an early
compression. The increased efficiency at light loads is due, how-
ever, more to the early cut-off than to the high compression,
although the high compression aids by heating the clearance
space, piston, and cylinder head, thereby keeping initial conden-
sation within more reasonable limits.

129. Setting the Slide-valve. On a small engine that can
be turned over easily by hand, the setting of a slide-valve is a
simple matter. With proper valve setting an engine should run
smoothly, should be easy to start, and each end of the cylinder
should furnish about half the power. If an indicator is at hand,
it should be employed in the setting. If no indicator is available,
the valve may be set by linear measurement. In the setting of
a slide-valve there are but two things to do, shift the eccentric on
the shaft, and lengthen or shorten the valve stem or rod.

VALVE SETTING BY INDICATOR. In order to get approximately
the same amount of work from each end of the cylinder, cut-off
for the two ends should be about the same. A rough adjust-
ment may be made easily by placing the eccentric somewhere
near its proper position, and adjusting the position of the valve
on the rod so that the engine may be started. After the engine
is started, take cards and then adjust the length of the valve
stem until the cut-off is the same percent on both ends. Next,
shift the eccentric until the desired percentage of cut-off is attained.
Shifting the eccentric ahead makes cut-off come earlier. By the
use of the indicator it is easy to get the exact setting desired.

Knowing the valve-travel and the dimensions of the valve, we
may compute by means of the Zeuner valve diagram the exact
amount the valve stem must be lengthened or shortened and the
angle the eccentric must be shifted to give a desired setting.
In the upper part of Fig. 92 are shown two cards taken with a
valve as now set. It is desired to change the setting so that
cut-off will be 50 per cent for each end. The cards are shown
in length equal to the valve-travel, but they need not have been,
as the percentages of stroke could have been scaled and the
corresponding crank positions found.

134

ENGINES AND BOILERS

From the cards, locate on the crank circle (assumed for con-
venience with its diameter the same as that of the valve-travel
circle), the crank positions at the different events. Draw lines
between the crank positions at admission and cut-off, and from
release to compression, for each end. The distance between the
admission-cut-off lines should be the sum of the steam laps as
measured on the valve itself. A radial line drawn perpendicular
to the line joining admission and cut-off establishes the angle of
advance a\. The laps may be measured from the diagram as
shown. Now construct a Zeuner diagram (Fig. 93) for the desired

c.e.ad/n

FIG. 92

FIG. 93

cut-off (50 per cent in the diagram) keeping the sum of the steam laps
the same as before. This amount, and also the sum of the steam
lap and the exhaust lap for each end, will be the same no matter
what adjustment is made. The angle of advance for the new
condition is 0:2. Then <* 2 a.\ is the angle that the eccentric must
be shifted forward. The difference X between the new and the
old head-end steam laps is the distance the valve must be moved
on the rod. Since the new head-end steam lap is larger than the
old, the valve rod must be lengthened if the valve is direct, and
shortened if the valve is indirect. The upper part of Fig. 93 shows
the cards that may be expected after the setting has been changed.

VALVES 135

VALVE SETTING BY MEASUREMENT. If an indicator is not avail-
able, the valve may be set by linear measurement. The steam chest
cover must be removed so that the measurements may be taken.
It is customary to set either for equal cut-offs or for equal leads.
We cannot have equal leads and equal cut-offs at the same set-
ting because of the angularity of the connecting rod.* In either
case adjust the valve on the stem so that the valve travels about
as far beyond the head-end port as beyond the crank-end port.

SETTING FOR EQUAL CUT-OFF. By means of marks on the guides
and on the cross-head the stroke may be determined, and that
proportion from each end at which cut-off is to take place may
be laid off on the guides. The procedure is as follows:

(1) Place the cross-head at the position for head-end cut-off.

(2) Loosen the eccentric on the shaft and turn it on the shaft in
the direction the engine is to run until the valve is just on the point
of cutting off. Fasten the eccentric to the shaft in this position.

(3) Turn the engine to the position for cut-off at the crank-end
and measure the distance from the valve to its correct position
to give cut-off on this end. Divide the error by two and take
up half the error by shifting the valve on the stem and the other
half by turning the eccentric on the shaft.

(4) Turn the engine back to the position for head-end cut-off
and check the setting. If there is an error left, repeat as ex-
plained above.

Be sure the eccentric is fastened firmly to the shaft and the
valve to the stem. Replace the cover of the steam chest.
SETTING FOR EQUAL LEAD.

(1) Place the engine accurately on head-end dead center.

(2) Loosen the eccentric on the shaft and turn it in the direc-
tion the engine is to run until the port is open a distance equal to
the desired lead. Be sure the valve will open the port if the eccen-
tric is turned ahead more. Fasten the eccentric to the shaft.

(3) Turn the engine to crank-end dead center and measure the
error. Divide the error by two and take up half by turning the
eccentric on the shaft and the other half by adjusting the length
of the valve stem.

(4) Turn the engine back to head-end dead center and check.
If there is an error, correct it by repeating as previously explained.

* This expression means the deviation from parallelism with the axis of the cylinder, of a
connecting rod of finite practical length, except when the crank is at one of the dead centers.

136

ENGINES AND BOILERS

The reason that half the error is taken up by moving the valve
on the stem and half by turning the eccentric on the shaft may
be explained as follows. Suppose the valve has been set to give the
correct cut-off on the crank-end, but that when turned over to
head-end position for cut-off there is an error as shown in Fig. 94.

The eccentric is now at E\.
By turning the eccentric from
Ei back to E 2) this error
would be adjusted, but nothing
would have been gained as will
be seen when the engine is
turned back to the position for
crank-end cut-off. The same
error now exists on crank-end as shown in Fig. 95. That is, the
eccentric is at E* and should be at E%. If we try to take up the
error by lengthening the valve stem (Fig. 96), nothing is gained
because the valve will be moved to the left a distance equal to the
error, and when it is turned back to the position for crank-end cut-
off, the valve will be open a distance equal to the error. If now
we divide the error by two, and move the edge of the valve to A,

FIG. 94

FIG. 95

FIG. 96

in Fig. 94, by turning the eccentric back from EI to D, we will
find that the other edge of the valve will be at B in Fig. 95, when
turned over to the crank-end cut-off position. If the other half
of the error be taken up in Fig. 94 by lengthening the valve stem
i.e. by moving the valve to the left on its stem, the valve will be
in the correct position for head-end cut-off. Moreover, moving
the valve to the left moves the crank-end edge (Fig. 95), from B to
the edge of the port, where it should be for crank-end cut-off.
Therefore, by taking up half the error in each way at the posi-
tion of head-end cut-off, we have made no net change in the
position of the valve for the crank-end cut-off position.

CHAPTER IX
GOVERNORS

130. General. The function of the governor is to keep the
engine running at nearly the same speed at all loads. It does
this by controlling the amount of steam admitted to the cylinder.
It is impracticable for the governor to keep the speed exactly
constant at all loads. This may be seen when we understand
how a governor must work. Under conditions of changing load,
the governor must change the amount of steam admitted to do
the work. At any instant we have a certain load on the engine,
the governor is admitting enough steam to carry that load. If
an extra load is thrown on, the engine will momentarily slow
down. This slower speed affects the position of the parts of the
governor, and this in turn allows more steam to enter to do the
extra work required. This change cannot be affected instantane-
ously, although some governors respond in a very short time.
Governors that are quick in making the change are said to be
sensitive, and those that are slow, sluggish.

Under full-load conditions, the engine usually runs slower than
that at no load, by an amount that depends upon the construction
of the governor. Governors are made that give an engine speed
which is nearly as great at full load as at no load. These are said
to give doss regulation. It is possible to design and construct
a governor that gives the same average engine speed at all loads,
in which case the governor is said to be isochronous. Such a
governor would tend to hunt, i.e. there would be a constant
fluctuation in speed as the governor attempted to regulate the
steam supply to balance the load conditions.

Practical considerations limit the nearness to which isochronism
may be approached. If the no-load speed is greater than the
full-load speed the governor is said to be stable. If the governor
is isochronous or gives a full-load speed greater than no-load
speed, it is unstable. An unstable governor is clearly undesir-
able. Various schemes are used to express the relation of speeds
at various loads. A common way is to express the variation of
speed from no load to full load and from no load or full load to
normal load in percent of the speed at normal load. We may

137

138 ENGINES AND BOILERS

then say, the percent of variation in speed from no load to full
load is equal to 100(ni n^/n, where HI and n% are the speed at no
load and the speed at full load, respectively, and n is the speed
at normal load.

131. Classification of Governors. Governors may be classi-
fied according to the following characteristics.

(a) As to the manner of regulating the steam supply: Under
this head we have (1) throttling governors, which regulate the
amount of steam admitted to the cylinder by controlling a throttle
valve, and (2) cut-off governors, which control the steam supplied
by changing the point of cut-off.

(b) As to the predominant controlling force in the mechanism:
We speak of centrifugal governors, inertia governors (although
inertia is not a force), and resistance governors. All mass has
inertia. If the mass of the moving parts is small and the inertia
effect is not used in governing, we call the governor a centrifugal
governor. If the inertia is large and its effect is used to aid in
governing, we have what we call inertia governors. Even in inertia
governors, the centrifugal force is a very important factor.

(c) As to the force used to balance the centrifugal force of the
rotating parts*: We have gravity-balanced governors and spring-
balanced governors.

(d) As to the arrangement of the mechanism: There are
spindle governors and shaft governors.

132. The Gravity-balanced Spindle Governor. The diagram
of Fig. 97 represents a simple gravity-balanced spindle governor.
This is sometimes called a conical pendulum, and is also often
called the Watt governor, because James Watt first used it on
his engines. Two flyballs, at the ends of arms, rotate about a
vertical spindle. The arms are pivoted to the spindle at 0. The
height of the balls is caused to control the steam supply. In
Fig. 97 this is done by raising or lowering the point A with the balls,
which, by a suitable mechanism, causes either the movement of
a throttle valve or a change in the point of cut-off.

A definite relation exists between the height hi of the cone of
revolution, and the speed of the spindle. To determine this rela-
tion consider one of the balls as a free body. At any certain
speed it may be considered to be in equilibrium under the action
of the following forces: the tension in the arm T, the weight W,

GOVERNORS

139

and the centrifugal force acting radially outward (W/g)X(v z /R).
Taking moments of these forces about the point 0, we have

X^Xh-WXR = 0.
g ti

But v = 2irRn, where n is the number of revolutions per unit time.
Therefore we may write

1 =WR

If we wish to express hi in inches and the speed of the governor
in r. p. m., our equation becomes

32.2X12X3600 35200
- . 9 9 zj

4?r 2 n 2 n 2

approximately.

From this equation it is seen that the height hi of the cone of rev-
olution does not depend upon the length of the arm. Figure 97

FIG. 97

shows the r. p. m. corresponding to the height /ii for two-inch
increments of /i, up to 20 inches. From these values it will be
noticed that for a certain vertical movement of the ball there will
be less speed variation with the ball in the lower positions. In
other words, to get a reasonable speed variation it will be neces-
sary to run the governor very slowly. At low speeds the gov-
ernor will not have much power unless the balls are made exces-
sively heavy. This practical limitation precludes the use of this
governor on modern engines.

140

ENGINES AND BOILERS

FIG. 98

If the governor arms are crossed, as shown in Fig. 98, it will
be noticed that much less variation in speed exists for the same
vertical movement of the balls than for the form shown in Fig. 97.
Moreover, this governor is nearly isochronous at 50 r. p. m.

By the proper selec-
tion of the pivots B
and C, we may get
quite satisfactory
speed regulation for a
certain limited range
of vertical movement
at any desired speed.

The pendulum gov-
ernor may be made
exactly isochronous by
making the balls swing up in the arc of a parabola, as in Fig. 99.
The subnormal of a parabola is constant, and it is seen in Fig. 99
that hi is the sub-normal of the parabola which is the path of
the balls as they swing upward. The balls may be made to take
the parabolic path by having the arms made flexible and to unwind
from the evolute of the parabola or by having them guided by
an arrangement of
cams. Of course it is
understood that, in
practice, the governor
never would be made
exactly isochronous,
but it is seen that iso-
chronism may be ap-
proached as nearly as
practical conditions
will permit.

In order to run a
spindle governor at
fairly high speed, and FlG> 99

still have a reasonably

small speed variation in the engine, it is customary to load it as
shown in Fig. 100. The load L tends to pull the balls down;
hence they must rotate faster, to get to the same height as
before. To find the relation between the height of the cone of

GOVERNORS

141

revolution and the speed, consider the load and the ball each as
a free body. With the forces acting on them as shown, we can
express the conditions of equilibrium as follows. Expressing the
fact that the sum of the vertical forces is zero for the load L,

(1)
or

2T 2 sin j8 = L,

T --

2sin/3

Considering the ball as a
free body, and taking
moments about as a
center, we have, since the
sum of these moments must
be zero,

W v 2
(2) -p

Q rt

-WXR- !F 2 sin
By (1) this may be written in the form
(3) ~^ Xhi ~ WR ~ f X# - | ctn

or, since v = 2irnR,

W T T

(4) j WRtfh, -WR-^R-^ctu

whence, since ctn /3 = R/h 2 ,

TT n l -
and finally, solving for n 2 ,

r^+K^+A.)] , 1 +

(6) * = i-

If /Z-i = ^2

= 0,

L1 )J

/Tf +ZA
V IT /

If hi be expressed in inches and n in r. p. m.,

W + L\ 35200

142

ENGINES AND BOILERS

The usual form of loaded governor is shown in Fig. 101. This
is seen to vary slightly from Fig. 100. Taking the load as a free
body, T 2 = L/(2 sin /3), as in (1) above.

FIG. 101

Taking the upper arm as a free body, and expressing the fact
that the sum of the moments about equal to 0,

(9) Xj^XH -WXR-

(10) -X^XH-

g R

in PX^R- T 2 cos ftXhi = 0.
Substituting the value of T 2 from (1), we have

6 2 ^b 2

But we have

(ID

hence

(12) ~X^H -

Since we have also v = 2irRn,

hi = H T and
o

_ w __ _

2 b 2/i

whence

GOVERNORS 143

Taking as our units inches and r. p. m., this becomes

(15) n 2 =

In the solution of a problem for this type of governor it is best
to make a drawing and to scale from it the values of H, h 1} and
hz at the different positions of the weights. With these values
substituted in the formula (15), the r. p. m. of the governor is
readily determined. The governor of Fig. 101 is the one commonly
used on Corliss engines.

133. The Spring-balanced Governor. In most high-speed
engines, the centrifugal force of the revolving weight is balanced

FIG. 102

by the force of a spring. Figure 102 shows a weight W, revolv-
ing about the center of a spindle or shaft. The centrifugal force
C of the weight is balanced by the spring tension S. The weight
is at a distance R from the center of rotation. Hence we have

For a certain value of n, C varies directly as R. In Fig. 103,
this variation is shown graphically. At speed n and with a
radius R, the centrifugal force is C. If R is doubled, C is doubled.
For other values of n, as HI and WQ, C will have different values
with the same radius R.

144

ENGINES AND BOILERS

In any kind of uniform spring the elongation, shortening, or
deflection is proportional to the force producing the deformation.
Figure 104 shows graphically the relation between the pull of the
spring and the elongation. If Figs. 103 and 104 be superimposed,
as in Fig. 105, we easily can see the relations that must exist if

of p0/h of
FIG. 103

FIG. 104

the spring pull equals the centrifugal force. For the three speeds
HI, n and n 2 , the spring pull equals the centrifugal force, as seen
at the points b, d, and /. That is, the spring pull will balance
the centrifugal force at the speed %when the elongation of the spring
is ei. The weight will then be revolving at a radius Ri=a+e,
where a is the distance the weight would be from the center

of rotation if there were zero
tension in the spring. At a
speed n, the forces will balance
when the elongation is e and
the radius is R=a-\-e. In like
manner, the forces balance at
the speed n 2 with an elongation
62 and radius R%. If the origin
A should be moved over to the
origin 0, i.e. the distance a be
made zero, it is seen that the
spring pull could balance the
centrifugal force at only one speed. This means that the gover-
nor would then be isochronous. If the origin A were moved to
the left of the origin 0, the governor would be unstable, because
the speed HI at no load would be less than the speed 712 at full
load. As stated before, an unstable or isochronous governor could
not be used in practice.

- Q .

FIG. 105

GOVERNORS 145

What is known as scale of spring is the force necessary to pro-
duce an elongation of one inch in the spring. If the spring pull
is Si at elongation e\, and $2 at elongation e 2 , the scale of the
spring is equal to

Suppose that we desire to find the scale of spring necessary for
a governor such as that shown in Fig. 102. Let us suppose that
the no-load speed HI and the full-load speed n 2 , and the corre-
sponding radii RI and R 2 , are known. First compute Ci and C 2
for the two speeds. Then, since the spring pull must equal the
centrifugal force at all loads, the scale of spring is seen to be

because R\ R^=ei e%.

In actual governors it is very seldom that the spring pull acts
in the same line as the centrifugal force. Figure 106 represents

FIG. 106

a more usual case. In the solution of this problem, moments will
be taken about the pivot point of the governor arm B. In the
full-line position, the moment of the centrifugal force equals the
moment of the spring pull about the point B. That is, CXh =
SXd. In like manner, CiXi=SiXf. The scale of spring equals
(S Si) divided by the elongation of the spring as the weight
goes from R to RI.

146 ENGINES AND BOILERS

134. Governing by Changing Position of Eccentric. Most
shaft governors regulate the steam supply by changing the per-
cent of cut-off. This is accomplished by changing the position
of the eccentric relative to the crank. In Fig. 107, a Zeuner
valve diagram is shown on which appear two positions of the
crank at cut-off. In the full-line construction, cut-off comes late
and the angle of advance is i, i.e. when the crank is on head-
end dead center, the eccentric will be at E\ (direct valve). By
shifting the eccentric forward to E^ the angle of advance is

FIG. 107 FIG. 108

changed to 0$. The dotted construction gives the position of
the crank with the new angle of advance. It is seen that the
cut-off comes earlier with the larger angle of advance. By turning
the eccentric on the shaft, the time of cut-off can be changed but
the value of the steam lap will be the same for all values of a,
because the only way to change the laps is to move the valve
on the stem.

Figure 108 shows the effect on cut-off of changing the valve
travel while keeping the angle of advance constant. The Zeuner
construction for a large valve-travel is shown in full line. Cut-
off is seen to come fairly late. With a smaller valve-travel, as
shown by the dotted construction, cut-off is seen to come earlier.
Hence it appears that late cut-off is obtained with large valve-
travel and earlier cut-off with small valve-travel.

A governor can regulate cut-off by changing either a or the
valve-travel. It may be so constructed that it will control by
making only the one change or it may change the valve-travel
and the angle of advance at the same time.

GOVERNORS

147

Changing the angle of advance affects the other events as well
as cut-off. When a is increased all events occur sooner. Thus
on one-valve engines that control the steam supply by shifting
the eccentric, it is found that a high compression accompanies
early cut-off, such as is characteristic of the Stephenson valve-gear.

135. Governing by Changing a. In Fig. 109, a governor is
shown that controls cut-off by turning the eccentric around the
shaft. The two weight arms are pivoted at the points G and F.

FIG. 109

Any rotation of these arms about their pivots causes the eccentric
to turn on the shaft. The weight arms are connected by the
links AC and BD to AB, which carries the eccentric. At light
loads the speed of the engine is greater than at heavy loads, and
the weights will be farther from the center of rotation. This
movement of the weights away from the center of rotation causes
the eccentric to turn in a clockwise direction relative to the
crank. It is evident that this increases a arid therefore makes
cut-off come earlier, as it should at light loads.

136. Governing by Changing both a and the Valve-travel. -

In Fig. 110 a governor is shown that changes both a and the
valve-travel at the same time. The pivot point on the flywheel
carries the governor arm. The eccentric is shown as a pin.
When the governor arm moves about the pivot point in a clock-
wise direction, it carries the center of the eccentric with it and

148

ENGINES AND BOILERS

makes a smaller and the valve-travel larger. It is known that
small a and large valve-travel both give late cut-off. Conversely,
a counter-clockwise movement of the arm about the pivot point
gives early cut-off because it makes a larger and the valve-travel
smaller. The centrifugal force acts through the center of rota-
tion and the center of gravity. Hence this force tends to give
the arm counter-clockwise movement about the pivot. At light
loads, and therefore higher speeds, the centrifugal force will be
greater than for heavy loads. This tends to give the arm counter-
clockwise rotation about the pivot and this makes cut-off come

FIG. 110

earlier for light loads, as it should. A great many other governors
besides the one shown in Fig. 110, change a and the valve-travel
in the same way.

137. Centrifugal and Inertia Governors. As has been previ-
ously stated, all governor weights have inertia. If the tendency
of the weight to keep moving at the same speed helps to effect
the change of position that causes the governing, the governor
will respond more quickly than it would if the inertia opposed
the change. In the gravity-balanced spindle governors that were
considered in 132, inertia acts against the rapid change of posi-
tion of the balls and so tends to make the governor sluggish.

In the governor of Fig. 110, the inertia of the arm assists in
the governing. If the load is suddenly thrown off the engine,

GOVERNORS 149

it will momentarily speed up. This means that the flywheel will
go ahead of the governor arm or rotate in a clockwise direction
relative to the arm. This swings the eccentric nearer the center,
which makes a large and the valve-travel small. Hence cut-off
occurs sooner, which tends to re-establish the conditions of equi-
librium. In Fig. 110 the arm is made very heavy, so that it will
have considerable inertia.

It must not be assumed that the inertia of the arm is the only
factor in the governing. The centrifugal force also plays its
part, as has been explained previously. The governor of Fig.
110, while it is very simple in construction, is at the same time
sensitive and quick in its action. It is widely used and is known
as the Rites inertia governor.

CHAPTER X
STEAM TURBINES

138. Introduction. The steam turbine of to-day is of as much
importance in the world of engineering as is the reciprocating
steam engine. Practically all large steam power plants which
produce electric current employ the turbine engine. The devel-
opment of the turbine has been remarkably rapid. However, it
would not be true to say that the turbine has crowded the recip-
rocating engine from the power-plant field. The fact is rather
that it has developed along new and different lines of use, and
now occupies a field that was never held by the reciprocating
engine, i.e. as the direct connected prime mover for high-speed
electric generating units.

The turbine and the alternating-current generator have devel-
oped together. Both the turbine and the alternating-current gen-
erator are well adapted to high-speed rotation. The cost of a slow-
speed electric generator is much higher than that of a high-speed
generator of the same capacity. Before the days of the turbine,
generators were nearly all of the direct-current type, which could
not run at very high speeds.

The electric system of power transmission is more economical
than the old belt system. Hence the turbine has replaced the
reciprocating engine in some manufacturing plants on account of
the development of systems of electric power transmission. On
land, large turbines are seldom used to drive anything but
electric generators.

139. History. The steam turbine is not a modern invention.
Hundreds of years ago people knew, as every child knows today,
that a pin-wheel would rotate when blown upon. There are rec-
ords of turbines built in the quite distant past. They were little
but toys, however, like pin-wheels, and of no practical importance.
The modern turbine dates from the years between 1880 and 1890.
During this period two types of turbines that have become of
great practical importance were developed.

DE LAVAL, the inventor of the cream separator, sought to drive
his separator by means of a turbine. After several experiments,
he perfected a type that was satisfactory for that purpose. The

150

STEAM TURBINES 151

same turbine, with improvements, has been used in large numbers
for driving centrifugal pumps, fans, and even small generators.

During practically the same time, C. A. PARSONS developed the
type of turbine that now bears his name. These two pioneers
were soon followed by other experimenters. Various forms of tur-
bines were developed. Some of these are still used. Others are
obsolete and are of interest only from an historical standpoint.

In the development of the turbine, there were two obstacles
that had to be overcome. The first of these was the lack of
knowledge of principles. The second was the need of better
mechanical means of manufacture. As will be shown later, the
velocity of the rotor in a turbine must be very high. This causes
large stresses, and makes necessary a very perfect balance. The
clearances between the rotors and the stationary parts must be
small to prevent undue leakage. This calls for an excellence of
design and construction that did not commonly exist for heavy
machines in the past. As in the development of any new machine,
satisfactory solutions of the problems grew out of the necessities,
so that the modern turbine is as reliable and dependable as any
piece of machinery in the power plant.

140. Fundamental Principles. Before making a study of the
common types of turbines now in use, we shall discuss the fun-
damental principles of the steam turbine. It is not our purpose
to give an exhaustive discussion, but only to present the principles
in their simplest form. The sketches of blades and nozzles are
not exactly correct in shape for the conditions assumed. They are
to be considered as only diagrammatic.

Steam under pressure contains a certain amount of usable heat.
The available amount depends upon the initial pressure, the
degree of superheat, and the pressure to which the steam may
be dropped. There is the same amount of available heat if the
initial and final conditions of the steam are the same, no matter
whether we are considering the reciprocating steam engine or the
steam turbine. The turbine, or the reciprocating engine, is effi-
cient, or is not, according as it uses a large or a small amount
of this available energy.

Since the turbine and the reciprocating engine both use the
same medium, it is not to be expected that one will be much
more efficient than the other. Both reciprocating engines and

152 ENGINES AND BOILERS

turbines may be made of about the same thermal efficiency. The
choice of type of engine depends upon other considerations than
efficiency.

The greatest loss in the reciprocating engine is due to the initial
condensation in the cylinder. Since the cylinder walls are made
of a heat-conducting material, they will never be as hot as the
incoming high-pressure steam, and they will be hotter than the
low-pressure steam leaving the cylinder. The relatively hot
steam coming to the cylinder strikes the cooler cylinder walls
and some condensation takes place, with a consequent shrinkage
in the volume. The condensed steam is mostly re-evaporated
before the steam leaves the cylinder, owing to an absorption of heat
from the then hotter cylinder walls.

The loss in the turbine is due to other causes, such as leakage,
friction, etc. The leakage occurs around the ends of the blades
or from stage to stage. The friction exists between the steam and
the parts of the turbine in the passage of steam through both
the stationary and the moving parts. There is also a windage
loss between the moving parts and the steam. This friction does
not cause a complete loss, because a part of the heat generated
may be used in later stages of the turbine.

141. Available Energy in Steam. In order to make clear the
nature of the available energy in steam, a concrete example will
be taken.

(1) Let us assume that the steam is dry saturated steam at a
pressure of 150 pounds gage (165 pounds absolute), and that it
is allowed to expand adiabatically to a pressure of 15 pounds
absolute.* The heat contents of a pound of dry saturated steam
at 165 pounds absolute is 1194 B.t.u. The heat contents of a
pound of steam at 15 pounds absolute, after expanding adiabati-
cally from 165 pounds, is 1019 B.t.u. The difference between
these values, which is the amount of heat available for doing
work, is 11941019 = 175 B.t.u. At 165 pounds pressure, dry
saturated steam occupies a volume of 2.75 cubic feet per pound.
At 15 pounds pressure, after the expansion just mentioned, the
quality is 87 per cent, and the volume is about 23 cubic feet.

In the reciprocating steam engine, this change in volume, work-
ing by its pressure, does work on the piston in forcing it forward.

* Adiabatic expansion is that in which no heat is added to the steam and none is extracted
except by the conversion of heat into work.

STEAM TURBINES 153

The velocity of the piston is immaterial. In the turbine, the
same change in volume takes place, but the steam is allowed to
acquire velocity in expanding. The energy of the steam due to
its velocity is imparted to the rotor of the turbine. The efficiency
of a perfect engine working on the Rankine cycle,* between the
pressures of 165 and 15 pounds absolute, is 175/(1194 181) =17
per cent. Since neither the reciprocating engine nor the turbine
is perfect, neither would have an efficiency as great as 17 per
cent when worked between the pressure limits named.

(2) Assume that the steam is allowed to expand adiabatically
from 165 pounds absolute to 1 pound absolute. The heat-drop
is 1194 871 = 323 B.t.u. and the efficiency on the Rankine cycle
is 323/(1194-70) = 28.6 per cent.

142. Velocity Due to Expansion. Let us next compute the
velocity of the steam if all the heat-drop goes to giving the steam
velocity.

(1) Suppose that the drop in pressure is from 165 to 15 pounds
absolute. Since one B.t.u. = 778 foot-pounds, the energy is 175 X
778 = 136,100 foot-pounds per pound of steam. The energy of
motion, or kinetic energy, is mv 2 /2 = (1/32) Xv 2 /2. This must be
equal to the value 136,100 foot-pounds just calculated. Hence

v 2 = 64X136,100 = 8,710,400,
v = 2950 ft./sec.

(2) If the drop in pressure is from 165 to 1 pound, we find, in a
similar manner,

K.E. = 778X323 = (1/32) Xv*/2,
whence

v = 4010 ft./sec.

In the steam turbine, the steam must be expanded, and the
velocity due to this expansion must be used by imparting its
kinetic energy to the rotor of the turbine. If this is done by
allowing the steam to expand in the stationary parts of the tur-
bine and imparting the velocity thus produced to the moving
parts, the turbine is said to be of the impulse type. If it is done

* To compare steam engines, the efficiencies based on the Rankine cycle are often used. The
efficiency of a steam engine operating on the Rankine cycle is given by the expression
(Qi Qi/Q\, where Q\ is the amount of heat required to make dry steam at boiler pressure from
water at the temperature of the exhaust, and Qz is the amount of heat rejected from the engine
minus the heat of the liquid at the temperature of the exhaust.

154 ENGINES AND BOILERS

by allowing the steam to expand in the moving parts, the unbal-
anced steam pressure reacting on the rotor, the turbine is
called a reaction turbine. In some turbines the steam expands
both in the stationary parts and in moving parts, and the turbine
is said to be of the impure reaction type.

143. Impulse and Reaction. In order to understand the prin-
ciples involved, consider the simplest cases involving the principles
of impulse and reaction in which the velocity is created and used.
It may be easier to think of the jet as a jet of water, for in that
case the fluid does not expand when the pressure on it is reduced.

Otherwise, the steam jet and the

""""*'* . e '^ * : i water jet follow the same laws of

impulse and reaction.
FlG jjj Suppose that water issues from

a nozzle, as in Fig. 111. The

nozzle is stationary, and the issuing jet has a velocity of v feet
per second. The unit mass m of water that will be considered is
that issuing from the nozzle in one second. A particle of water
in the jet will move v feet in one second. The kinetic energy
of this unit mass will be mv 2 /2.

(1) Consider the case in which the jet strikes a stationary flat
surface (Fig. 112). After striking the flat surface, the water flows
or splatters out to the sides at right angles to its former direction
of motion, that is it loses all its

velocity in the direction of the

jet. The force exerted by the | ^"," "' * ^ *

jet on the flat surface may be

measured by the force F neces-

sary to hold the flat surface

stationary. Since force = mass X change in velocity per second,

and since the time in which mass m emerges from the nozzle

and strikes the plate is one second, we have F = mv. The force

exists in the case of the stationary plate, but no work is done

because the plate does not move.

(2) Suppose that the flat surface moves with a velocity V (Fig.
112). Then the force F=mX(v-V). The quantity (v-V) is
the velocity of the jet relative to the flat surface. It is seen that
F is less than before, and will be zero if the velocity of the sur-
face is the same as the velocity of the jet. The work done in one

Force F

* t/t/. &

STEAM TURBINES 155

second by the jet on the plate equals F times the distance the
plate moves, or

W=FxV=m(v-V)V.

The velocity of the plate at which the work is a maximum may
be found by equating to zero the first derivative of the work
with respect to V. This gives

Hence the maximum work occurs when V =v/2, that is, when
the velocity of the plate is half that of the jet.

(3) If instead of striking a flat surface, the jet strikes a station-
ary curved surface, such that

the jet is turned completely
back on itself, or through an
angle of 180 (Fig. 113), the
force F exerted on the surface

is mX2v, which is twice that "*'' ^~* f

,, a , f PIG. 113

exerted on the nat surtace.

Since the curved surface is stationary, there is no work done.

(4) Suppose the curved surface (Fig. 113) is moving with a
velocity V. The velocity of the jet relative to the curved surface is
(vV). It follows that the absolute velocity of the jet leaving
the surface is (v V) V = (v 2 V ) . Consequently the change in
the velocity of the jet is v-\-(v 2V) = 2(v V), because the
direction of motion is completely reversed. As before, we have

F=mX2(v-V),
and the work done per second is

W=FV = 2m(v-V)V = 2m(vV-V*),
whence

For maximum work we must have

^=0

Hence

v=2V, or V=v/2.

That is to say, the curved surface should move at half the velocity
of the jet for the production of maximum work. If the latter condi-

156 ENGINES AND BOILERS

tion exists, the absolute velocity of the jet as it leaves the sur-
face is zero. That is, all of the velocity of the jet has been used.
(5) In Fig. 114 we have a tank that is free to move horizontally
upon a track. In one side of this tank is placed an orifice or
nozzle. The water issues from this nozzle due to the pressure of
the water from above. If the tank is stationary, the water leaves
the tank with an absolute velocity v. The force F, due to the

unbalanced pressure of the water
in the tank, tends to force the
tank to the left, but since the tank
| is held stationary, the force F

^: fe/.v-^ does no work. If the tank is
' allowed to move to the left with

.....,, ,,,JJj$,,,,,,^), f ,,,,,,,,,,, r a ve l c ity V, however, the work

p iG 114 done will be FV. The absolute

velocity of water leaving the tank

is (vV). The maximum work will be done when V =v, that is
when the escaping water has no absolute velocity.

In the first four cases considered, the jet impinged on a sur-
face. Work was done by the jet striking and moving the surface.
A turbine in which the pressure-drop occurs in a stationary nozzle
or part is said to be an impulse turbine (142) because the energy
is given to the moving parts by the impulse of the jet. In the
fifth case the drop in pressure occurred in the moving nozzle.
When this occurs in a turbine, it is said to be a reaction turbine
(142). A comparison of the above simple examples shows that
the velocity of the moving parts of a reaction turbine must be
nearly twice as great as that of the impulse type, other factors
being equal.

144. Bucket Shapes. In the common types of steam tur-
bines, buckets or blades are mounted on the periphery of a wheel
or rotor. The shape of these blades is something like that of
the curved surface considered in (3), 143. Of necessity the jet
cannot be completely turned through an angle of 180 as in
(3), 143, because the steam must have a velocity in the direc-
tion of the axis of rotation of the rotor in order to get to the
bucket and to leave it.

In Fig. 115, let Vi denote the velocity of the jet relative to the
bucket or blade at the point where the jet first strikes it, and let a

STEAM TURBINES

157

denote the angle it makes with the tangent to the rim of the rotor.
Let v% and 0, respectively, denote the velocity and the angle upon
leaving the bucket. We see that the component of the velocity
of the jet relative to the bucket in the direction of the tangent
is vi cos a and the component in the direction of the axis of the
rotor is v\ sin a. In like man-
ner, the relative velocity v^
has similar components v% cos /3
and t>2 sin /3. These compo-
nents Vi sin a and v% sin /3
must be large enough to get
the jet through the row of
buckets on the rotor, in order
that the following buckets
shall not interfere with the
flow.

If the relative velocity of FIG. 115

the jet is Vi and the angle that

it makes with the tangent is a, the absolute velocity v of the
jet makes a different angle 6 with the tangent (Fig. 116). If V
denotes the tangential velocity of the bucket, v is the resultant of
the two velocities Vi and V, and 6 is the angle that this resultant
v makes with the tangent.

In like manner, the resultant of v% and V at the exit is the abso-
lute velocity v' at the exit, and
it makes an angle <f> with the
tangent.

If we assume that the jet
strikes the bucket in the direc-
tion of the tangent to the rim
of the rotor, the preceding par-
agraph shows that an error will
be introduced. In order to
make the calculations as simple
as possible, however, we shall
assume that the jet does strike
tangentially, and we shall bear in mind that some error has
been introduced. The results previously derived for steam veloci-
ties for certain heat drops will now be applied to the problem
of the turbine.

FIG. 116

158 ENGINES AND BOILERS

145. The Single-stage Turbine. In a single-stage impulse
turbine, the steam is expanded in a stationary nozzle, and is
directed against the moving buckets or blades, which are mounted
on the rim of the rotor. In the single-stage reaction turbine, the
rotor itself carries the nozzles, and the steam expands in passing
through them.

If the expansion of the steam is from 165 to 15 pounds absolute
we have seen in (1), 142, that the steam or jet velocity is 2950
feet per second. For maximum work done, the bucket velocity
of the impulse turbine is half that of the jet velocity ( 143).
Hence the peripheral velocity of the rotor should be 2950/2 = 1475
feet per second. With a reaction turbine, the peripheral velocity
of the rotor should be the same as the jet velocity, or 2950 feet
per second.

If the rotor speed is assumed to be 3000 revolutions per minute,
or 50 revolutions per second, the diameter of the rotor should be

1475

-^r- =9.4 feet

50-7T

for an impulse wheel. This is obviously very much too large.
For a reaction wheel, the diameter should be 18.7 feet, which is
absurd. If a speed of 24,000 r. p. m. is assumed, the diameter
of the rotor for an impulse turbine should be

1475

or 14 inches. These values for the speed and the diameter of
the rotor are not far from those which are used in the DeLaval
single-stage turbine.

The preceding examples show what a very high peripheral veloc-
ity is necessary for a fair efficiency in a single-stage turbine.
With a vacuum, a much larger velocity should be used. Since
immense stresses are induced in the wheel by these high velocities,
it is readily seen that a single-stage reaction turbine is almost out
of the question. If such turbines were operated, their efficiency
would necessarily be very low. Hence they are not used.

In Fig. 117, a diagram of the single-stage impulse turbine is
shown. Steam enters the nozzle from the left, expanding as
it passes through. As the pressure drops, a high velocity is
imparted to the steam. The steam leaves the nozzle at low pres-
sure and at a high velocity. The steam now impinges upon the

STEAM TURBINES

159

buckets or blades of the rotor, imparting to the rotor its velocity,
and therefore its kinetic energy. Upon leaving the rotor, the
absolute velocity of the steam is quite low.

The graphs at the lower part of the diagram show the changes
in the pressure and in the velocity. The steam pressure is shown
by the full line, and the steam velocity by the dotted line. While
single-stage impulse turbines are widely used, they are never
made in large sizes.

The diagram of Fig. 118 represents a single-stage reaction tur-
bine. The steam passes from the left directly to the rotor. The
rotor carries blades so shaped that the spaces between them act
as nozzles. The steam expands in these spaces or nozzles. As
it expands, its pressure drops, and it reacts upon the blades. This

Single-Stage Impulse
FIG. 117

Single - Sfagre Reac tioh
FIG. 118

force of the steam on the blades causes the rotor to move and to
absorb the energy liberated by the expansion. It will be noticed
from the graphs that the velocity does not change much in pass-
ing through the rotor, but the pressure drops during the passage.
In order to decrease the peripheral velocity of the rotor, and
at the same time to expand and use all the velocity of the steam,
more than one set of rotor blades or buckets are employed. This
is called staging. The steam passes successively through the
sets of blades in each stage, giving up part of the energy to each set.

160

ENGINES AND BOILERS

Staging. In multi-stage impulse turbines, two methods
are in use. The first method is to expand the steam in one set
of stationary nozzles, and to take out part of the velocity in
each stage. This is known as velocity-staging.

The second method is to expand the steam partially in one set of
stationary nozzles, using up the velocity caused by this expan-
sion in one stage, then to expand the steam again in another set of
stationary nozzles, using the velocity thus generated in the
second stage, and so on. This scheme is called pressure-staging.
A combination of pressure-staging and velocity-staging is also
used, in which there are two or more velocity stages in each pres-
sure stage.

147. Multi-stage Impulse Type with Velocity-staging. The
diagram in Fig. 119 shows the velocity-stage impulse turbine.

- Velocit y - Stage /mpulse .

FIG. 119

The steam enters from the left and passes through the stationary
expanding nozzle, where the pressure drops and the velocity is
acquired in exactly the same manner as in the single-stage im-
pulse turbine of Fig. 117. The rotor in this case, however, has
much less velocity than the rotor shown in Fig. 117. Hence the
steam loses only a part of its velocity in passing through the first

STEAM TURBINES 161

set of buckets. Emerging from the first set of buckets, it passes
through a set of stationary blades or vanes which change the
direction of flow of the steam, but not its velocity. These sta-
tionary blades are necessary, because the steam has a large up-
ward component of velocity after leaving the rotating buckets
of the first stage; and since the velocity of the rotor is downward,
the direction of flow must be reversed so that the steam may
impinge on the second set of rotating buckets. In passing through
the second set of moving buckets, more of the steam velocity is
taken up by the rotor. The direction of flow is again changed
by the stator, and so on, till the steam finally emerges from the
turbine with its velocity practically all expended.

Suppose, for example, that the downward velocity of the steam
as it leaves the nozzle is 4000 feet per second, and that the bucket
velocity downward is 500 feet per second. As it leaves the first
set of buckets on the rotor, the steam will have an upward velocity
of 4000-2X500=3000 feet per second, the effect of friction
being neglected. In going through the first set of stator blades,
the direction of flow is reversed, but is unchanged in magnitude.
Upon leaving the second set of rotor buckets its velocity will be
upward, and its magnitude will be 3000-2X500=2000 feet per
second, and so on for the other two stages. Each set of moving
buckets takes out 1000 feet per second of its velocity, and it
emerges with no vertical component of velocity.

It was shown in 145 that the rotor of a single-stage turbine
has to have an absurdly large diameter unless it has a very high
speed or the efficiency is very low. The objection to such high
speed is that the turbine must have a reducing gear in order that
the power may be used. With a multi-stage turbine we can
choose the diameter and also the speed, and make the number of
stages such that all the velocity can be used.

Let us assume that the speed is 3000 r. p. m., and that the
diameter of the rotor is 3 feet. Then the peripheral velocity
must be (3000/60) X37r =471 feet per second. Neglecting the
effect of friction, each stage will absorb a steam velocity of twice
the bucket velocity, or 942 feet per second. If the steam expands
from 165 to 15 pounds absolute, the steam velocity is 2950 feet
per second. Hence the number of stages necessary will be
2950/942 =3.1, and three stages should be used. If the turbine is
condensing, and the pressure drops from 165 to 1 pound absolute,

162 ENGINES AND BOILERS

the steam velocity will be 4010 feet per second, the number of
stages 4010/942 =4.2, and four stages should be used.

In a velocity-stage turbine, the efficiency is very low after the
first two stages, principally because the jet is broken up by its
passage through the blades. As a result, more than two velocity
stages are seldom used. It must be remembered also that a very
high steam velocity produces a very great friction between the
steam and the surfaces of the blades, thereby causing a consid-
erable loss.

Impulse Type with Pressure-staging. If,

instead of expanding the steam completely in one nozzle, we ex-
pand it only a little in the first nozzle, then use its velocity, ex-
pand it some more in a second nozzle, and again use the velocity
generated, and so on, the process is called pressure- staging (146).
Figure 120 shows the method diagrammatically. In this diagram
there are five sets of nozzles and five pressure stages. The steam
enters from the left and passes through the stationary nozzle.
The pressure and velocity lines below show that the drop in pres-
sure is accompanied by an increase in velocity. The steam with
its acquired velocity impinges on the blades of the rotor. The
velocity is absorbed in the rotor. As the steam leaves this rotor
with low velocity, it is collected and led to a second stationary
nozzle in which the pressure is again dropped, and velocity is
acquired. The second rotor absorbs this velocity and the steam
passes on through the following stages, until its pressure and
velocity are practically all used up at its exit.

Making the same assumptions for speed and diameter of the
rotor as in the preceding type, let us compute the number of
stages necessary with the pressure-stage type. If the diameter
of the rotor is 3 feet and the speed is 3000 r. p. m., there is a
steam velocity of 942 feet per second to be absorbed per stage
(147). If a pound of steam loses a velocity of 942 feet per
second, it gives up

~XrX942 = 13900 foot-pounds

Z oZ

of kinetic energy. This is equivalent to 13,900/778 = 17.9 B.t.u.
For the non-condensing condition assumed in 147, there was a

^ .

STEAM TURBINES

163

heat-drop of 175 B.t.u. available in the whole turbine. The num-
ber of stages will then be

For the condensing conditions assumed in 147, the number of
stages will be

17.9"

In making a comparison of this type with that of 147, it
might seem at first sight that the velocity-staging were the better,
as the number of stages is so much smaller. While this is an
advantage, it is overbalanced by the fact that the pressure-stage
type is more efficient. This type is used very extensively in

Multi- Pressure -Stage Impulse.

FIG. 120

medium-sized turbines. It is often called the multi-cellular type
because each pressure stage is composed of a cell that is steam-
tight except for the openings through the inlet and outlet nozzles.
It is evident that it is necessary to keep each cell steam-tight in
order to prevent a leakage of steam from one stage to the next.
Where no difference in pressure exists, there is no tendency to
leak. In the velocity-stage type, the pressure was the same
throughout the whole turbine beyond the expanding nozzles, and
so there could be no leakage.

164

ENGINES AND BOILERS

149. Multi-stage Impulse Type with Combined Pressure-
staging and Velocity-staging. Very often a combination of
pressure-staging and velocity-staging is used, with the result that
some of the advantages of both types are utilized. The diagram
of Fig. 121 shows this arrangement. In the sketch three pressure
stages are shown, and each pressure stage has two velocity stages.
The drop in pressure occurs in the three stationary nozzles. After
expanding in the nozzle, the steam passes through the buckets

/*" Pressure

g & Pressure 5fe?e 3 & Pressure 5hj<?e

Multi- Pressure, Multi -Velocity Stage (Cvrf/sJ
FIG. 121

of the rotor wheel, where a part of the velocity is absorbed. It
then emerges from this rotor, and its direction of flow is reversed
in the stationary vanes, as in the velocity-stage impulse type.
It then passes through a second set of moving buckets, where
most of the remaining velocity is absorbed. The steam is now
collected and expanded some more in the next nozzle, and the
process of the first pressure stage is repeated.

By dropping the pressure in three successive nozzles, the velocity
generated is not nearly so great as it is in the velocity-stage type
of Fig. 119. This means that there will be less friction loss due
to excessively high steam velocities. Ordinarily two velocity

STEAM TURBINES 165

stages are used for each pressure stage, so that the drop in
efficiency due to a disturbance of the jet is not so great as in the
pure velocity-stage type.

Applying the above problem to this combination type, let us
determine the number of pressure stages required. Assuming, as
before, that the speed is 3000 r. p. m., and that the diameter is
3 feet, we find that the bucket velocity is 471 feet per second
(147). Since there are two velocity stages for each pressure
stage, it is seen that 2X2X471=1884 feet per second of steam
velocity per pressure stage can be absorbed. The heat-drop per
pound of steam that corresponds to this velocity is

yXX 1884 x l/778= 71.3 B.t.u.

With a pressure-drop from 165 to 1 pound absolute, in which
there are available 323 B.t.u. for doing work, there will be
323/71.3 = 4 pressure stages.

Comparing this with other types, it is seen that the number
of pressure stages is the same as the number of velocity stages
in the velocity-stage type. But as there are two rows of rotor
buckets in each pressure stage, there will be twice as many rotor
wheels as in the former type. With pressure-staging, there were
18 rows of rotor buckets, or more than twice as many as in the
mixed type. This combination type is more efficient than the
velocity-stage type, and at the same time it is more compact than
the pure pressure-stage type. Due to these distinct advantages,
the combination type is extensively used.

150. Multi-stage Reaction Type. In the pure reaction tur-
bine, all of the expansion of the steam occurs in the moving
parts. At the present time, no pure reaction turbines are used.
The so-called reaction turbine in use expands its steam both in
the stationary and in the moving parts. It therefore employs a
mixture of the impulse principle and the reaction principle. This
mixed type is shown diagrammatically in Fig. 122. The steam
enters from the left and passes through a set of stationary blades
in which there occurs some drop in pressure. In passing through
the next set of blades, which are moving, a further drop in pres-
sure occurs. In this first set of rotor blades, the velocity gen-
erated in the stationary blades, and also that produced in the
rotor blades is absorbed. Hence there is a drop in pressure

166

ENGINES AND BOILERS

throughout the whole length of the turbine. As the velocity is
used up as fast as it is produced, no very high steam velocity
exists at any time during its passage through the turbine. The
graphs show the drop in pressure and also the change in velocity
as the steam passes through the turbine. The exact ratio of
pressure-drop or velocity-change is not shown in the curves, as
the scales used in the curves on the small cut are only illustrative.

Multi 3tage Impure Reaction
FIG. 122

It has been shown previously ( 143) that the velocity of the
blades of a pure reaction turbine should be the same as that of
the steam relative to the blades, in order to obtain the highest
efficiency. This means that, for the conditions we assumed in
147, the steam velocity to be used up per stage should be 471.
This velocity corresponds to a kinetic energy per pound of steam
of

= 3470 foot-pounds.

This is equivalent to 3470/778 =4.46 B.t.u. per pound of steam,
which is one-fourth of the value that was obtained for the im-

STEAM TURBINES

167

pulse turbine. For a heat drop of 175 B.t.u. (165 to 15 pounds
absolute), it will require 175/4.46=39 stages. Under the con-
densing conditions assumed in 147, 323/4.46 = 72 stages
would be necessary. For the pressure-stage impulse type, the
values were 10 and 18. Since the reaction turbine under discus-
sion employs a combination of the impulse principle and the
reaction principle, the values for the number of stages may be
taken as a mean between the values for the pure impulse type
and pure reaction type, that is 25 for the non-condensing condi-
tion, and 45 for the condensing condition.

The preceding computations show how much greater length
the reaction turbine must have than the impulse turbine. The
disadvantage of great length is offset to some extent by the fact
that the loss due to friction is less in the reaction type, since the
steam velocities are low.

151. Summary. To recapitulate, assuming that the speed is
3000 r. p. m., and that the diameter of the rotor is 3 feet, the
necessary stages for each of the various types is shown in the
following table.

Number of rows of buckets or
blades on the rotor

Heat-drop of
175 B.t.u.

Heat-drop of
323 B.t.u.

Velocity-stage impulse

3
10
4

40
25

18
8
72
45

Pressure-stage impulse . .

Mixed pressure- and velocity-stage impulse .
Pure reaction
Impure reaction

152. Change of Area of Steam Passage Space. It is neces-
sary to design a turbine so that the area of opening for the passage
of steam gives the proper velocity at all times. As the pressure
drops, the volume of steam increases. Hence there must be a
very much larger area for the passage of steam in the later stages
than in the first stages. This increase is accomplished in various
ways. If the diameter of all the rotor wheels is kept the same,
the area may be increased by having only partial peripheral
admission of the steam in the early stages. That is, the nozzles
or stationary vanes may extend only part way around the perim-

168

ENGINES AND BOILERS

eter of the stator. In Fig. 123 a is the opening for the passage
of steam in the first stage of the turbine. In the next stage, the
opening extends farther around the stator, thus giving a large
area for the passage of steam, and so on till, in the later stages,
the opening extends entirely around the stator.
The area may be increased by having full
peripheral admission in all the stages, but
having the length of nozzles and blades or
buckets increase with each successive stage.
Figure 124 shows this scheme. Or if full
peripheral admission is given and the blades
and nozzles are kept the same length, the
area for the passage of steam may be in-
creased by increasing the diameter of each successive stage, as
shown in Fig. 125. ,Ih practice, combinations of these methods
of increasing the area for the passage of steam are used. This
increase in area to allow for the increase in the volume of steam
passing, must not be confused with the increase in area in a

FIG. 123

FIG. 124

FIG. 125

FIG. 126

velocity-stage turbine which is necessary to allow for the decrease
in steam velocity for the different velocity stages.

153. Leakage. The leakage of steam through an opening
depends upon the difference in pressure that exists on the two
sides of the opening, and upon the area of the opening. In most
steam turbines, there are necessarily differences in pressure, and
openings through which steam may escape. With the rotor of a
turbine moving at a high speed, it is not possible to have tight
joints at all places between the stationary and the moving parts,
for then the friction between these parts would create an even

STEAM TURBINES 169

greater loss. In design and construction the clearances are kept
as small as is consistent with economy and safety, but even then
some leakage is sure to occur.

In Fig. 126, a difference in pressure exists between the two
sides of the stator at c and d; hence there will be a leakage of
steam at a. If there is a difference in pressure between d and e,
a leakage occurs at 6, and so on for the various stages. In the
velocity-stage impulse turbine there is no difference in pressure
after the steam leaves the nozzles; hence there is no tendency
to leak at either a or b. In the pressure-stage impulse type, there
is a difference in pressure between c and d, and therefore leakage
at a; but there is none between d and e.

In the impure reaction type there is a difference in pressure
between c and d, and also between d and e; hence there is leakage at
both a and b. The leakage at a can be kept at a minimum by
making the opening there as small as possible. This opening at a
depends upon the closeness of the fit and upon the diameter of
the shaft. Carbon or other form of packing is sometimes used
here to give a close fit. In most reaction turbines, however, the
shaft is very large, in fact it is even a drum, and in that case the
area of the opening is quite large. Packing cannot well be used
with large diameters.

In the reaction turbine the difference in pressure is not so
great between c and e as in the impulse type. In the reaction
turbine a leakage occurs at 6; hence the clearance there is kept as
small as is consistent with safety and economy. A certain amount
of spillage occurs at b because the rotor throws some of the steam
out by centrifugal force and it escapes through g, even if there
is no difference in pressure between d and e.

154. Loss Due to Running at Partial Capacity. Every tur-
bine is designed to expand the steam from a certain initial pres-
sure to a certain back pressure. This does not mean that the
turbine will run only under the pressure conditions assumed in
its design, but it does mean that the efficiency will be low if
there is any great difference from the assumed conditions. For
instance, if a turbine is designed to run non-condensing, and is
operated condensing, it means that the efficiency will be consid-
erably less than it would be if the turbine had been designed to
run condensing. The reason for this is easy to understand.

170 ENGINES AND BOILERS

Every nozzle and each steam passage is designed to carry a cer-
tain volume of steam with a certain velocity. We have shown
that velocity is caused by drop in pressure, so that if the pressure-
drops are not as designed, the velocities are not such that the
efficiency will be a maximum.

When a turbine is designed for a certain load, and that load
is greatly increased or decreased, it is evident that the efficiency
will be decreased. In the design of the governor the aim is to
make this drop in efficiency as small as possible, while at the same
time maintaining uniform speed for all loads. It is also essential
that the governor be reliable and therefore not too complicated.
The simplest form of governor is a throttling governor, and many
of the smaller turbines are equipped with that kind. However, with
light loads, the throttling governor causes a large loss in efficiency
because the range in pressure between admission and exhaust
is so largely decreased.

On the larger machines some other means than throttling us-
ually is used. If the machine is of the impulse type, it has sta-
tionary nozzles, and the governor can be arranged to open the
proper number to take care of the load. If the machine had a
single stage, or if a proportional number of nozzles could be kept
open in the later stages of a multi-stage turbine, this would seem
to be an ideal arrangement. But the large machines never are
made single-stage and a governor to control all the nozzles in
all the stages would be very complicated.

In practice, as in the ordinary Curtis turbine, the governor
controls the nozzles of the first stage only. At light loads only
a few nozzles of the first stage are open, while at maximum load
all the first-stage nozzles are open. This arrangement assures
good economy in the first stage at light loads, but as all the later-
stage nozzles are open, there will be a loss there.

In the single-stage DeLaval turbine, a throttling governor is
used, but an effort is made to maintain the best economy by
having some of the nozzles controlled by hand-operated valves.
By this method, most of the nozzles can be shut off at small
load and opened up by hand for heavy load. The governor is
incapable of controlling the load entirely if manual control is
attempted.

In a reaction type of machine in which there are no stationary
nozzles, the preceding method of governor-control cannot be used.

STEAM TURBINES 171

The Westinghouse Company, on their Parsons type of turbine,
has attempted to secure good efficiency at light loads by having
the governor admit the steam in puffs, which is the plan intro-
duced by Parsons. That is, the steam is admitted at full pres-
sure for a short time and then entirely cut off. The interval
between puffs is practically constant, but the length of time the
full steam pressure is on during the puff is controlled by the
governor. In the later stages, the effect of this kind of governor
approximates that of a plain throttling governor. In other makes
of reaction turbine a throttling governor is commonly used.

Large overloads are often carried in the various types of tur-
bines by turning full steam pressure into a later stage. In this
way the machine is able to carry a large excess of load at a re-
duced efficiency. If an electric generator is attached to the tur-
bine, care must be exercised that it is not overloaded long
enough to get too hot. This method is used only in emergencies,
and then only for short periods of time.

155. Summary of Losses in the Steam Turbine. The losses
that occur in a turbine have been mentioned in 140, 147,
153, 154, and elsewhere. We shall now make a summary of the
more serious causes of loss.

FRICTION LOSSES. Losses due to friction occur as follows :

(1) Between the shaft and the bearings, and in the packing
rings where the turbine is made steam-tight. With proper design
and construction this loss is quite small.

(2) Between the steam and parts of the turbine. The steam
friction varies directly as the steam pressure, and as the square
of the velocity between the steam and the parts. It also varies
as the amount of surface in contact with the steam. Hence the
amount of surface in contact with the steam should be kept as
small as possible and the velocity should be kept as low as possible.
The steam-friction loss in the first stages may be partially reclaimed
in the later stages since the heat generated tends to raise the tem-
perature of the steam. The steam friction occurs as the steam
flows through the nozzles and blades.

(3) There is also a friction loss between the rotor discs and the
steam surrounding them, at d, e, /, and g in Fig. 126. This loss
is called windage; it may be reduced by having the rotor discs
smooth and polished.

172 ENGINES AND BOILERS

LEAKAGE LOSSES. Leakage occurs wherever a difference in pres-
sure exists on the two sides of an opening. It is thus seen that
there may be a leakage out of the casing or through the joints
between the pressure stages, or around the balance pistons of the
reaction turbine ( 160).

Another loss occurs under condensing conditions because air
leaks into the low-pressure parts of the turbine. This air tends to
lower the vacuum in the condenser, and may thereby cause a loss.

DISTURBANCE OF FLOW. A breaking up of the jet of steam, and
the consequent formation of eddies, causes a considerable loss in
some types of turbines. This is one of the causes of the low
efficiency of the velocity-stage impulse turbine ( 147).

LACK OF PROPER VELOCITY. It has been shown that the proper
relation must exist between the velocity of the buckets and that
of the jet to obtain the highest efficiency ( 143). If the velocity
of the blades or buckets is not correct, there will be a loss. Throt-
tling of steam by the governor will cause a loss. (See 154.)

EXIT VELOCITY. The turbine derives its energy by absorbing
the steam velocity. A high velocity of steam in the exhaust en-
tails a decided loss. The turbine should be designed and operated
to extract nearly all the steam velocity, and to leave only enough
in the exhaust to cause the steam to flow to the condenser.

156. Common Commercial Types. A great many makes of
turbines have been used in the past. Some of these were not
economical and have been replaced. Others have ceased to exist
for other reasons. At the present time there are a great many
different forms in use, but space will not permit us to consider
all of them. Some of the more common forms will be explained.

157. DeLaval Single-stage Steam Turbine. The DeLaval
single-stage turbine is the oldest of the types used at present.
It dates back to the years 1880-1890. DE LAVAL, the inventor of
the cream separator, sought to drive that device by means of a
direct-connected turbine. In his experiments he developed the
type that bears his name. Some minor changes have been made,
but its essential features remain the same. The DeLaval tur-
bine used in America is manufactured by an American company
that originally produced their machines under the DeLaval
patents. The same company now makes a multi-stage machine,
but it is not to be confused with the single-stage type.

STEAM TURBINES

173

The essential features of the DeLaval turbine are as follows:

(1) The expanding nozzle in which all the pressure-drop occurs.

(2) The rotor wheel which carries a single row of buckets.

(3) A slender, flexible shaft that carries the wheel and transmits
the power to the gears.

(4) A set of reduction gears which lowers the speed so that it
is usable.

One style of the DeLaval turbine is shown in Fig. 127. The

FIG. 127

bucket wheel R is mounted on a flexible pinion shaft which is
supported by the bearings B, B'. This small shaft also carries
the small pinion P which is supported by the bearings &', 6'.
The pinion meshes with the two large helical gears shown in
the figure. The shafts that carry the large gears are supported
by the bearings b, b. The shafts of the electric generators or pumps
to be driven by the turbine are coupled to the large-gear shafts.
The governor also is driven

from one of these shafts.

The expanding nozzle is
shown in Fig. 128. The steam
enters from the left and the

FIG. 128

area of the opening gradually increases to a size that produces the
proper pressure and velocity. The amount of flare given to
the nozzle is governed by the drop in pressure that is desired.

174 ENGINES AND BOILERS

The nozzle used if the turbine exhausts to the atmosphere is
called a non-condensing nozzle. If the turbine exhausts into a
condenser, the nozzle is called a condensing nozzle. The flare in the
non-condensing nozzle is less than that in the condensing nozzle.

One or more of the nozzles are open at all times; the rest are
opened or closed by hand, depending upon the amount of the
load. The governor is of the centrifugal type, and governs by
throttling the steam. The nozzles are placed in the casing par-
tition at N (Fig. 127). The steam chest is at $; the steam passes
from it through the nozzles, then through the row of buckets on
the rotor, and out into the exhaust space E. From E, the steam
is led to the condenser or to the atmosphere.

We saw in 145 that the bucket velocity of a single-stage tur-
bine must be very high in order to extract a reasonable part of
the kinetic energy of the steam. The high bucket-velocity causes
large stresses in the rotor. The rotor disc therefore is made of high
quality alloy steel and is very carefully designed and constructed.
A sufficient factor of safety exists at the rated speed, but any
large increase above this speed will cause failure in the wheel.
It is customary to make the rotor weakest at a point just inside
the rim where the buckets are attached, so that, if the breaking
speed is reached, only the rim of the rotor will tear loose. When
that happens, the speed will die down of itself because the buckets
are gone. If the shaft becomes sprung, the safety bearings around
the hub keep the main part of the rotor in place, so that no great
damage results. The steel casing around the rotor is strong
enough to keep any small fragments from breaking through; but if
the whole rotor should break in two, it is likely
that considerable damage would be done.

The buckets are drop-forged. They are at-
tached to the wheel as shown in Fig. 129.
Slots are machined into the perimeter of the
rotor and the buckets are forced into the slots
from the side. Should an individual bucket
become damaged, it may be removed and an-
FIG. 129 other put in its place.

At high speed, there is a tendency for the
rotor to vibrate because it is impossible to get the center of gravity
exactly in the center of rotation. There is a critical speed at which
the vibration is a maximum. At a speed above this critical value,

JM*!M of a-t

STEAM TURBINES 175

the shaft or the bearings yield slightly, and the center of gravity
of the rotor comes into the axis of rotation. The rotor then runs
smoothly again. For smooth running, the speed must be either
very much less, or considerably greater, than the critical value.
Making the shaft small tends to reduce the critical speed. The
shaft of this turbine is made very small so that it may give
easily, and so run smoothly at its desired speed, which is normally
above the critical speed. While the diameter of the shaft is small, it
is ample to carry the power at the high velocity at which it runs.

The rotor speed in the smallest sizes is as high as 30,000 revo-
lutions per minute. For a 300-horsepower turbine, the speed is
10,000 revolutions per minute. It will be seen that it is imprac-
ticable to run an electric generator or a pump at such speed,
and to utilize the power developed, a speed reduction must be
used. Helical gears are used for this reduction in order to insure
smooth, quiet running. In some models, one large gear is used;
in others there are two, as shown in Fig. 127. The latter neces-
sitates two generators or pumps. When the gears are new, the
loss of power in the reduction is small. This loss increases as
the wear increases. Reduction gears are now used in some other
turbines, though they were used originally only in the DeLaval
turbine.

Aside from its use in the cream separator, the single-stage De-
Laval turbine has been used in driving electric generators, cen-
trifugal pumps and blowers. It is not used in sizes above 500
horsepower. The size is limited because the buckets would have
to be unreasonably long, or else the diameter of the rotor would
have to be too large, in order to get enough nozzles to play on
the buckets,

158. The Multi-pressure-stage Impulse Turbine. We have
seen that there is a practical limit to the size of single-stage tur-
bines. Because it gives the designer greater liberty of choice
of speeds, velocities, and bucket lengths, the multi-stage impulse
turbine is quite common in small and medium sizes. Moreover,
the efficiency of this type is comparatively high in these sizes.
For these reasons, it is quite common in sizes up to 5000 horse-
power.

Since it has a number of pressure stages or cells, this kind of
turbine is sometimes called the multi- cellular type. It is also

176

ENGINES AND BOILERS

called the Rateau type, because RATEAU was the first to develop
it, but there is no essential difference between the Eateau tur-
bine and the numerous other multi-cellular turbines.

Figure 130 shows an Economy turbine, made by the Kerr Tur-
bine Company. As seen in the figure, there is a rotor shaft which
carries a number of rotor wheels or discs. Each wheel runs in a
pressure cell. The cells are separated by the heavy diaphragms
shown at a. The joint between the diaphragm and the shaft is
kept as nearly steam-tight as possible by making as close a fit
as is practicable under the circumstances. The buckets are fastened

FIG. 130

on the rim of the wheels in much the same manner as in the single-
stage DeLaval turbine. The shaft is supported by the bearings
B, B. Nozzles are located in the cell partitions as at N. The
number of nozzles increases from the first to the last stage, in
order to allow for the increased volume of the steam as the pressure
drops.

The steam enters the steam chest S, and passes through the
nozzle N t where it is partially expanded. Leaving the nozzle, it
passes through the first row of buckets 6, on the rotor, then
through the second set of nozzles, and so on till it arrives at the
exhaust E. Leakage of steam from the high-pressure end of the
turbine, and of air into the exhaust, is prevented by the stuffing-
boxes G, G.

The casing is lagged with a non-conducting material to prevent
the loss of heat by radiation. The bearings are equipped with

STEAM TURBINES 177

ring oilers. As in other impulse turbines, there is no difference
between the pressure on the two sides of the rotor. Hence there
is no tendency to leak steam over the ends of the buckets, and it
is unnecessary to make the clearances between the buckets and
the stationary elements small. This obviates the necessity of
careful adjustments, and the danger of contact due to any un-
equal expansion of the parts. The diaphragms which compose
the cell walls are made rigid so that they will not spring as a
consequence of the differences in pressure on their opposite sides.

The turbine is supplied with a throttling governor of the cen-
trifugal type, driven from a worm on the main shaft. In the
smaller sizes it is direct-acting, that is, the position of the gov-
ernor weights directly controls the opening of the throttle valve.
In the larger sizes, it has been found desirable to have a relay
arrangement by which the throttle valve is thrown by steam or
by hydraulic pressure, which in turn is controlled by the position
of the governor weights. This arrangement gives better speed
regulation and does not require so large a governor. Consider-
able force is required to operate the large throttle valves.

Machines of this type made by other companies closely re-
semble the one just mentioned in essential features. Designers
often increase the length of the blades or buckets with each succes-
sive stage, thus helping to allow for the increase in volume. The
diameter of rotor wheels is sometimes made larger in the later
stages for the same purpose. Various devices are used to keep the
leakage from stage to stage at a minimum.

Carbon packing rings are sometimes used. In other makes,
labyrinth joints are used. No large mechanical pressure is allow-
able at the contact points between stages. The result is that
there is often a considerable loss from leakage of steam around
the wheels, sometimes as high as 15 to 20 per cent.

Thrust bearings are also used to prevent end-play of the shaft,
which might cause the buckets to rub and to become injured.

The practice of cutting holes in the rotor disc to insure an equal-
ization of pressure on the two sides of the rotor is common. The
shaft should be made large enough so that the speed at which it
is to run is well below the critical speed, and then excessive vibra-
tion will not occur. It is necessary to use good workmanship and
make the rotor well balanced; otherwise, severe strains would be
induced at the high speeds at which the turbine is run.

178 ENGINES AND BOILERS

It is not our purpose to describe the details of construction of
all the various makes of turbines. It must be remembered that
each make has its own minor variations in construction, but the
preceding description holds in general for all makes of this type of
machine.

159. The Curtis Steam Turbine. The original patents for the
Curtis turbine were issued about 1895, and the General Electric
Company started production of this machine shortly afterwards.
For several years it was built in this country only by them. More
recently, however, the Curtis principle has been used largely in.
the high-pressure stages of some other machines. At first the
General Electric Company built the larger turbines with the axis
of the rotor vertical. The generator, which was direct-connected
to the rotor shaft, was placed on top of the turbine, and the con-
denser was placed directly beneath it. This was claimed to be an
ideal arrangement, and in some respects it was excellent, but all the
weight of the rotating parts of both the generator and the turbine
had to be supported by a step-bearing. With every precaution,
and the best of design, this step-bearing would sometimes fail, and
this failure often caused the buckets to strip. For very large,
fast-speed turbines, it was difficult to secure sufficient mechanical
rigidity for the bearing supports in vertical machines. The makers
therefore have come to prefer the horizontal type. A great many
of the vertical turbines are still in successful operation, however.
The following description is taken from General Electric Co.
Bulletin No. 4883.

Figure 131 shows diagrammatically the progress of the steam
in a Curtis turbine. Entering at A from the steam pipe, it
passes into the steam chest B, and then through one or more
open valves to the bowls C. The number of valves open de-
pends upon the load, and their action is controlled by the gov-
ernor. From the bowls C, the steam expands through diverg-
ing nozzles D, entering the first row of revolving buckets of
the first stage at E, then passing through the stationary buckets
F, which reverse its direction and redirect it against the second
revolving row G.

This constitutes the performance of the steam in one stage,
or pressure chamber. Having entered the first row of buckets
at E with relatively high velocity, it leaves the last row G

STEAM TURBINES

179

with a relatively low velocity, its energy between the limits
of inlet and discharge pressure having been extracted in passing
from C to H. It has, however, a large amount of unexpended
energy, since the expansion from C to E has covered only a part
of the available pressure-range. The expansion process is, there-
fore, repeated in a second stage.

The steam, having left the buckets G, and having had its ve-
locity greatly reduced, reaches a second series of bowls H,
opening upon a second series of nozzles /. Through these the

FIG. 131

steam expands again from the first-stage pressure to some lower
pressure, again acquiring relatively high velocity in its expan-
sion through these nozzles, leaving them at J and impinging
upon and passing through the moving and stationary buckets
K, L and M, precisely as in the first stage. Again the velocity
acquired in the nozzle is expended in passing through the mov-
ing and stationary buckets, and the steam leaves the second row
M with relatively low velocity.

This process is continued in most large turbines through
several stages. Curtis machines as now constructed have a

180 ENGINES AND BOILERS

single stage in the very small sizes, up to six or seven stages in
the larger ones.

Again referring to Fig. 131, it should be noticed particularly
that the pressure of the steam has not changed in its passage
from E to H, that is the pressure is practically constant at all
points in the stage. This fact leads to one of the principal
structural advantages of the Curtis type. . . . For, since the
pressure is uniform at all points, there is no tendency for the
steam to pass elsewhere than where directed by the nozzles,
i.e. through the buckets. Hence there is no necessity of main-
taining a close clearance between the ends of the revolving
buckets and the turbine casing. In practice free clearance is
provided from one to two inches. The reaction type, to se-
cure high economy, must be provided with a minimum clearance
at this point.

Also since the pressure on both sides of each wheel, i.e. at
E and H, Fig. 131, is the same, the wheel is in perfect equilib-
rium, there being no tendency for the steam to force the wheel
in an axial direction. As this is true of each wheel, the entire
rotor is in equilibrium, and there is practically no unbalanced
thrust.

As the steam expands from stage to stage, its volume rapidly
increases, and a greater area of steam passage must be pro-
vided. This is accomplished in two ways. First, by increas-
ing the height of the buckets and second, by increasing the
number and area of nozzles from stage to stage. Referring to
Fig. 131, it should be noted that the primary admission noz-
zles D actually extend around a small portion of the first stage
periphery; therefore only those buckets adjacent to the nozzles
at any instant are carrying active steam. This applies equally
to the stationary row and the second revolving row; in fact,
the stationary or intermediate buckets, as built, extend over a
small arc not much larger than the nozzle arc. In the second
stage, however, the nozzle arc becomes longer and wider, thus
permitting the flow of steam through a greater number of re-
volving buckets and necessitating a longer arc of stationary
buckets. Finally, when the low-pressure stages are reached,
the nozzles and stationary buckets extend all the way around
the circumference.

As previously mentioned, greater area for the steam flow is

STEAM TURBINES 181

also provided by increasing the bucket lengths. For example,

the first-stage, or high-pressure, buckets are generally less than

an inch long, while those in the low-pressure stages may be

eight or ten inches in length.

It will be noticed that the length of the second row of moving
buckets in each stage, G and M, is greater than that of the first
row. This is made so, not to allow for any expansion of the
steam, but to provide for a decrease or velocity of the same vol-
ume of steam.

Customarily the buckets of the Curtis turbine are dovetailed
into the rim of the rotor wheel, somewhat as shown in Fig. 131.
At intervals the dovetail channel in the rim of the rotor is open
for the insertion of buckets. These openings are afterwards filled
with a spacing blank, and closed up. After the buckets are as-
sembled a shroud ring is riveted to their outer ends. The func-
tion of this ring is partly to stiffen the complete row and to reduce
vibration, but more especially to assist in retaining the steam
flow in the bucket space. Centrifugal force tends to throw the
steam out to the end of the bucket.

The governor is of the centrifugal type and controls the steam
supply by opening and closing some of the nozzles of the first
stage. Those nozzles that are open, are wide open, and those
that are closed, are tight shut. This scheme is positive and
reliable. It gives close speed regulation, and high efficiency at
light loads. In addition to the governor, the machine is equipped
with an emergency stop, whose function is to prevent excessive
speeds, should the governor fail for any reason. It consists of
an unequally weighted ring attached to, and revolving with, the
shaft. At any speed up to the normal speed, the weights are held
concentric with the shaft by springs, but at excessive speeds the
force of the springs is overcome. Then the ring revolves eccen-
trically, and trips the valve mechanism, causing the main throttle
valve to close instantly, thereby shutting off the steam supply.

Figure 132 shows a marine Curtis turbine. In marine service,
the speed must be very much less than is common in land prac-
tice, if the rotor is coupled directly to the propeller shaft. To
get this low speed, it is necessary to use very large diameters or
a great number of stages. Usually both, schemes are combined.
As has been stated previously, the efficiency after the second
stage of a velocity-staged impulse turbine is low, but in order to

182

STEAM TURBINES 183

reduce the speed properly, it may be desirable to use more than
two velocity stages, even if the efficiency is reduced. In land
practice this concession is not often made. In the forward tur-
bine shown in Fig. 132, there are four velocity stages in the first
pressure stage, three velocity stages in each pressure stage from
the second to the sixth, and two velocity stages in each pressure
stage from the seventh to the fourteenth. In the reverse tur-
bine, there are four velocity stages in each of the pressure stages.

It is customary in marine turbines to have the reverse turbine
mounted on the same shaft with the forward or ahead turbine.
It is put on the exhaust end of the shaft so that when it runs
idle, the rotor is in a high vacuum and therefore offers as little
resistance as possible to rotation. Since the friction loss varies
as the steam pressure, the loss is not great for low pressures.

The efficiency during reverse operation is quite poor because
there are only two pressure stages in the reverse turbine. The
reverse is in use only briefly and rather unfrequently. Hence the
the low efficiency of the reverse turbine is a negligible factor.

After a study of the previous types, Fig. 132 should be largely
self-explanatory, hence no detailed description need be given.
It is to be noted that the buckets of the seventh to fourteenth
stages are carried by a drum. As the pressure on the right end
of this drum is greater than that on the left end, it is seen that
there will be an end-thrust of the shaft. This thrust is taken
up by the thrust bearing T. The thrust bearing T is for the
turbine only, and not for the propeller. On every propeller shaft
there is a thrust bearing to take up the thrust of the propeller.
The bearings are provided with a water jacket to prevent heating.
In all large turbines, the bearings are cooled either by means of
water or oil. When oil is used, it is often cooled by a device
similar to the surface condenser.

160. The Parsons Steam Turbine. The Parsons turbine is
not only one of the oldest, but also one of the most common types.
It is usually made in medium and large sizes. In small sizes the
Parsons turbine is expensive, and is not very efficient. All of the
previous types have operated on the impulse principle, but this
one uses a mixture of impulse and reaction. It is ordinarily
called a reaction turbine. There are no distinct nozzles, as in
the impulse turbine. Instead, there are alternate fixed and mov-

184

STEAM TURBINES 185

ing blades, and expansion occurs in both. It is made by the
Westinghouse Company and by the Allis-Ch aimers Company.

Figure 133 shows one style of Parsons Turbine. Instead of
rotor wheels as in the previous types, the shaft carries a drum
on which the moving blades are mounted. Steam enters at the
steam inlet and passes through the governor valve to the left
of the first set of blades. Passing through the alternate fixed
and moving blades, it leaves through the exhaust outlet.

Full peripheral admission is used. The increase in passage
area is accomplished by increasing the length of blades from
stage to stage. The first-stage blades are very short, usually less
than an inch in length. After a large number of stages, the blades
are of considerable length. To avoid further increase in blade
length, the diameter of the drum is increased, and the first blades
on the larger diameter are made smaller so as to give the proper
passage area. Progressing to the right the length of blades again
increases, and again the diameter is increased. In the last stage
the blades are quite long. In the larger machines, the final blades
may be as much as a foot long.

Where the drum size is increased, there is an area exposed to
steam pressure from the left. This pressure causes an end-thrust
on the rotor to the right. To balance this end-thrust dummy
or balance pistons are placed on the left end of the drum, each
one being made of such a diameter that the steam pressure on its
right side produces the proper force to the left. Initial steam
pressure acts on the right of the smallest piston PI. An equaliz-
ing passage, EI, leads from the left of the second stage on the
drum to the right side of the middle balance piston P 2 , so that the
same steam pressure exists at 6 as at c. In like manner, the pres-
sure at e is the same as at d, on the right side of the largest balance
piston PS. A third equalizing pipe, E$, connects the exhaust space
with the left side of the piston P 3 . Of course there is some leak-
age of steam by the balance pistons, but this is minimized by
cutting annular grooves in the pistons and having rings on the
casing extend into these grooves to form a labyrinth packing.

Since the pressure drops in both the fixed and movable blades,
a leakage takes place, both between the ends of the fixed blades
and the drum, and over the ends of the moving blades. It is
therefore essential that the radial clearance be made as small as
is safe. Even with the smallest clearance possible, there is bound

186 ENGINES AND BOILERS

to be some leakage. The proportion of radial clearance space to
the area of the steam passageway through the blades is propor-
tionally greater with the short blades than with the longer ones.
Hence more leakage occurs in the first stage than in the last.
It follows that there is better economy in a reaction turbine in
the low-pressure stages than in the high-pressure stages.

In some makes of Parsons turbines a shroud ring is fastened
over the ends of the blades. In others, the shroud is left off, but
the blades are lashed together by means of wires that pierce the
blades. This wire is comma shaped in cross section, and the tail
of the comma is caulked down on each side of the blades, thereby
keeping them in the same relative position. These shroud rings
or lashing wires do not add strength to resist the centrifugal force,
but they keep down the vibration or flutter of individual blades.
With long slender blades, the flutter might be more than the axial
clearance and the contact at high speed might cause the blades
to be torn loose. With only one blade loose, the whole system
of blading might be almost instantly torn out. With this type
of turbine, it is of the utmost importance that each blade be
properly secured and adjusted. With the previous types, the
stripping of blades may be localized to one stage, but in this,
the damage is apt to be more general.

The machine represented in Fig. 133 is equipped with an over-
load valve. If the load is more than the turbine is ordinarily
able to carry, this valve is opened, allowing high-pressure steam
to enter at the second step, as shown by the dotted arrows. The
steam consumption will be greatly increased, but a much larger
load may be carried. Since the governor still controls, the speed
may be considerably reduced. The overload valve is for emer-
gency operation only, and is not supposed to be used often.

The two joints between the shaft and the casing are made
tight by a water seal. As a turbine is ordinarily run condensing,
there is a tendency for air to leak in at these joints. This leak-
age of air into the turbine is likely to produce a greater loss of
efficiency than would a leakage of steam outward. The reason
for this is that the vacuum in the condenser is greatly impaired
by the air in the steam. If the seal is kept full of water, the
leakage inward will be of the water, which will have little or no
bad effect on the economy of the turbine. In some turbines low-
pressure steam is used in place of water with the same effect.

STEAM TURBINES 187

As the weight of the rotor is very great, the bearings B, B must
be well constructed and kept cool. As in the previous type, the
bearings are usually kept cool by means of a circulation of oil.
With small clearances, no end-play can be allowed. To keep
the rotor from moving axially, a thrust bearing is used. The ad-
justment of the thrust bearing is made by means of the two
screws shown in the figure.

The number of stages in a reaction turbine is very great, which
necessitates a long rotor. With a long rotor, the changes in
length due to temperature changes is considerable. This distor-
tion increases with the amount of superheat. Hence little super-
heat can be used with some long reaction turbines.

To use more superheat, and also to limit the length of blades
in the later stages, the designers sometimes resort to a scheme
called compounding, i.e., the turbine will be cut into two separate
parts. The steam passes first through the high-pressure turbine,
and then is led to the low-pressure turbine. If the high- and
low-pressure turbines are both placed on the same shaft, the
machine is called a tandem-compound turbine. Since they are
on the same shaft, they must both have the same angular speed.
Sometimes better results can be obtained by mounting each tur-
bine on its own shaft and running the two at different speeds.
With the latter arrangement, the machine is called a cross-com-
pound turbine. In marine service, cross-compound turbines are
sometimes used, and then each turbine is connected to its own
propeller shaft.

In order to get rid of the balance pistons, Parsons turbines are
sometimes made double-flow. In the double-flow turbine, the
steam enters at the center of the casing and half flows to the
right, while the other half flows to the left. The two halves of
the drum are exact duplicates and any end-thrust on one half
is balanced by the thrust on the other. While this adds to the
total number of blades in the turbine, it does away with the
dummy pistons. Figure 134 shows the double-flow arrangement,
and is self-explanatory. Quite often the low-pressure turbine in
the compound arrangement is made for double flow.

The governor used on the Parsons turbine made by the West-
inghouse Company is of the blast type. As mentioned in 154,
the steam is admitted in blasts or puffs. The speed of the gov-
ernor is much less than that of the shaft, since it is reduced by

188

ENGINES AND BOILERS

a worm gear from the main shaft. A diagram of this type of
governor is shown in Fig. 135. The rod C is given an up and

FIG. 134

down motion from an eccentric on the governor shaft. The pivots
D and E being fixed, the reciprocating motion is communicated
by means of the links and levers to the small pilot valve A. The

recess on the valve A
allows steam to enter
periodically through
the pipe, to pass
through the ports, and
to push up on the
under side of the pis-
ton B. The fit around
the rod running down
from B is loose, so
that the pressure soon
drops. The vertical

p IG 135 position of the point F

is controlled by the

position of the balls of the governor. This in turn controls the
length of time the pilot valve admits steam under the piston B.

STEAM TURBINES

189

The piston B is connected to the main steam valve of the turbine.
When B is down, the main steam valve is wide open. When it is
up, the steam is shut off. The length of the time it is up for each
puff is seen to depend upon the position of the governor weights.

161. The Westinghouse Turbine. Aside from the Parsons
turbine just described, the Westinghouse Company has made for
several years a turbine which they call the Westinghouse. The
same type is made under other names, and has become quite
popular abroad. It consists of one impulse stage, such as exists
in the Curtis turbine, i.e., one pressure stage, with two velocity
stages, and the remainder of the turbine of the Parsons type.
This Curtis stage may be used with a single-flow Parsons, or a
double-flow Parsons, or with single-flow intermediate Parsons
stages combined with double-flow Parsons in the final stages.

Figure 136 illustrates the Westinghouse type in which there is

FIG. 136

a Curtis stage combined with a double-flow Parsons. The steam
enters the inlet at A and passes through the first stage exactly
as in the Curtis turbine; it then divides, half going to the right
and half to the left, through the reaction stages, and comes out
to the exhaust at the two ends of the casing. Partial peripheral
admission is used in the Curtis stage and the governor controls
the number of nozzles in use, as in the Curtis type.

There are certain advantages to be gained by the mixed Westing-
house type. First, the length of the rotor is shortened, because

190 ENGINES AND BOILERS

the length taken up by the Curtis stage is only a small part of
that which would be required by the same pressure drop in the
reaction type. Second, the efficiency in the high-pressure part
of the turbine is increased. We have seen that there is much
leakage in the high-pressure stages of the Parsons type because
the blade lengths are short in these stages, and the leakage is
proportionately large. Third, it allows the governor to control
the steam supply by shutting off nozzles, which is superior to
either the throttling or blast governing.

Turbines are also in use that have one or two Curtis high-pres-
sure stages, and low-pressure stages of the Rateau type. This
combination is not common at present in this country, but it is
found in some turbines in Europe.

There are many factors that determine the choice of type of
a turbine. A satisfactory discussion is impossible here. Given
the size, the kind of service, and the various operating conditions,
the designer is able to say which type is the best.

162. Other Types. In addition to the types heretofore de-
scribed, there are various other small turbines in use. They are
principally of the Pelton type, i.e. they are built on the same
lines as a Pelton water wheel. Buckets are cut in the outer
rim of a rotor and the steam enters them in a direction nearly
tangential to the rotor. They are usually built in small sizes
and are used for driving fans, blowers, and pumps. In such
service simplicity counts for more than high efficiency.

163. Low-pressure Turbines. The reciprocating steam engine
utilizes economically a greater amount of the available energy
of the steam at high pressures than at low pressures. That is,
the efficiency of a reciprocating engine is relatively higher at a prac-
tical range of pressures above the atmosphere than below atmos-
phere. This does not mean that the non-condensing engine is
more efficient than the condensing, but that of two engines, one
taking steam at a high pressure and expanding to atmospheric
pressure, and the other taking steam at atmospheric pressure
and expanding down to that of a good vacuum, the former will
be the more efficient. There is about the same amount of energy
available for doing work in expanding from a medium boiler
pressure down to that of the atmosphere, as there is in expand-
ing from atmospheric pressure to that of a good vacuum. How-

STEAM TURBINES 191

ever, the amount of expansion is much larger in the second case. To
utilize this great amount of expansion at low pressure, the cylinder
would have to be very large; consequently a large cylinder loss
would occur. With the turbine, the reverse is true, i.e. the turbine
suffers relatively less loss at low pressures than at high pressures.
It has been shown by experience that the power out-put of a
plant using non-condensing engines can be increased between 80 and
100 per cent by taking steam from the engines and sending it
through what are called low-pressure turbines which exhaust into
a notably excellent vacuum. This power increase is obtained
from the steam without any extra coal cost or any increase in
the size of the boiler plant, through changing the exhaust pres-
sure from that of the atmosphere to that of the vacuum by the
introduction of the turbine and its condenser. It sometimes
happens that an existing plant, equipped with reciprocating en-
gines, is to be enlarged, and in some instances the low-pressure
turbine has been used to solve this problem. In case the existing
engines are already condensing, it has been found that a net gain
in power of 25 to 40 per cent may result by using their exhaust for
the turbines, on account of the more perfect vacuum used with
turbines since they are not subject to cylinder condensation and
since the air leakage is less. The low-pressure turbine may take
steam from engines, pumps, air compressors, hoists, etc.

164. Mixed-pressure Turbines. The mixed-pressure turbine
is much the same as the low-pressure turbine. It uses low-
pressure steam, but it may also use some high-pressure steam
at the same time. The high-pressure steam is admitted in vary-
ing amounts, to make up any deficiency in the supply of the low-
pressure steam. This may be done by throttling the high-pressure
steam down to the low-pressure before using it, or the turbine
may be equipped with high-pressure stages to use the steam
without throttling. The latter is the more efficient way of using
the high-pressure steam.

165. Bleeder Turbines. In some plants equipped with tur-
bines, a supply of low-pressure steam is required for heating or
for some manufacturing process. Rather than take high-pressure
steam and throttle it down to the required pressure, the low-
pressure steam may be drawn from the turbine at an intermediate
stage. Then the machine is said to be a bleeder turbine.

192 ENGINES AND BOILERS

166. The Use of Superheated Steam. Turbines are, in gen-
eral, well adapted to the use of superheated steam. Comparative
tests show a marked increase of efficiency when superheated steam
is used. It has been claimed that this is due partly to a lessening
of friction between the steam and the parts, but it is due mainly
to the increased available energy in the steam which enters the
turbine. Superheated steam also gives an increased efficiency
when used in the reciprocating engine, but its use there is accom-
panied by an added difficulty in lubrication. With the turbine,
lubrication is no more difficult with superheated steam than with
saturated steam, since the steam-wet metal moving parts are not
in rubbing contact with other metal parts. Superheating the
steam used for turbines reduces the wear on the blading, com-
pared with steam which is saturated or wet at the inlet.

167. The Marine Turbine. The turbine has been used in
marine service since the early years of its development. Many
factors make the turbine particularly well adapted to this service.
It occupies less space than a reciprocating engine, and it is lighter.
It gives a uniform torque on the propeller shaft and does not
cause as much vibration of the ship's hull as does the recipro-
cating engine. On the other hand, the turbine is a high-speed
machine, while for good efficiency the propeller must be run at
rather a low speed. Direct connection of the turbine to the pro-
peller shaft is, of course, the simplest arrangement. When this
is done, it is necessary to design the turbine to run as slowly as
possible, and to design the propeller to run as fast as possible.
Even then a compromise often has to be made: the propeller
has to run too fast, and the turbine too slow, for best efficiency.
Hence direct-connected turbines are limited to swift boats. One
way out of this difficulty is to reduce the speed by means of gears.
While this is sometimes done, it is not entirely satisfactory, since
the gears are not highly efficient when worn.

A method of drive that is being used to a considerable extent
is the electric drive. With this, turbines similar to those used
on land are used to drive electric generators, and the current
is used by motors direct-connected to the propeller shafts. This
allows both the turbine and the propeller to run at the proper
speed. It also permits great flexibility of arrangement, and con-
venience in steering and in maneuvering.

CHAPTER XI
GAS ENGINES

168. Introduction. The small-sized internal-combustion en-
gine is perhaps better known to the average person than any
other prime mover. During the past twenty years it has exerted
a very marked effect upon our manner of living. It has made
possible the automobile, the motor truck, the gasoline tractor,
and the airplane. It has replaced the more expensive, small
hand machines and horse machines on our farms. It is indeed
hard to estimate the value to mankind of the gasoline engine.

In plants in which there are combustible waste gases, such as
those from the blast furnaces and coke ovens, large-size gas
engine units have come into use, and they are there more eco-
nomical than steam engines. In marine service, where space is
a prime consideration, as in the submarine, the internal-combus-
tion engine is generally used.

169. History. Many years ago, men of an inventive turn
of mind dreamed of gunpowder engines or explosive engines, and
many patents were taken out for these devices. Records show
that as early as 1680 HUYGHENS produced a working model of a
gunpowder engine. It was of no practical importance. Many
other inventors produced various forms of engines, but not till
1860 was an internal-combustion engine produced commercially.
At that time LENOIR started building gas engines. In the course
of a few years four or five hundred of these engines were built,
but the engine was not very efficient as it lacked compression for
the unignited gases.

The first really scientific work done on the gas engine was that
of the French engineer, BEAU DE ROCHAS, who laid down the fol-
lowing four conditions as being essential to high efficiency.

(1) The largest cylinder volume, with smallest exposed surface,
i.e., the proper relation of diameter to length of stroke.

(2) The greatest possible rapidity of explosion, i.e., the maxi-
mum piston speed.

(3) Highest possible pressure at the beginning of the expansion.

(4) The greatest possible expansion of burnt gases.

The same engineer proposed to obtain the above results by
means of a single cylinder and to operate his engine upon the

following cycle.

193

194 ENGINES AND BOILERS

(1) Draw in a charge of mixed air and gas through an entire
stroke.

(2) Compress this mixture during the next stroke.

(3) Ignite the compressed combustible mixture at the beginning
of the third stroke, and expand the products of combustion dur-
ing the stroke.

(4) Discharge the burned gases on the following stroke.

The above is known as the four-stroke cycle, or the Otto cycle,
and is the one most commonly used. It will be considered more
in detail later.

A few years after Beau de Rochas secured his patent, two in-
ventors, OTTO and LANGEN, produced an engine that had a ver-
tical cylinder with a free piston. The explosion of uncompressed
gases drove this piston upward. On its downward stroke, a rack
attached to the piston engaged, through a clutch, with a spur gear
that drove the machinery. While this engine was more efficient
than any produced before, it was noisy. Although several thou-
sand were produced, the design was abandoned after a new design
using compression of the admitted gases was produced by Otto.

The new Otto engine was shown at the Paris Exhibition in
1878. It operated on the cycle formulated by Beau de Rochas,
and may be considered as the first modern gas engine.

A few years later, DUGALD CLERK brought out an engine with a
two-stroke cycle. Many machines of this type are in use, es-
pecially in motor boats and the like.

Mention should be made of the Brayton engine, which was
produced at about the same time as the Otto-Langen engine.
GEORGE B. BRAYTON was an American, and many of his engines
were used in this country. The principle of its operation is quite
different from that of the others mentioned, but it need not be
explained here.

The Diesel engine dates back to 1892, but it was not perfected
until some time later. Since then, the semi-Diesel and other oil-
burning engines have come into use.

These historical notes are given, not to explain the principles
of operation of the early engines, but to indicate the length of
the period in which the internal-combustion engine was evolved.

170. Cycles of Operation. Modern internal-combustion en-
gines operate upon either one of two cycles: the four-stroke

GAS ENGINES

195

cycle or the two-stroke cycle. These terms are usually abbre-
viated to four-cycle and two-cycle. The four-stroke cycle is some-
times called the Otto cycle, and the two-stroke cycle is occa-
sionally known as the Clerk cycle. Each of these cycles is used
with gas, gasoline, Diesel, semi-Diesel, and other oil engines.

171. The Four-stroke Cycle. In the four-stroke cycle there
are four strokes of the piston, two forward, and two backward.
These occur as the shaft makes two complete revolutions. Figure

Met

fxhoust

FIG. 137

137 represents an engine cylinder with its piston P. The inlet
valve is at A, and the exhaust valve at B. When the piston is
at its crank-end dead-center position, the volume to the left of
it is the piston displacement plus the clearance. Both valves, A
and B, are closed, and the space to the left of the piston is filled
with a mixture of air and fuel.

During the first stroke, the piston moves from its crank-end to
its head-end dead center, compressing the mixture into the clear-
ance space. Near the head-end dead center, the compressed mix-
ture is ignited, whereupon it burns, and its pressure suddenly
increases.

On the second stroke the piston moves from the head-end dead
center to the crank-end dead center, while the burnt gases expand
and do work upon the piston. Near the end of the second stroke,
the exhaust valve B opens.

As the piston moves to the left on the third stroke, the burnt
gas is forced out to the exhaust. The burnt gas that is left in
the clearance space is not expelled. At the end of the third
stroke the exhaust valve closes.

At the beginning of the fourth stroke, the inlet valve A opens.
As the piston moves to the right, a fresh charge of air and fuel

196

ENGINES AND BOILERS

is sucked into the cylinder. At the end of the fourth stroke the
inlet valve closes. This completes the cycle of operation.

Figure 138 shows the indicator card of the four-stroke cycle
engine. Starting from E, at a little below the atmospheric pres-

Atmospheric Line

FIG. 138

sure, the charge is compressed during the first stroke to the
pressure at the point A. From A to B, the charge is burned
and the pressure rises. From B to C expansion takes place. From
C to D the burnt gases are expelled. The suction stroke is rep-
resented by DE. The four strokes then represent compression,
burning and expansion, scavenging, and suction. On the card, L
represents the length of stroke, and F the clearance.

172. The Two-stroke Cycle. This cycle is completed in one
revolution. Figures 139 and 140 show a two-stroke cycle engine

FIG. 139

such as is often used on motor boats. Due to the fact that the
piston covers and uncovers the admission and the exhaust ports,
it is often called a valueless engine.

GAS ENGINES

197

In Fig. 139, the piston is shown at the crank-end dead center.
Both the inlet and exhaust ports are open. There is a small
compression in the crank case and the combustible mixture of
air and fuel is forced up and into the cylinder through the port A.
There is a baffle on the top of the piston which deflects the in-
coming mixture to the top of the cylinder. At the same time the
burnt gases from the previous stroke are escaping through the

FIG. 140

exhaust port B. Naturally a little of the unburned gas will es-
cape before the exhaust port is closed.

As the piston moves upward on its first stroke, it covers the
ports A and B, and then compresses the mixture into the clear-
ance space. At the head-end dead center, Fig. 140, ignition takes
place, and the piston is forced downward on the second stroke
by the pressure produced. The burnt gases expand until the
exhaust port B is uncovered. They then escape to the exhaust.
As the piston moves on down, the inlet port A is uncovered, and
the fresh gas coming in at A sweeps out more of the burnt gas.
As the piston moves upward a slight vacuum is formed in the
crank case. When the piston gets nearly to the top of its travel,
a port communicating to the carburetor or fuel and air supply
is uncovered, and the combustible mixture is sucked into the
crank case. Upon the piston's downward stroke this mixture is
compressed enough to force it into the cylinder when the port A
is uncovered.

In Fig. 141 another two-stroke cycle engine is shown. In the
large size of these engines, separate pumps for air and gas are

198

ENGINES AND BOILERS

used instead of the compression in the crank case. The engine
is double-acting, and the exhaust port is placed around the cylinder
midway between the two ends. The piston P uncovers this ex-
haust port near the end of each stroke. Gas and air is com-
pressed in the pumps shown. The piston valves of the pumps
deliver the gas and air alternately to the two ends of the cylinder.

FIG. 141

The air and gas are mixed just before they enter the admission
valve I. In Fig. 141, gas and air are compressed in the left ends
of the pumps, and are forced into the left end of the engine cylinder.
As the piston starts to the left, the left admission valve closes,
and the mixture is compressed into the clearance space. At dead
center the charge is fired, and expansion takes place as the piston
moves to the right. When the piston uncovers the exhaust port,
the burnt gases escape. In Fig. 141, the amount of gas forced
into the cylinder is controlled by butterfly valves whose position
is controlled by the governor.

The indicator card of the two-stroke cycle engine of Figs. 139
and 140 is shown in Fig. 142. Compression takes place from
E to A, burning from A to B, and expansion from B to C. From
C to D, and from D to E, the 'exhaust of the burnt gases takes
place. The admission of the charge occurs from H to D and

GAS ENGINES

199

from D to G. It is seen that the exhaust port opens a little
sooner than the inlet port.

173. Classification from Fuel Used. Internal-combustion en-
gines are called by different names, according to the fuel used.
We have the gas, gasoline, oil, Diesel and semi-Diesel engines.
The fundamental principle is much the same in all of them, dif-
ference being largely a matter of detail in design and in the feed
of the fuels.

In the earlier gas engines, city or coal gas was often used.
Then, in the days of the natural gas booms in this country, gas
engines became quite common. With these kinds of gas, it was

-& Atmospheric /ne

FIG. 142

not possible to give the mixture of fuel and air a very high com-
pression, because the temperature during compression might
be so high that ignition would occur before the end of the com-
pression stroke. With the use of producer gas or blast-furnace
gas the compression can safely be carried higher; hence we find
it common to use a less amount of clearance with engines designed
to use a lean gas for fuel. With a small clearance the compres-
sion is higher.

The internal combustion engine with which we are most familiar
is the gasoline engine. Gasoline will vaporize partially at ordi-
nary temperatures; hence a spray of the liquid fuel is mixed
with the air as it goes to the cylinder. While this spray is com-
monly not all vaporized before reaching the cylinder, it is sufficiently
vaporized so that an explosive mixture results, and the charge
is fired. If the spray is fine enough and if sufficient time is given
during the stroke, the gasoline will be very nearly all buined.
Of course there are differences in gasolines, some kinds being more
volatile than others. The device that introduces the spray into the

200

ENGINES AND BOILERS

air intake is called a mixing valve, or a carburetor. With the less
volatile liquid fuels, such as kerosene, heat is sometimes applied
to help in vaporizing it before it is introduced into the cylinder.

When liquid fuels heavier than gasoline are used, it is common
to spray them into the cylinder during the compression stroke.
Often the spray strikes a hot plate in the cylinder and the fuel is
sufficiently vaporized so that ignition can take place at the end of
the stroke. The compression in low-pressure oil engines is no
greater than in some gas engines, usually about 60 pounds per
square inch.

In the Diesel engine the clearance is much less and a com-
pression of 500 to 600 pounds per square inch is attained. Prac-

df/nospheric Line
FIG. 143

tically all liquid fuels will partially burn at a temperature at-
tained by the compression of air to 200 to 300 pounds per square
inch. To prevent premature ignition, the liquid fuel is sprayed
into the cylinder at the beginning of the forward or working
stroke in Diesel engines.

Air alone is compressed during the compression stroke. At a
pressure of 500 to 600 pounds its temperature will be in the
neighborhood of 1000 F. This is sufficient to ignite and com-
pletely burn the oil that is sprayed in. The length of time the
oil is injected into the cylinder usually is regulated by the gov-
ernor. The Diesel engine may operate on either the four-stroke
or on the two-stroke cycle.

The indicator diagram for the four-cycle Diesel engine is shown
in Fig. 143. On the first stroke, air is compressed from a pressure
a little below that of the atmosphere, shown at E, to 500 or 600

GAS ENGINES

201

pounds per square inch, as shown at A. From A to B, the fuel
is injected and burned. Expansion of the burnt gas takes place
from B to C. The cylinder is scavenged from C to D, and a
fresh charge of air is drawn in from D to E.

The high pressures necessary in the Diesel engine have been
found troublesome from a mechanical standpoint, and there has
been a tendency to reduce the high compression. With a com-
pression around 200 to 300 pounds the term semi-Diesel is used.
There is but little difference in the theory of operation between
the Diesel and semi-Diesel. Oil is injected at the opening of
the expansion stroke as before. However, on account of the

FIG. 144

lower temperature of compression, aid to the vaporization of the
oil is given by the addition of a hot bulb located in the head
of the cylinder.

Figure 144 shows this type of engine. Before starting, the
bulb is heated by means of the burner B. After the engine has
started the bulb will be hot enough without the aid of the burner.
In Fig. 144, air is drawn into the crank case and compressed as
the piston moves forward. With the piston near the crank-end
dead center, the exhaust port is uncovered and the air is blown
in through the ports at the top of the cylinder. As the piston
returns to the left, the charge of fresh air is compressed. With

202 ENGINES AND BOILERS

the piston on head-end dead center, the oil pump injects the fuel
into the cylinder, and the spray strikes the lip of the hot bulb.
The temperature of the bulb is high enough so that the fuel is
ignited and practically all burned.

This engine operates on the two-stroke cycle. On the larger en-
gines the four-stroke cycle is more common. On heavy loads,
there is often a tendency to knock in the semi-Diesel engine. This
is relieved by the injection of a small amount of water with the
fuel, which does not seem to lessen the efficiency of the engine.

174. Efficiency. The efficiency of an engine operating on the
Carnot cycle is (T\Tz)/Ti, where T\ is the absolute tempera-
ture during combustion, and T^ the absolute temperature of the
gases in the exhaust. Of course none of our engines operate on
the Carnot cycle, but their efficiency does depend upon the range
of temperature in the cylinder. Other factors being the same, it is
evident that the highest thermal efficiency will occur in that engine
which has the largest range of temperatures during the working
stroke. From this it is seen that those engines with the highest
compressions, and therefore the highest combustion temperatures,
will have the highest efficiency. This also explains the fact
that gas engines using a lean gas, such as blast-furnace gas, may
have a higher efficiency than those that operate on natural gas
or gasoline vapor. Of all the internal-combustion engines, the
Diesel has the highest thermal efficiency. The efficiency based
on the brake horsepower ranges from 30 to 35 per cent. Other
gas engines give somewhat lower efficiency.

175. Fuels. Gas engines may operate on almost any com-
bustible gas. Naturally the more expensive gases are but little
used. Ordinary city gas is commonly a mixture of coal gas and
water gas. As ordinarily produced it is too expensive for exten-
sive use in gas engines. Where natural gas is plentiful and cheap,
it is commonly used for gas engines.

At some blast-furnace plants, the gas from the furnace is used
for fuel in the gas engines that drive the blowers. This gas is
not of what we commonly call high quality, that is its heat con-
tent per cubic foot is much lower than that of natural gas or
coal gas, but it gives excellent results when used in the engines.

The by-product gas from coke ovens is being used more and
more as the efficiency of these plants is looked after. There are

GAS ENGINES 203

also quite a number of plants, especially in the east, where pro-
ducer gas is used for gas engines.

Natural gas, while it varies in composition, may be said to be
composed mainly of marsh gas, CH 4 . In some natural gases,
there is a considerable free hydrogen and also an appreciable
amount of olefiant gas, C 2 H4. The heat value of natural gas usu-
ally is between 900 and 1000 B.t.u. per cubic foot.

The illuminating gas used in cities varies widely in its com-
position, depending upon how it is made. It usually contains
about 40 per cent H2, 30 per cent CH 4 , and varying amounts of
CO and C 2 H 4 besides CO2 and N 2 . The heating value aver-
ages from 500 to 600 B.t.u. per cubic foot.

Coke-oven gas contains about 50 per cent H2 and 35 per cent
CH 4 . Its heating value is about the same as that of illuminating
gas.

The principal combustible substance in blast furnace gas is CO,
which ordinarily runs about 25 per cent. The heating value of
blast-furnace gas is but little over 100 B.t.u. per cubic foot.

Producer gas, while it is variable, contains about 15 per cent
H 2 , and 25 per cent CO. It has a heat value of about 145 B.t.u.
per cubic foot.

Practically all these gases contain a little 02 and varying amounts
of C02 and N2. These substances in the gas add no heating value
to it.

The liquid fuels used in internal-combustion engines vary from
crude oil to more refined products, such as gasoline. Many
Diesel engines seem able to burn crude oil very well, but some
of the semi-Diesel and oil engines do better on the more volatile
product, such as kerosene. The great demand for gasoline has
led to a gradual lowering of the flash-point of this product. Car-
buretors that used to give good results with the gasoline sold a
few years ago, now have trouble in using the commercial grades.
The development of carburetors has had to keep step with the
change in the quality of the product they have to handle.

176. The Gasoline Carburetor. There are a great variety of
mixing valves or carburetors on the market. We cannot hope
to describe all of the various types in this course. Only one simple
form will be described, but the fundamental principle is the same
in all. This principle is to divide the liquid fuel into as fine a

204

ENGINES AND BOILERS

spray as possible and to mix it thoroughly with air. Some of
the liquid is evaporated, but it is doubtful whether it is ever all
vaporized.

Evaporation is not necessary if the liquid particles are finely
enough divided and are thoroughly mixed with the air. Even
solid combustible matter is explosive when mixed with air, as
is shown by flour-mill explosions and the explosions in mines due
to dust of inflammable materials.

Figure 145 shows a gasoline carburetor. The main body of
the carburetor B is partly filled with gasoline. The height of

FIG. 145

the liquid is kept very near to a constant level by means of a
float F, which is attached by means of a lever to the valve H
which leads from the supply. Air enters through the opening
at the top and passes down through the passage C. Turning to
the left, it passes out to the pipe leading to the engine. A spray
of liquid is injected into the air through the needle valve D.
The opening in the needle valve D is adjusted by the handle E.
The opening of the needle valve into the air passage is a little
above the liquid level in B, so that no gasoline runs out unless
there is a current of air to suck up the liquid. There is at all
times an opening for the air to enter the carburetor at A, but
this opening may be increased when a large supply is needed by
the suction pulling back the valve A. The valve A is held on
its seat by a light spring 0. The tension in the spring may be

GAS ENGINES 205

adjusted by turning the screw M. By the proper adjustment
of the air valve A and the needle valve Z), any richness of mix-
ture may be secured. The amount of mixture of air and gaso-
line leaving the carburetor to enter the engine is regulated by the
throttle K. The bowl of the carburetor may be drained by open-
ing the cock T.

177. The Gas Producer. A large amount of publicity has
been given to the subject of gas producers for the gas engine,
and much research work has been done on the subject. It is only
just, however, to say that the producer plant has been somewhat
of a disappointment in America. While trial tests show up re-
markably well, and the claims of makers are unusually good,
actual experience has shown that the time has not yet come when
they can replace the steam plant. It is not safe to predict as
to the future, and every power plant engineer should be some-
what acquainted with the subject.

When air is passed through a bed of hot carbon, combustion
takes place. If there is sufficient air, the combustion is complete
and CC>2 is formed. If not enough air is supplied, the burning
is only partial and CO is formed. This CO may later be burned
to CO2 by the addition of more air, which is done in the engine
cylinder in the case of the producer and engine plant.

If steam is passed through a hot carbon bed, a decomposition
of the steam takes place. The hydrogen is liberated as H 2 and the
oxygen combines with the carbon to form CO. Both of these
gases are valuable as fuel, and the mixture is often called water
gas.

When air is passed through the hot carbon bed, and CO is
formed, heat is generated, so that the bed gets hotter and hotter.
On the other hand, when steam is passed through, heat is ab-
sorbed and the bed gets cooler and cooler. By the proper pro-
portioning of air and steam passed through, it is possible to keep
the fuel bed at the proper temperature. This is what is done
in the gas producer. It is evident that most of the gases in
producer gas will be CO, H and N, the nitrogen coming from the
air and being inert.

The fuel used in the producer may be coke or coal. Better
results are obtained by using anthracite coal than by using bitu-
minous coal. This is partly due to the fact that bituminous coal

206

ENGINES AND BOILERS

tends to cake and needs constant working or poking to keep holes
from burning through the cake, thereby letting excess air get
through. Moreover, bituminous coal gives off various tars when
it is heated. If these are not removed, they clog up the pipes
and the engine. The removal of the tar is not easy. It is some-
times done by throwing the tarry material out of the gas by
centrifugal force by means of a kind of fan arrangement, or by
passing the gas through scrubbers. Devices have been tried

FIG. 146

whereby the distilled products are passed through the hot fuel
bed and their composition thereby changed. This last method
resembles the underfeed furnace used in steam plants.

Figure 146 shows a form of gas producer. Coal is fed into the
hopper A. From this, it is dropped into the chamber B. Pass-
ing out of the bottom of B, it is scattered in a uniform layer over
the fuel bed by means of the spiral spreader C, which is rotated
by the bevel gear Q.

The fuel bed may be divided roughly into three zones. The
top zone is the green-coal zone, where distillation takes place.
The volatile products pass on out with the other gas.

As the volatile products are driven off, and the bed settles, the
fuel reaches the coke zone. It is here that the burning and de-
composition mentioned above take place. As the carbon is burnt

GAS ENGINES 207

out, the ash settles to the bottom of the producer, where it is
raked out through the water seal R.

Air from the duct E is led to the bottom of the fuel bed through
D, and passes up through the coke zone and the green-coal zone.
Steam is admitted to the air supply through the pipe G. Holes
are provided in the furnace walls at F for the working of the fuel
bed.

The gas leaves the producer through the opening at 7, and
passes down the pipe shown to K. From K it passes through
the pipe L to the wet scrubber. The wet scrubber is filled with
coke, Pj which is continuously sprinkled with water from the
nozzle N. The gas passing up through the wet coke is cooled
and deposits dust, tar and other impurities. The gas leaves the
wet scrubber at 0, and may go either directly to the engine or
else to a dry scrubber. The dry scrubber is filled with wood
shavings or excelsior, which takes out the remaining tar.

Upon starting the producer, the gas is vented to the roof through
the valve J. As soon as the quality of the gas becomes good
enough, the valve J is closed and the engine is started. The coal
in B is kept cool enough to prevent it from burning, by a water-
jacket S. The producer is lined with firebrick.

Air either may be blown into the producer, or it may be drawn
through by the suction of the engine. In the former case a
storage tank for the gas is necessary. With the suction type the
storage is unnecessary, as the engine draws through only what it
needs. When a water seal is used at the bottom, the producer
is called a wet bottom producer. The poking of the fuel bed
may be done by hand or mechanically.

178. Cooling of Cylinders. The cylinder walls of the internal-
combustion engine must be kept cool enough to insure proper
lubrication. This cooling is commonly done by circulating water
through a jacket around the cylinder, as shown at W, in Figs.
137, 139, 140, 141 and 144. As far as the efficiency of the engine
is concerned, the hotter the cylinder walls the better, so that it
is evident that they should not be cooled any more than is neces-
sary to insure lubrication. In small engines, the cylinder is
sometimes placed in the lower part of a hopper filled with water.
Fresh water is added as it boils away in the hopper.

Another method that is used occasionally to cool the cylinder

208

ENGINES AND BOILERS

Met

FIG. 147

is by having the outer walls of the cylinder and head covered with
fins. These present a large surface to the air and the heat is
radiated from them. To increase this transfer of heat, a cur-
rent of air is kept moving over the
surface of the fins by means of a
fan. This latter method is called
air-cooling. Figure 147 shows an
air-cooled cylinder.

179. Ignition. The charge of
combustible mixture in an internal-
combustion engine is fired in vari-
ous ways. A method formerly used
quite extensively but now not very
common is the hot-tube method.
This method is illustrated by the
sketch in Fig. 148. The tube which
connects with the cylinder is heated
by a gas jet. Arrangement is made

so that the flame can be shifted along the tube, heating it closer

or farther away from the cylinder.

To explain how the scheme works, suppose the charge has been

fired. The tube will then be filled with burnt gas. After the

fresh charge has been drawn in

and compression started, the fresh

gas is forced into the tube. When

it is compressed into the tube far

enough to strike the heated part,

ignition occurs. Naturally, this

scheme can be used only when the

load is constant.

As has been explained previ-
ously, ignition may be had by

using a high compression, so that

the temperature of compression

will be high enough to fire the

charge. This method is used

mostly in oil engines and is assisted in those using the lower

pressures by a hot bulb or plate. The hot bulb acts somewhat

in the same manner as the hot tube just mentioned. With gas

FIG. 148

GAS ENGINES 209

or the more volatile liquid fuels this method is not satisfactory,
since early ignition is apt to occur.

The most common method is electric ignition. There are two
general types of electric ignitors, the jump-spark and the make-
and-break. In the former, a spark plug is used (Fig. 149), which
has two fixed terminals exposed to the gases of the cylinder.
At the proper time for ignition a current with voltage high enough
to jump the gap is introduced in the circuit. The
heat of this spark ignites the charge. The details of
timing and of producing the current will not be dis-
cussed here, except to say that the current may be
furnished either by dry-cell or storage batteries, or by
a magneto.

In the make-and-break system, two electrodes are
brought into contact within the cylinder and are sepa-
rated at the proper time for ignition. As the circuit is
broken a spark is formed between them. The details
of the many schemes used will not be discussed here.
This make-and-break system does not require as high
an e.m.f. as the jump-spark system, but it is limited to the slower
engine speeds.

180. Valves. The earlier types of gas engines were equipped
with slide-valves. These required lubrication, which is some-
what difficult at high temperatures. Except for a few engines
that use sleeve-valves, gas engines of today have the so-called
lifting or poppet valves , which are commonly lifted .by cams on
a cam shaft. In the four-stroke cycle engines the cam shaft is
geared to run at half the speed of the crank shaft, so that the
cam shaft makes one revolution per cycle. The cams are placed
on the cam shaft so that the valves open and close at the proper
time. Each valve is kept seated by means of a spring around
the valve stem. Figure 147 shows the cams, and how they are
made to lift the valves. In some of the slower speed engines,
the inlet valve is not opened by means of a cam, but by the
suction inside the cylinder. With this arrangement a strong
spring cannot be used to seat the valve.

181. Governing. As far as the governor itself is concerned,
the gas-engine governor does not differ materially from the steam-
engine governor. Both centrifugal and inertia governors are used.

210 ENGINES AND BOILERS

Depending upon how the governor regulates the amount of work
done in the engine cylinder, we have three general types of gas
engine governors: (1) hit-and-miss governors, (2) quantity gov-
ernors, and (3) quality governors.

HIT-AND-MISS GOVERNOR. With this type of governor there
is a working stroke for every cycle under conditions of maximum
load. At lighter loads the governor mechanism fails to admit
a charge occasionally, giving what might be termed a blank cycle
in which no work is done by the cylinder. This drops the speed
of the engine and the governor acts so that fuel is again taken in
as before. With this scheme the engine either operates under
conditions of maximum efficiency, or it does not fire at all. This
method of governing gives better economy at light loads than the
other methods, but it does not give close speed regulation. When
the engine misses, the exhaust valve is commonly held open so
that there is no work done in useless compression.

QUANTITY GOVERNOR. For every engine there is a ratio of gas
to air, which is nearly constant, with which the engine gives the
best efficiency. With the quantity governor, this ratio is kept
(theoretically) constant. Regulation is accomplished in two ways :
(1) by the cut-off governor, and (2) by the throttling governor.
With the cut-off or throttling governor, the normal mixture is
allowed to enter the cylinder only during a part of the suction
stroke at light loads. The length of time the mixture is ad-
mitted is controlled by the governor. With the throttling gover-
nor, the normal mixture is taken in during the whole of the suc-
tion stroke, but the opening is throttled so that not as much
enters at light loads as at full load.

QUALITY GOVERNOR. This governor changes the ratio of the
fuel to the air at different loads. At full load, a rich mixture is
used, and at light load a lean mixture. Mechanically, this scheme
is quite simple, but it has the disadvantage of giving low effi-
ciency at light loads. If the mixture gets too rich or too lean,
it may be impossible to secure ignition. Oil engine governors
commonly control the amount of oil admitted per working
stroke. This is seen to be the same as the quality-governing
scheme.

It should be mentioned that the speed of an engine can be
regulated by changing the time of ignition. With either too
early or too late ignition the full power is not developed in the

GAS ENGINES 211

cylinder. The speed of motor-boat engines is often controlled
in this way. With high speeds, the spark should occur earlier
than in low-speed engines. With a variable-speed engine, such
as exists in an automobile or truck, the time of ignition should
be adjusted to the speed in order to get the best results.

182. Determination of Horsepower. The indicated horse-
power of a gas engine is determined in the same way as for the
steam engine, with the exception that for four-stroke cycle en-
gines only half the r.p.m. is used in the computation. If a
hit-and-miss governor is used on the engine, the number of hits
per minute must be counted rather than the r.p.m. of the shaft.

183. Multi-cylinder Engines. The single cylinder, single-act-
ing four-stroke-cycle gas engine has one impulse stroke in two revo-
lutions. The double-acting steam engine has two impulse strokes
per revolution. Thus it is seen that the single cylinder gas en-
gine has a much greater variation in angular acceleration of crank
shaft than does the steam engine. For some kinds of service this
variation in angular velocity is immaterial; in other cases it is
a serious disadvantage. For instance, in the generation of elec-
tric current to be used for lighting, a single-cylinder engine is
impracticable, unless an exceedingly heavy flywheel is used. To
approximate the uniformity of torque that exists in a single-
cylinder steam engine, it is necessary to use four cylinders on the
gas engine.

In automotive service, a fairly uniform torque is desirable, and
therefore four or more cylinders are used. If more than four
cylinders are used, there will be less variation in the angular
torque, and the engine speed may be controlled more easily by
the throttle. Too large a number of cylinders may cause a de-
crease in the efficiency of the engine. This may be explained
by the fact that with a multi-cylinder engine, there is more area
of cylinder wall exposed to the burned gas for the same volume
of gas than there is in the single-cylinder engine. Principle I of
169, as set forth by Beau de Rochas, is a statement of this
same fact. While a lowered efficiency may result from the use
of a large number of cylinders, this loss may be more than com-
pensated by the added smoothness of running. Many of the
higher grade automobiles made at the present time are equipped
with six, eight, or even twelve cylinders.

212

ENGINES AND BOILERS

PROBLEMS 213

PROBLEMS

1. (a) What is the pressure in pounds per square inch that corresponds to
a mercury column 16 inches high?

(6) What is the atmospheric pressure when the barometer reads 27.4 inches?

2. A steam gage is used to show the pressure in a steam line and is at-
tached as shown in Fig. A. If the small pipe leading to the gage is full of
water and the gage reads 183 pounds, what is the pressure in the steam line?

3. If the pressure gage on a
boiler reads 150 pounds and the
barometer reads 29.3 inched, what
is the absolute pressure in the boiler
in pounds per square inch?

4. The vacuum gage on a con-
denser reads 27.2 inches and at the
same time the barometer reads 29.1
inches.

(a) What is the absolute pres-
sure in the condenser in pounds per
square inch? (6) What is the vac-
uum-gage reading reduced to a JT IG
30-inch basis?

6. A condenser with its air pump is guaranteed by the manufacturer to
produce a vacuum of 28.5 inches (on the basis of a 30-inch barometer). Dur-
ing the acceptance test the barometer read 28.73 inches. What should be
thevacuum-gage reading to maintain the guaranteed vacuum?

Q) A boiler-feed pump is located 14 feet below the water line of the boiler.
The pump draws water from a tank located 7 feet below the pump cylinder.
If the pressure in the boiler is 40 pounds gage, neglecting friction losses due
to the flow of water, etc., what is the least total head the pump must act
against: (a) In feet? (6) In pounds per square inch? (c) What is the
least height of water level in a standpipe above the boiler in order that the
water will flow into the boiler by gravity?

(Y) Reduce: (a) A temperature reading of 50 Centigrade to the corre-
sponding Fahrenheit reading. (6) A temperature of 320 Fahrenheit to
Centigrade. (c) 75 (great) calories to B.t.u. (d) 46 B. t. u. to calories.
(e] 33000 foot-pounds to B.t.u.

8. If 1,576,000 B.t.u. are given to an engine in an hour and if the engine
can convert 6 per cent of this heat into work, what is the horsepower of the
engine? (One horsepower is 33,000 foot-pounds per minute.)

9. A sample of Indiana coal gave the following proximate analysis:

Moisture =3.81%, Fixed carbon = 76. 16%,

Volatile combustible = 13.62%, Ash = 6.41%.
The same sample when dried gave the following ultimate analysis:
Carbon =84.26%, Oxygen =1.73%, Sulphur = 1.22%,

Hydrogen= 4.38%, Nitrogen = 1.75%, Ash =6.66%.

The oxygen calorimeter gave a calorific value of 14682 B.t.u. per pound of
dry fuel. Find the calorific value of the preceding sample per pound of dry
fuel, from (a) the proximate analysis, (6) the ultimate analysis.

214 ENGINES AND BOILERS

A sample of West Virginia coal gave the following proximate analysis :
Moisture =4.85%, Fixed carbon = 68.36%,

Volatile combustible = 16.31 %, Ash = 10.84%.
The same sample when dried gave the following ultimate analysis:
Carbon =80.34%, Oxygen = 3.11%, Sulphur = .49%,
Hydrogen = 4.00%, Nitrogen = 1.05% Ash =11.01%.
The oxygen calorimeter gave for a similar dried sample a calorific value of
14,180 B.t.u. per pound. Find the calorific value per pound of dry coal from
^ (a) the proximate analysis (b) The ultimate analysis.

(llj Find the theoretical weight of air required to burn completely a
pound of the dry coal of Problem 10.

12. A boiler test was run using the coal from which the sample of Problem
10 was taken. During the test the analysis of dry flue-gas was as follows:
Carbon dioxide = 8.58%, Carbon monoxide = .05%,

Oxygen =11.32%, Nitrogen = 80.05%.

Find the approximate weight of air used to burn one pound of dry coal.
(13? In the test of Problem 12, the temperature of air entering the furnace
was'64 F., and the stack temperature was 559 F. Find the percentage of
available heat carried up the stack by the dry flue-gas.

14. Water is fed to a boiler at a temperature of 170 F. The pressure
gage reads 140 pounds, and the barometer 29.6 inches. How many B.t.u.
are needed to generate a pound of dry steam? What is the temperature of
thesteam generated? The volume per pound?

Q$ Steam at a gage pressure of 135 pounds is generated from water at
120 F. The temperature of the steam is 490 F. The barometer reads
29.3 inches. Find the B.t.u. required to generate one pound of steam.

16. If the temperature of steam in a condenser is 115 F., what is the great-
estjaossible vacuum-gage reading, if the barometer reads 29.16 inches?

Qj) If water boiling under a pressure of 185 pounds gage is allowed to
escape to the atmosphere (as in a boiler explosion), what percentage of its weight
turns to steam? What is the ratio of its new volume to the old? Assume
that the barometer reading is 29.6 inches.

18. Dry steam leaves a boiler at a pressure of 180 pounds gage and reaches
the engine with a quality of 98 per cent, and a pressure of 177 pounds gage.
What percentage of its heat contents has it lost in its passage through the
pipe? What percentage of its volume? Assume that the barometer reading
is 29.43 inches.

Q9^ If one pound of steam of 95 per cent quality at atmospheric pressure is
mixed with 8 pounds of water at 70 F., what will be the resultant tempera-
ture? Assume that the barometer reading is 29.00 inches.

20. Dry steam enters a turbine at a pressure of 180 pounds gage; leaving
the turbine it passes into a condenser in which the vacuum is 27.6 inches (30-
inch basis). The quality of steam as it leaves the turbine is 87%. Neglect-
ing all losses, find how many foot-pounds of work may be obtained from each
pound of steam that passes through the turbine.

/2L) A frictionless piston weighing 7000 pounds is placed in a vertical
cyimcler 10 inches in diameter. Two pounds of water at 70 F. are placed

PROBLEMS 215

under the piston. If 800 B.t.u. are added to the water, how far will the pis-
ton move? The barometer reads 29.6 inches.

22. If, in Problem 21, 2500 B.t.u. are added to the water, what will be
the weight of steam formed? What will be its temperature? How far will
thejMston move?

2iP A sample of steam is taken from a steam line in which the pressure
is 150 pounds gage and is led to a throttling calorimeter in which the tem-
perature is 230 F. and the gage pressure is 3 pounds. The barometer reads
29.4 inches. What is the quality of steam in the line?

24. A horizontal water-tube boiler (B. and W. type) has 10 vertical rows
of four-inch tubes, 9 tubes to the row. The tubes are 18 feet long, and the
steam drum is 24 feet long and 42 inches in diameter. Find the heating
surface and the rated horsepower (a) by using as heating surface the out-
side surface of the tubes and one-half the surface of the drum; (6) by
RuJle3, p. 26.

j2J5c A horizontal return tubular boiler 60 inches in diameter and 18 feet
long has 44 four-inch tubes. Find the heating surface and rated horsepower
(a) by Rule 1, p. 26, (6) by Rule 3, p. 26.

26. A Scotch marine boiler-shell is 16 feet 3 inches in diameter and 12
feet long. There are three furnaces, each 43 inches in diameter. The boiler
contains three sections of tubes, each section consisting of 110 three-inch
tubes 10 feet long. Find the approximate heating surface and the horse-
power.

27. A vertical fire-tube boiler (exposed-tube type) has a diameter of 30
inches and a height of 6 feet. The furnace is 25 inches in diameter and 27
inches high. There are 55 two-inch tubes 45 inches long. The normal water
level is 10 inches from the top of the tubes. Find the heating surface and
rated horsepower by Rule 2, p. 26.

^5P In a test of a B. and W. boiler with a hand-fired furnace at the Sew-
age Pumping Station, Cleveland, Ohio, the following data were taken:

Rated horsepower of boiler 150

Grate surface 27 square feet

Duration of test 24 hours

Steam pressure 156.3 pounds gage

Temperature of feed water 58 F.

Quality of steam formed 99 per cent.

Total weight of coal fired (wet) 15078 pounds.

Moisture in coal 7.5 per cent.

Total weight of water fed to boiler 105100 pounds.

Find:

(a) Factor of evaporation.

(6) Dry coal per square foot of grate surface per hour.

(d) Equivalent evaporation per hour per square foot of water-heating
surface.

(e) Boiler horsepower developed.

(L Percentage of rated capacity developed.
(297 In the test of Problem 28, the dry coal had a calorific value of 12292

216 ENGINES AND BOILERS

B.t.u. per pound, and the cost delivered at the boiler room was $3.50 per
ton of 2000 pounds. Find:

(a) Equivalent evaporation from and at 212 per pound of dry coal.
Combined efficiency of boiler, furnace and grate.
Coal cost per 1000 pounds of equivalent evaporation.
In a test of a B. and W. boiler the following data were taken:

Rated horsepower of boiler 508

Grate surface 90 square feet

Duration of test 16.25 hours

Steam pressure 199 pounds gage

Temperature of feedwater 48.4 F.

Superheat 136.5 F.

Total weight of coal fired (wet) 39670 pounds

Moisture in coal 4.22 per cent

Total weight of water fed to boiler 336200 pounds

Find:

(a) Factor of evaporation.

(6) Dry coal per square foot of grate surface per hour.

(d) Equivalent evaporation from and at 212 per hour per square foot
of water-heating surface.

(e) Boiler horsepower developed.
Percentage of rated capacity developed.

The coal in the test of Problem 30 gave the following proximate anal-
ysis when dry: volatile combustible, 19 66 per cent; fixed carbon, 75.41 per
cent; ash, 4.93 per cent. The cost delivered to the boiler room was S3. 75 per
ton of 2000 pounds. Find:

(a) Equivalent evaporation per pound of dry coal.
(6) Combined efficiency of boiler, furnace and grate.

Coal cost per 1000 pounds of equivalent evaporation.
Is the boiler of Problems 28 and 29 working harder than that of
Problems 30 and 31, or conversely? Give the reason for your answer.

33. Find the size of a pop safety-valve with a 45 seat for a 60-horsepower
return-tubular boiler which is to carry a gage pressure of 75 pounds. Assume
that the maximum evaporation is 5 pounds of water per hour per square foot
of wetter-heating surface, and that the lift of the valve is 1/30 of the diameter.
(3p How many 2.5-inch pop safety-valves would one 4.5-inch valve re-
place, assuming that the lift is proportional to the diameter?

35. How many 3.5-inch pop safety-valves are required for the boiler
of Problem 30? Assume the rate of maximum evaporation as 6 pounds of
water per square foot of water-heating surface per hour, and that the lift
is 1/30 of the diameter.

36. What should be the size of the pop safety-valve for the boiler of
Problem 28

(a) Computed as in Problem 35?

(6) Computed from the P. G. Darling formula? See p. 59.

(d) Computed from the city of Philadelphia formula? See p. 59.

PROBLEMS 217

(e) Computed from the U. S. Supervising Inspectors' formula? See p. 59.

(/) Computed from the A. S. M. E. Boiler Code Committee's require-
menjt^? See Report of Boiler Code Committee of A. S. M. E.
/3T,/ What should be the size of a steam pipe leading from a 250-horse-
power boiler if the pressure carried is 160 pounds gage? Assume a velocity
of florin the pipe of 5000 feet per minute.

(ZSy/A 5000-kw. steam turbine requires 16 pounds of dry steam per
hour per kw. at 160 pounds gage pressure. The vacuum in the exhaust
of the turbine is 27.5 inches of mercury (30-inch barometer). The quality
of steam in the exhaust is 85%. If the velocity of flow of steam to and
away from the turbine is to be 7500 feet per minute, what should be the
size _ol steam and exhaust pipes?

/39.) If a steel steam pipe is to carry steam at a pressure of 200 pounds
gageand may be as cold as 30 F. when the steam is cut off, how far apart
should expansion joints be placed if each joint gives a 3-inch movement?

40. If 9536 pounds of water at a temperature of 60 F. are mixed with
1160 pounds of steam at 3 pounds gage pressure, the steam being of 90 per
cent quality, what will be the resultant temperature of the mixture?
^41. The exhaust from a 65-horsepower steam engine is led to an open
feedwater heater. The engine uses 30 pounds of steam per hour per horse-
power, and the quality of the exhaust steam is 80%. The heater is at atmos-
pheric pressure; water enters at 50 F. and is heated to 200 F.

fa) What horsepower of boilers will the heater supply?

(6) What should be the size of steam and water pipes leading to the
heater? Assume a steam velocity of 5000 feet per minute and a water veloc-
ity of 150 feet per minute.

42. A 4000-kw. steam turbine is equipped with a surface condenser
The turbine uses 16 pounds of steam per kw. per hour, which enters the
condenser at a quality of 85 per cent. The vacuum to be maintained is 28
inches (30-inch basis). The circulating water enters the condenser at a
temperature of 60 F., and leaves at a temperature 10 cooler than that of
the incoming steam. (a) How much circulating water is needed per hour?

(6) If the same amount of water is circulated as in part (a), but if it enters
at 90 instead of 60, and leaves at 10 cooler than the incoming steam, what
vacuum can be maintained?

43. An 18"X24" steam engine has a piston rod 2.75 inches in diameter.
Find the head-end and the crank-end piston displacements in cubic feet.

44. If it takes 10.6 pounds of water to fill the head-end clearance space
and 11.2 pounds to fill the crank-end clearance space of the engine in Problem
43, what is the percentage of clearance for each end of the engine?

45. Find the volume of steam back of the piston of the engine of Problem
43: when the piston is at 12.4 per cent of the head-end stroke; when it is
at 14.0 per cent of the crank-end stroke.

^ 46. Find the weight of dry steam back of the piston of a 24" X 36" engine
when it is at 30 per cent of the head-end stroke. The head-end clearance is
4 per cent and the steam pressure back of the piston at the above position
is 105 pounds gage. If we know that at this time there is actually 1.06 pounds
of wet steam back of the piston, what must be its quality?

218

ENGINES AND BOILERS

Construct a hypothetical indicator diagram, using the following data.
Length of diagram =4 inches (this does not include clearance).
Initial pressure = 150 pounds per square inch (gage).
Back pressure =5 pounds per square inch (gag).
Cut-off = 25 per cent, Release =95 per cent.
Compression = 15 per cent, Admission = 2 per cent.
Clearance = 7 per cent.

Atmospheric pressure = 15 pounds per square inch,
as a scale of pressure 60 pounds per inch.

(Construct a hypothetical indicator diagram for a uniflow engine
108), using the following data.
Length of diagram =4 inches. Initial pressure = 170 pounds.
Back pressure = 12 pounds (engine is running condensing).
Cut-off = 20 per cent. Release and compression each =90%.

Admission = 2 per cent. Clearance = 3 p r cent.

Also show, by a dotted line on the same diagram, the compression curve when
the engine runs non-condensing (back pressure = 0).

State in what ways this excessive compression may be relieved.
49. Compute approximately the percentage of head-end and of crank-end

clearance of the engine from which
the cards of Fig. B were taken. Use
two methods. Cards were taken
with an 80-pound spring.

50. Compute the engine con-
stant, or the horsepower constant
(LA/33000), for the head end and
for the crank end for a 10"X14"
engine with a 2" piston rod. Your
answer must be correct to within

onephalf of one per cent.

-p IG -g (51/ Find the indicated horse-

power (i. hp.) of the steam engine of

Problem 50 when the head-end mean effective pressure is 34.2, and the crank-
endgjke.p. is 35.4 pounds per square inch. The engine is running at 260 r.p.m.
(52/ A test was run on a 14" X 18" steam engine with a 2" rod. The
head-end m. e. p. was found to be 35.2 pounds per square inch and the crank-
end m.e.p. 34.6 pounds per square inch. The speed of the engine was
250 r.p.m. The power was absorbed by a Prony brake whose arm is 6' 5"
long. The effective weight of the brake arm on the scales was 45 pounds.
During the test the pressure on the scales was 382 pounds. Find (a) the indi-
cate4\h rse P wei S (&) the brake horsepower; (c) the mechanical efficiency.

H>3/ The test of Problem 52 was run for 45 minutes, during which time
thVengine used 2750 pounds of steam at a pressure of 120 pounds gage, and
at a quality of 97 per cent. Find:

(a) Dry steam used per indicated horsepower per hour.
(6) B.t.u. per indicated horsepower per minute,

(d) Thermal efficiency based on b. hp.

PROBLEMS

219

The indicator diagram in Fig. C and the following data were taken during

cut-off

re/ease

compression
FIG. C

a test of a Buckeye engine.

Size of engine, 7.75" X 15", \\\"

rod.
Radius of Prony brake arm =6.02

feet.

Room temperature = 73.5 F.
Temperature in throttling calo-
rimeter =221.5 F.
Steam pressure at throttle = 128.7

pounds per square inch gage.
Steam pressure in calorimeter =

1.125 pounds gage.
Barometer = 28.5 inch.
R.p.m.= 222.5.
Net brake load = 140 pounds.
Scale of indicator spring = 80 pounds.
Steam used per hour = 1161 pounds.

54. Find the m.e.p. of the cards by the mean-ordinate method.
65. Find the indicated horsepower, head-end, crank-end, and total.
Find the brake horsepower.
Find:

Mechanical efficiency.
Pounds of steam per i. hp. per hour.
B.t.u. per i. hp. per minute.
Thermal efficiency based on b. hp.

Determine from each card the percentage of stroke and the steam pres-
sure for each of the following events :
(a) Cut-off.
(6) Release
(c) Compression.

59. Determine the weight of dry steam back of the piston for each end

at the events of cut-off, release, and
compression.

60. Find the amount of re-evapo-
ration or condensation per hour dur-
ing expansion.

61. Find the weight of dry steam
per hour per indicated horsepower
accounted for by the cards.

56.
57.

(a)
(6)
(c)
(d)
58.

ZO pound s/orinq.

FIG. D

62. Combine the indicator dia-
grams shown in Fig. D, and deter-
mine the diagram factor. The cards
of Fig. D were taken from an 8.02"
X15"X24" cross-compound Corliss

steam engine, running at 85 r.p.m. The head-end clearance of the high-pres-
sure cylinder is 7.4 per cent and the head-end clearance of the low-pressure
cylinder is 6.01 per cent.

220

ENGINES AND BOILERS

Determine the size of cylinders for a compound, two-cylinder, double-
acting steam engine (receiver type), assuming the following data: i. hp. = 120,
r. p.m. = 100, cylinder ratio = 1/3, piston speed = 600 feet per minute, initial
steam pressure = 140 pounds absolute, termin^ljgressure of hypothetical dia-
gram =14 pounds absolute, vacuum = 24 inches (30-inch basis), and diagram
f actor = .85

QJ4,/ In a certain two-cylinder compound steam engine the number of ex-
pansions is 10, the initial steam pressure is 120 pounds absolute and the back
pressure is 5 pounds absolute. The receiver pressure is 30 pounds abso-
lute. The cylinder ratio is 1 to 3. Neglecting clearance and piston rods,
compare the work done in the two cylinders and the stresses on the two
piston rods.

65. Given a cross-compound steam engine, show by means of a graph the

l/j/ve shown in mid-position

"4* >l

\ head
end

/ohfon

-*r

cnwft \

end

FIG. E

variation in power distribution when the governor varies the cut-off equally
in each cylinder (choose at least three cut-offs).

66. Proceed as for Problem 65, but assume that the governor varies the
point of cut-off in the high pressure cylinder only.

67. Consider a 12"X18" steam engine (section of cylinder and valve
shown in Fig. E), with the following given data.

Connecting rods 6 feet long.

Val ve- travel = 6 inches.

Head-end lead = crank-end lead = .25 inch.

Head-end steam lap = 1.25 inches.

Head-end exhaust lap = .5 inch.

Width of port = 1.75 inches.

(a) Draw the valve on its seat, the crank position, the eccentric position,
and the position of the piston in the cylinder when the crank is on head-end
dead center. (Make your drawing y actual size.)
(6) Draw the same parts for head-end cut-off.

(d) Draw the same parts for head-end release.

(e) Draw the same parts for head-end compression.

(/) Determine the percentage of the stroke for each of the above events.

PROBLEMS

221

'Consider a 14"X16
L aruTwith the following data.

engine, running over, with direct slide-valve

valve-travel = 4 inches,

lead = 1 A inch,'

steam lap = 1 inch, v

exhaust lap = K inch/

/ (a) Valve ellipse. Draw the crank circle K actual size, and about the
same center draw the eccentric circle full size. Choose 12 equidistant crank
positions and find the corresponding eccentric positions.

For any crank position (as C, Fig. F 2 ), the piston is at a distance x from
its mid-position, and at the same time the eccentric is at a distance y from its

Jfeam /ap+/ead

Crank on
head-end
dead center

FIG. F 2

FIG. F 4

mid-position. Plot y vertically and x horizontally, for all 12 positions of the
crank.

Connect the points thus found by a smooth curve. Label on this diagram
the following details: the crank position at each event of the stroke, the lead,
the steam lap, the exhaust lap, the maximum port-openings, and the angle
of advance.

(6) Bilgram diagram. Draw the crank and eccentric circles and choose
12 equidistant crank positions as in (a) . For each crank position (as C in Fig.
F 3 ), draw a dotted line parallel at a distance y from the crank. The inter-
section of these dotted lines is the Bilgram construction point P. About this
point P, draw in the steam-lap and exhaust-lap circles.

Show on this diagram the crank position at each event, the lead, the steam
lap, the exhaust lap, the maximum port-openings, and the angle of advance.

(c) Zeuner diagram. Draw the crank and eccentric circles as before, and
choose 12 equidistant crank positions. Lay off radially on the crank from the
center of the crank circle the eccentric displacement y (Fig. F 4 ) ; connect all
points thus found by a smooth curve.

Show on this diagram the crank position at each event, the lead, the steam
lap, the exhaust lap, the maximum port-openings, and the angle of advance.

222 ENGINES AND BOILERS

69. Consider an engine with the following given data.

Direct slide-valve. Head-end steam lap = 1| ".

Engine running over. Crank-end steam lap = 1 inch.

Valve-travel = 5 inches. Head-end exhaust lap = J4 inch.

Head-end lead = 1/8 inch. R/L = 1/5.

Find the head-end and crank-end crank positions, and the percent of stroke
at each event by means of

(a) The valve ellipse, (6) the Bilgram diagram, fc) the Zeuner diagram.

70. Consider an engine with a direct slide-valve and with the following
given data:

Engine running over. Crank-end cut-off =50 per cent.

Valve-travel = 3 inches. Head-end compression = 25 per cent.

Head-end admission = 1 per cent. Crank-end compression = 25 per cent.
Head-end cut-off = 50 per cent. R/L = 1/6.

Find the percentage of stroke at all events, the angle of advance in degrees,
the steam laps, the exhaust laps, the maximum port-openings, and the leads,
by means of

(a) The Bilgram diagram, (6) the Zeuner diagram.

Draw the eccentric circle full size and the crank circle to such a scale that
it is the same size as the eccentric circle. Label all of the dimensions asked
for directly on the diagrams, also label the head end of the diagram and in-
dicate by an arrow the direction of rotation of the crank.

71. Consider an engine with an indirect slide-valve and with the follow-
ing given data.

Engine running over.

Valve-travel = 4 inches.

Head-end lead = | inch.

Crank-end lead = Y inch.

Head-end cut-off = 35 per cent.

Head-end compression = 15 per cent.

Sum of steam lap and exhaust lap the same for both ends. R/L = %.
Find the percentage of stroke at all events, the angle of advance in degrees,
the steam laps, the exhaust laps, and the maximum port-openings, by means of
(a) The Bilgram diagram, (6) the Zeuner diagram.

72. Consider an engine with a direct slide-valve and with the following
data.

Engine running over.

Head-end admission = 2 per cent.

Head-end cut-off = 60 per cent.

Head-end maximum port-opening = 1 .25 inches.

Crank-end maximum port-opening = 1.25 inches.

Head-end compression = 20 per cent.

Crank-end compression = 20 per cent.

Find the valve-travel, the angle of advance, and each of the laps, by means
of

(a) The Bilgram diagram, (6) the Zeuner diagram.

Also draw to scale the valve on the seat in its mid-position.

PROBLEMS

223

73. Consider an engine with a direct slide-valve and with the following
data.

Engine running under.

Head-end lead = J^ inch.

Crank-end lead= f inch.

Head-end cut-off = 55 per cent.

Head-end compression = 20 per cent.

Crank-end compression = 20 per cent.

Head-end maximum port-opening = 1.25 inch.

Find the valve-travel, the angle of advance, and each of the laps by
(a) The Bilgram diagram, (6) the Zeuner diagram.
Draw the valve to scale in mid-position.

FIG. G

74. Consider an engine with a valve whose dimensions and seat are as
shown in Fig. G. The valve is not shown in mid -position. The valve-travel
is 4 inches; R/L = l/5; the engine runs over.

The cards of Fig. H are taken with the valve as now set. Find the angle
through which the eccentric must be shifted (state whether backward or for-

Cut-otf

Cuf-off

Crank-end card

FIG. H

ward), and the amount the valve stem must be lengthened or shortened (state
which), in order to give a cut-off of 25 per cent on each end. Draw the ap-
proximate cards for the new setting.

224

ENGINES AND BOILERS

75. Consider an automatic shaft governor with the following given data
(The Rites inertia governor is shown in Figs. I and J.) :
Head-end steam lap = 1.75 inches.
Head-end exhaust lap = 0.

Lead at normal position of eccentric = 5/32 inch.
Distance of eccentric center from pivot (R) = 12".
Distance from center of shaft to pivot point (x) = 13f ".

FIG. I

Location of point DE = 25", /=18".
Location of point A: c = 30", 6 = 52".

Cut-off at no load = 10 per cent.
Cut-off at full load = 65 per cent.
Direct slide-valve, engine running over.

Draw the governor analysis (full size) and find the valve-travel and

angle of advance

(a) At 10% cut-off, (6) at 65%
cut-off, (c) at normal cut-off.

(d} Also find percent of normal
cut-off.

II. Draw the head-end Zeuner
diagram, or the Bilgram diagram, for
(a) 10% cut-off, (6) normal cut-
off, (c) 65% cut-off.

From the events thus deter-
mined, draw the theoretical indicator diagrams, using 6 per cent clearance,
150 pounds initial steam pressure, 5 pounds back pressure, and 80 pounds per
inch as the scale of spring.

Find the elongation of governor spring (drawing J/ size) .
(a) From no load to normal, (6) from normal to full load.

FlG

PROBLEMS

225

76. Consider a four-valve engine, such as is shown in Figs. 60 and 61, p.
104, with head-end valves as shown in Fig. KI and K 2 , and with the follow-
ing data.

Radius of steam valve arms = 5".

Maximum diameter of steam or cut-off eccentric circle = 4".

Diameter of exhaust eccentric circle = 4".

With the crank on head-end dead center, the pivot point of the governor
arm for the cut-off eccentric is on the horizontal center line 8^" beyond
the center of the shaft. Radius of locus of eccentric centers = 7 5/16".

Cut-off at maximum load = 65 per cent. Compression = 15 per cent.

Cut-off at normal load = 25 per cent. Release = 95 per cent

va/ve in extreme open position
(Fu// /oad)

'exhaust rtt/ye !/r
extreme c/osed joosift'on

FIG. K

At normal load the steam-valve arm is vertical when the valve is in extreme
closed position. The exhaust-valve arm is vertical when that valve is in ex-
treme open position. R/L = 1/Q. Engine runs over.

Find the angle of advance for each eccentric at normal load.

Find the location of the steam-valve arm when in mid-position, at cut-
off, at admission, and when it is in extreme open position at normal load.

Find the maximum port-opening at normal load, and the lead at normal
load.

Draw the head-end steam valve in extreme position, and in open position
at normal load.

Draw the head-end exhaust valve in extreme open position.

226

ENGINES AND BOILERS

77. The necessary dimensions of a Corliss engine are given in Figs. L
and M.

Consider such an engine with the following data.

A 12"X24" Corliss engine running at 150 r.p.m.

All valves operated by one eccentric.

Engine runs over.

FIG. L

Diameter of all valves =3" (d, Fig. L).

m, Fig. L=sy 2 ";

k, Fig. L = 15".

Length of valve arms = 4' / .

Center of wrist plate is equidistant from all valves.

Radius AO and BO on wrist plate = 5".

Radius HO on wrist plate (0 =6".

Angles EC' A' and FD"B" = 90.

Release = 98 per cent.

PROBLEMS 227

Compression = 4 per cent.

Crank angle at admission = 3.

Throw of eccentric = 6f."

Radii of rocker arms are equal for eccentric and hook rods.

Normal cut-off = 20% .

Width of admission port = %".

Width of exhaust port = l".

Single-ported valves.
In Fig. M,

Radius of arm EG =3%'

Radius of arm EH =4^".

Radius of arm El =4".

Radius of cam EJ =2".

Radius of latch EK = %1".

Center G is V/i" above horizontal center line at trip position for

normal cut-off.

Make the general layout % actual size, and that of the trip mechanism
full size.

Find the lengths of the steam rod AC and the exhaust rod BD, the
angle of advance, the steam lap, the lead, the exhaust lap, the maximum

FIG. M

cut-off with trip working, the maximum cut-off when beyond control of trip,
the maximum port-opening for maximum cut-off, the maximum port-open-
ing for normal cut-off, the maximum port-opening for 10% cut-off, the move-
ment of the governor rod from normal to 10% cut-off, and the movement of
governor rod from normal to maximum trip cut-off.

228

ENGINES AND BOILERS

PROBLEMS 229

78. The necessary dimensions of a Stephenson link are as follows. (Fig. N.)
Valve-travel at full gear = 5>".
Steam lap = I".
Exhaust lap = T y.
Lead at full gear=0".
Steam port = W.
Exhaust port = 2Y 2 ".
Bridge = 1".
= l/7.5
a = b = 45".
m = 3".
e=U".

h=U".

.7 = 1713/32".
k = l8".
#==48.8".
Make the drawing one-half actual size and proceed as follows.

(1) Make a template of the link as shown in Fig. N.

(2) Locate the center of the link-block with the crank at head-end dead
center at full gear.

(3) Find the center of travel of the link-block, neglecting for the time
being the angularity of the eccentric rods.

(4) Place the center of the rocker shaft above the point found.

(5) Place the crank and eccentrics in their positions at 40% head-end cut-
off, running forward.

(6) Find by trial the position of the link (template) for the preceding
position of the crank. (Remember the center of the link-block is now at a
distance equal to the steam-lap from its mid-position.) Locate the saddle-
block pin and the position of the bell crank.

(7) Assume twelve equidistant crank positions and the corresponding ec-
centric positions for the preceding cut-off. Then draw in the center of link
for each crank position, by trial by means of the template.

(8) Draw a valve ellipse from the valve displacement found above. Locate
all events.

(9) Check as to the assumed cut-off.

(10) Find the amount of slip between the link-block and link when running
at the assumed cut-off.

230

ENGINES AND BOILERS

PROBLEMS 231

79. The arrangement of parts and the necessary dimensions of a Wal-
schaert gear are shown in Fig. O. The valve is of the piston type and has
inside admission. This is the common type used on modern locomotives.
Fig. O shows the piston in its mid-position and the link-block set at mid-gear.
As the link is pivoted to the frame of the engine at the point L, there will
be no motion of the link-block when set at mid-gear. Hence all the motion
the valve gets at this position of the block comes from the cross-head. There-
fore, as the cross-head is in mid-position, the valve will also be in mid-position.

Suppose that such a gear is used on an engine with the following data.

26J/TX30" engine.

Engine runs forward (under).

Diameter of valve = 14."

Maximum valve-travel = 6>".

Steam lap = l 1/16".

Exhaust lap = 0.

Lead =3/16".

Dimensions as in Fig. O.
Make the drawing one-fourth actual size, and proceed as follows:

(1) Lay out the gear in the position shown in Fig. O, with the piston in
mid-position and the link-block set at mid-gear. Indicate each of the parts
by its center line only.

(2) Make a template of the link.

(3) Set the valve and the point H for head-end cut-off. Then place the
crank at 40% of forward stroke, assuming that the engine is running forward.

(4) Locate the eccentric E at 90 back of the crank, and the point F, and
draw in the center line of the link.

(5) With the cross-head K in position for 40% cut-off, the location of the
point / will be determined. Since H was located in (3), the point G is found
by connecting / and H. The distance that G is to the right of the mid-
position gives the distance that the link-block center is to the right of its
mid-position. This locates the link-block for 40% cut-off. Now locate the
points D, M and B.

(6) With the link-block set for 40% head- end cut-off, take twelve equidis-
tant crank positions and find the valve displacement for each position. Plot
these valve displacements against the corresponding piston displacements
either as in a Zeuner diagram or as in a valve-ellipse diagram, and connect
the points thus found by a smooth curve.

(7) Draw in the laps on the diagram just constructed, and check the cut-off
with the assumed value of 40%.

(8) Find the amount of slip between the link-block and the link when the posi-
tion is that of 40% cut-off, with the engine running forward.

80. The Russell Engine Co. makes a four- valve engine. In this engine,
the exhaust is taken care of by oscillating or Corliss valves, and the admis-
sion by a direct slide-valve. This slide-valve admits steam and carries on
its back a rider-valve that cuts off the steam. The main valve is driven
by an eccentric keyed to the shaft, while the rider-valve is driven by an
eccentric whose angle of advance is controlled by the governor. The gov-
ernor simply rotates the eccentric about the shaft, changing the angle of

232

ENGINES AND BOILERS

advance, but affecting in no way the absolute travel of the valve. Hence,
it is necessary to consider the relative motion of the rider-valve and the main
valve in making an analysis of the cut-off valve.

The necessary dimensions of the valves and the seat are given in Fig. P.
The throw of both eccentrics is 5K inches, and the angle of advance of the
main eccentric is 32.5 degrees.

Proceed with the analysis in the following manner.

(1) Draw the two eccentric circles about the same center and locate the
extremity of the diameter of the valve circle for the main valve.

Trgyf/ grf both IM/HCS S4?

Seaf

FIG. P VALVES AND SEA^T OF RUSSELL FOUR-VALVE ENGINE

(2) Place the crank for cut-off at 25% of the head 7 end stroke, and locate
the main eccentric. Then determine from the dimensions in Fig. P how far
from the mid-position the cut-off eccentric must be to give the proper posi-
tion of rider-valve for cut-off. This determines the angle of advance of the
cut-off eccentric at this particular cut-off.

(3) Take twelve equidistant crank positions and determine the relative
position of the rider-valve to the main valve for each position. Plot these
distances as in a Zeuner diagram. It will be noticed that the diameter of
the relative valve-circle is equal in amount and parallel in directio'n to a
line connecting the extremities of the diameters of the valve circles in the
Zeuner diagrams for the main valve and the rider- valve. Also determine
the relative steam lap, which is negative. It will be found that this is equal
to the distance between the working edges of the rider-valve and the main
valve when both are in mid-position.

PROBLEMS

233

(4) Now determine the diameters of the relative valve-circles in amount
and in direction by repeating the process of (3) for eight cut-offs (0% to 70%),
and draw the locus of the extremities of the diameters of the relative valve-
circles. It is seen that this locus is the arc of a circle whose radius is equal
to the eccentricity of the rider-valve eccentric and whose center is the ex-
tremity of the diameter of the valve circle for the main valve.

(5) Make the Zeuner analysis for cut-offs of 10%, 30% and 60%, finding
the relative angle of advance and the relative valve- travel by drawing a per-
pendicular to the crank position at the point where the crank cuts the rela-
tive steam lap. The point where this perpendicular intersects the locus of
the extremities of the diameters of the relative valve-circles determines the
construction point for the relative Zeuner diagram.

81. A gravity-balanced spindle governor built as shown in Fig. Q has
arms 20 inches long. At normal speed the arms are at an angle of 45 with
the horizontal.

' (a) Find the normal speed of the governor.

(6) Find the percentage of variation in speed from no load to full load.
(c) If the normal speed is increased 30 per cent and the range of vertical
movement of the point A is the same as before, what is the percentage of varia-
tion in speed from no load to full load?

- S/o/oaJ

- formal

- - FU///04J

FIG. Q

FIG. R

82. In the cross-armed gravity-balanced spindle governor shown in Fig. R,.
the upper arms are 30 inches long and the lower arms are 20 inches long.

Find the percentage of variation in speed from no load to full load, and
compare the result with that of Problem 81.

83. The governor of Problem 81 is now loaded with a weight of 60 pounds.
If the normal speed is now 100 r.p.m., and the vertical movement of the point A
is the same as before, what is the percentage of variation of speed from normal?

84. A Corliss engine is governed by a loaded gravity-balanced spindle
governor. The pulley on the governor and pulley on the engine are both
10 inches in diameter. It is desired to change the speed of the engine from
100 to 120 r.p.m. In what three ways may this be done without affecting
the speed regulation? Give your calculations.

234

ENGINES AND BOILERS

86. Find the percentage of variation in speed from no load to full load for
the governor shown in Fig. S.

86. Suppose that the spring of a spring-balanced centrifugal governor is
fastened directly to a weight of 40 pounds, as shown in Fig. T. If the speed

FIG. S

FIG. T

at no load is 206 r.p.m., and at full load 200 r.p.m., what must be the scale
of spring?

87. If the spring in Problem 86 is replaced by one of 40 pound scale, and
the speed at full load is 200 r.p.m., what is the speed at no load? Is the
governor stable or unstable?

88. What scale of spring would make the governor of Problem 86 iso-
chronous at a speed of 200 r.p.m.?

89. The no-load speed of a governor is 300 r.p.m. and the full load speed

is 290 r.p.m. At no load, the tension
in the governor spring is 500 pounds.
At full load, the spring is 2 inches shorter
than at no load. The scale of spring is
100 pounds.

If the spring is tightened by shorten-
ing it up half an inch, what will be the
effect on the no-load and full-load speeds?

90. With the governor of Problem 89,
how much would the spring have to be
taken up to make the governor isochro-
nous? What would the speed then be?

91. In the steam-turbine governor
shown in Fig. U, the weights of 8 pounds
each are 4 inches from the center of the
spindle at full load when the speed is 600
r.p.m. At no load, the weights are 5 inches

FIG. U

from the spindle and the speed is 610 r.p.m. The weights are pivoted at
points A and A .

(a) Compute the scale of spring.

(6) Design the spring, i.e. find the size of spring wire, the diameter of the
coil, and the number of turns that will give the correct force and scale.

M512399 TJ255
B9
Forestry