oPPLEMENTY
MATER
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INCRUASING THE CYLINDE: PE PORMMAN CE
OF THE GaS ENGILivis
Thesis for Degree of M. Ee
Harold Madison Jacklin.
1919
(
.
N.
1. During the past few years the demand for
gasoline has increased far more rapidly than the supply of raw
material from which it is produced. Consequently, there has been
@ general change in refinery practice with the production of
80-called heavy gasolines, which are actually of relatively low
volatility. Although this change in quality is not necessarily
deterioration, it has brought to the attention of designers ani
users of internal combustion engines, problems that did not
exist in the past. The main problem has been that of carburetion.
Other troubles, such as the rapid format ion of carbon and the
dilution of the lubricating oil by the less volatile parts of the
fuel are the result of poor carburetion or rather incomplete
vaporization of the fuel.

The fuels marketed at present as gasoline are produced
in various ways and are as follows:

l. Straight refinery gasoline

Ze Oracked gasoline

3- Casing head gasoline blended with kerosene

4. Casing head gasoline blended with heavy straight

refinery gasoline

5. Casing head gasoline blended with naptha
Re . These fuels can be used for starting gas engines
under practically all conditions as about twenty per cent of the
fuel will vaporise at temperatures below 100 or 200° F. If
means were provided for completely vaporizing the fuel soon after
the engine gets started, they should give good service. The
difficulty lies in the fact that in order to vaporize the least
volatile part of the fuel and have a “dry” mixture enter the

the engine, temperatures ranging from 320 to 550° F. are necessary.
L002 27
e
This temperature is commonly called the "end point” of the

fuel. In order to properly vaporize these fuels, the mixture
of gasoline vapor and air should enter the cylinder ata
temperature near if not at that of the end point of the fuel
being uged. Few carburetors or inlet manifolds give any such
temperature. This is evidenced by the fact that passenger cars
require changing of the lubricating oils in the crank case at
least once a month, andi trucks about every other week, because
of the excessive dilution of the oil and consequent lubrication.
The remedy, of course, is to heat the mixture more highly. In
the case of automobiles, the economy is increased somewhat by
by doing this but there is a falling off in maximum or full
load power of the engine. It should be remembered that this
discussion relates to full load conditions in all cases.

H. L.- Horning of the Waukesha Motor Co. stated that "It is
conservative to estimate a reduction in horsepower of 10 to

20 per cent because of such heating, mt economy of fuel is
increased considerably" in the March 22, 1917 issue of"®he
Automobile", now called “Automotive Industries." It is evident
that heating the mixture ent ering &@ gas engine cylinder will
reduce the volumetric efficiency. There are various proposed
definitions for volumetric efficiency. In this paper, it will
be considered as the ratio of volume of air and fuel drawn into
the cylinder at atmospheric pressure and 62°F. to the "suction
volume.” In the case of present engines, suction volume is
represented by the peton displacement. Later on it will become
clear why this term is used in place of the term "piston
displacement."

Be In the average case, it is found that the mixture
in the cylinder of present engines at the beginning of the
compression stroke is at a temperature of 390° C abs. This
figure is the result of investigations by three competent
authorities quoted by O. Ae Malychevitch in the September 19
amd 26, 1918 issues of “Automotive Industries." This is
equival@mt to a temperature of 702° F. absolute. This temperature
is the temperatwe resulting when the new cool charge is
heated by coming in contact with the cylinder walls am by
mixing with the hot gases left in the clearance space at the
end of the previous exhaust stroke. If it be considered that
the charge entering the cylinder when these temperatures are
obtained is at a temperature of 100° F.; and that, in order
to vaporize present fuels, a temperature of 400° F. is needed,
less mixture will be drawn in. The ratio of the amounts of
mixture (by weight) will be equal to the ratio between the
absolute temperatures of the cylinder contents at the beginning
of the compression stroke. fThis ratio is found as follows.
4. With the charge m@mtering the cylinder at a
temperature of 100° F. with grades of fuel as used in the past,
the temperature of the cylinder contents becomes 702° F. abs
at the beginning of the compression stroke. The temperature
of the charge has been increased 702 - (460 + 100) = 142° BF.
5. : The mixture entering the cylinder at 400° F.
will have the temperature raised the same emount, thus its
temperature at the beginning of the compression stroke will be
460 + 400 + 142 = 1002° F. abs. The ratio between the absolute
temperatures of the charges is 702 . Therefore 702 x 100 = .69%
1602 — I00E
per cent as much mixture enters the cylinder when the charge is

heated. This charge will contain only 69% per cent as much fuel,
and, therefore, the full load power of the engine will be
reduced by about 30 per cent simply by pre-heating the charge as
shown.
6. Therefore, the proper use of present fuels means
that an engine of a given sise will give considerably less
horsepower. Nor is there any relief in sight as regards fuels.
There are no great unfieveloped fields that can be drawn on, and
even if there were, it is a question as to whether or not it
would be advisable to change the fuel back to that which was used
in the past. National economy of resources seems to demand that
the present grades be used. In fact, the U. 8. Bureau of Mines
is recommending a so-called "straight run" fuel which shall include
the present grades of gasoline and a good per cent of the present
kerosene and possibly some distillate. Present indications are
that this will gradually come about in the relatively near future.
Temperatures of the mixture will then have to be such that the
temperature of the gases at the beginning of the compression
stroke will be held at 400° F, at least, if not 500° F. in order
to insure good operation. Evidences of what this will mean have
already come to light in the case of a great many tractors that
could not be made to pull their rated load on kerosene, the owner
being forced to use gasoline in order to do so. It has been found
in many cases that an engine delivered approximately three-fourths
as much power on kerosene as it did on gasoline. In other words,
a given size cylinder produces less horsepower. The weight in
an ordinary commercial engine cannot be reduced proportionately,
since many of the dimensions of these engines are determined by
considerations having to do with foundry practice rather than

strength. Then too, gasoline is used in order to warm up an
engine before attempting to use kerosene, and the engine

should be designed to take care of the higher explosions incident
to operating with the cool mixture. It is probable, also, that
higher explosions wuld result in the engine after being warmed
up in case it missed a few times. The cylinder would then be
filled completely with a combustible charge instead of one
diluted with exhaust gases so a higher explosion pressure would
have to be cared for in the desig. Suppose that a given engine
for mtomotive use weighs sixteen pounds for each horsepower
developed when operating with a relatively cool mixture. Certain
parts, such as the connecting rod, crank shaft and possibly the
flywheel might be made lighter. Perhaps one pound could be
removed in this way figuring on the output being the same as at
present. The engine would then weigh fifteen pounds per horse-
power developed as rated on the cool mixture, but whem operating
on the heated mixture the actual weight per horsepower developed
would be close to twenty pounds per horsepower developed if the
loss be considered as 25 per cent. Therefore, it is probable
that the “weight of engine per horsepower developed” has
increased and that it will increase further unless means other than
the makeshift of increasing the speed is provided to counteract
the tendency. In other words, some means should be provided to
cause the output from a given cylinder to be increased and, if
possible, to bring about conditions whereby the parts are
subjected to the maximum strains at all times instead of only

at the time the engine is being warmed up or when it misses once
in a while, as in the case of a hit-and-miss stationary engine.
7. By increasing the compression pressure to the

limit for the fuel, the power output will be increased somewhat
but the other conditions remain. The volumetric efficiency

is increased by the use of higher compression pressures, but in
any case, the charge consists of a goodly percentage of burned
gases from the previous cycle mixed with the new charge, so that
the explosion pressures while running are lower than when starting
or after one or two "misses." It is found in the average case
that the volumetric efficiency is seldom over 75 per cent and
that 65 per cent to 710 per cent is more often obtained.

8. By using water injection, the compression can
be carried above that permissible without. MThere are cases where
it has been carried as high as 150 pounds per square inch with
gasoline as the fuel. This is at least 60 pounds per square

inch above that at which pre-ignition is a danger without the
use of water, even in a refined design. The use of water would
improve the volumetric efficiency a great deal, but it should

not be resorted to. Few users are expert enough nor are they
willing to care for the necessary adjustments in most cases.

On automobiles, trucks, tractors and airplanes, it is practically
impossible to use water as it adds considerably to the weight

of the vehicle even if it were always easily procurable in
sufficient quantities.

9. However, the author believes that the real
remedy lies in increasing the volumetric efficiency, but by
means other than increasing the compresai on pressure which
entaile the chance of trouble from pre-ignition in case of the
slightest trouble in the cooling system, or by the use of water.
In 1913, the Automobile Club of America made some rather
extensive tests on a Moline Knight four-cylinder automobile

A
engine. The cylinder dimensions of this engine were 4 bore and
6" stroke. Among the tests were some made to ascertain the
volumetric efficiency. The air entering the carburetor was
measured by means of a venturi meter from which the air flowed
to an air box surrounding the carburetor. Various readings
were taken as indicated in the following table. Tests were also
made with the carburetor on but with the engine driven by a
dynamometer, and also without the carburetor, the engine being
driven in the same manner. Since we are concerned with the
actual performance, i.e. the operation with the carburetor on
and the engine running under its own power, these latter two
tests are not shown.

Volumetric efficiency of Moline Knight 4 x 6 automobile
engine. Data'taken from Table 4 of the report of the Automobile
Club of America test made in December 1913.

Test Duration R.P.M. B.H.P. Vol. eff.
A, 6 min. 1715 52.5 70.5
By 5 1523 51.1 7164
C., ‘5 1318 45.1 75.9
De 5 1108 38 62 78.0
Bs 5 902 30.8 79.0
Fz 4 701 25.1 78.6
Ge 5 491 16.8 72.9
Hg 5 295 9.5 51.0
10. The volumetric efficiency of poppett valve

engines of the automobile type is generally much lower than that
of the Knight engine because the valve openings are smaller and
less direct than those in the latter type. Where particular
care has been taken in laying out the cylinder and valve
arrangements, higher volumetric efficiencies do obtain as

in the case of modern airplane motors. Since noise from valve
parts is not objectionable in these engines, quick opening and
closing cams may be used, thus making the valves more efficient.
Then too, the cooling facilities are much better, in that the
water circulation is much better. The cylinder is generally

of steel, consequently it may be relatively thin, so that the
inner surface is not as hot as the inner surface does not have
as great an effect in heating the incoming charge from the
carburetor. However, the volumetric efficiency even under these
conditions cannot be conceivably greater than 85 per cent.

li. Since the tests on the Moline Knight were made
over five years ago when a good quality of fuel was easily
procurable, it was unnecessary to pre-heat the charge to insure
vaporisation of the fuel. It is probable that, if tested today,
the figures would be somewhat different because pre-heating

is absolutely necessary unless some method of direct injection
of the fuel is used. It is interesting at this point to consider
the probable increase in power of the Moline Knight if the
volumetric efficiency were 100 per cent. From test Bs, it is
geen that the volumetric efficiency is 71.4 per cent at 1523 R.P.Me
The probable increase in power is then 100 - 71.4 , 100 = 40%.
12. Taking 85 per cent as the volumetric efficiency
of the best aviation engine, the probable increase in power
because of increasing the volumetric efficiency to 100 per cent is
100 ~ 85 x 100 8 17,67%.

— a

13. From the last two paragraphs, it is apparent
that there is a great deal of lee-way in which to set to work

to fit engines with new accessories so that more mixture is forced
into the engine, or to redesign engines so that they may draw
in a greater charge during their intake stroke. As it stands at
present, approximately one-fourth of the intake stroke of the
piston is used in expanding the gases left within the clearance
space at the end of the exhaust stroke and is therefore ineffective
in drawing in a new charge. Air scavenging, precompression, and
the author's design (which incorporates an auxiliary piston in
the cylinder) are ways in which this condition may be partially
if not entirely corrected.
14. Air scavenging might be worked out as follows;
A third valve could be arranged in the cylinder in such a position
that air could be pumped through it into the cylinder, blowing
the products of combustion from the previous cycle out through
the exhaust valve, thus leaving the clearance space filled with
air instead of with a spent or burned gases. A rich mixture
could then be drawn into the cylinder from the carburetor to
mix with this air and to form a correctly proportioned mixture
for the next explosion. A possible valve timing that suggests
itself is-

Exhaust opens 45° to 55° before outer center

Air valve " 45° to 20° before inner center

Exhaust closes 15° to 25° after inner center

Air valve " 15° to 25° after inner center

Inlet opens 18° to 30°

after innér center

Inlet closes 30° to 45° after outer center

The auxiliary air pump or blower would need to have a
capacity of at least double the clearance volume in order to insure
through scavenging. This means that its capacity would be 40 per

cent to 60 per cent as great as the paton displacement of the
oe

s
LO
engine itself. On the face of it, it appears that the increase

in weight of the complete power plant would counterbalance the
increase in power, to eay nothing of the extra complication

and attention necessary, due to the pump and extra valve parts.
This type, therefore, will probably not come into any great
UBC.

15. A pump or blower with a capacity greater than
the piston displacemamt of the engine might be arranged to
pre-compress the charge. Thug more mixture would be forced

into the cylinder. Because of the possibility of loss of fuel
from condensation, @ pump, if used, wold probably be arranged
to force the air through the carburetor. The carburetor bowl
would have to be sealed and connected to the inlet pipe else
there would be difficulty in procuring fuel. On the other hand,
@ blower would assist in breaking up the fuel and could be
placed between the carburetor and engine to good advantage on
this account. The pump would add greatly to the weight of the
power plant and would not be used on that account. It is
conceivable that the blower might be a drag on the engine rather
than an aid to it at low speeds. It is also a question whether
or not the blower could be made so efficient that there would
be any net gain.

16. The author has worked out a design which he
believes will solve the problem with a minimum of complication.
In thie paper, he will show his design of a 5” bore by 7. stroke
six cylinder aviation engine using sliding sleeve valves to
control the gases in a manner somewhat similar to the Knight
engine which most engineers are now familiar with. As already

stated, an auxiliary piston is also used. This piston is
ll
operated by an eccentric or crank which rotates at one-half
the speed of the crank shaft. The stroke of this piston
is so proportioned that the clearance between it and the main
piston is very small at the end of the latter's exhmst stroke;
and so that the two are the proper distance apart at the end
of the compression stroke to give a clearance volume that
insures the desired compression pressure. Figure I shows the
arrangenent of the piston and sleeves at the left, the
auxiliary piston taking the place of the cylinder head. At
the right, the distance between the two pistons is show
throughout the cycle of operation. It is seen that the
auxiliary or head piston is in its outer position at the time
of explosion and that it is in its inner dead center position
when the main piston has returned on its exhaust stroke. A
variable clearance throughout the cycle is the result. ‘The
head piston assists in clearing the cylinder of the burned
gases during the exhaust stroke and assists in drawing in the
new charge during the suction stroke of the main piston. Since
the clearance volume is very small at the end of the exhaust
stroke, a very small quanitty of hot exhaust gases will be left
in the cylinder with which the new charge will be diluted.
Therefore, the new charge will not have its temperature raised
appreciably by mingling with these spent gases; the charge in the
cylinder at the end of the motion stroke will be cooler so
that a greater weight of mixture will be handled by the engine.
Since the cylinder contents at the time of ignition will always
be made up of a much greater proportion of new unburned gases
than spent gases, ignition will always result even when the
engine is operated with the throttle nearly closed.
 

 

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17. It should be clear from Figure I why the

author uses the term "suction volume” in his definition of
volumetric efficiency. The suction volume in this case is
somewhat greater than the piston displacement of an ordinary
engine of the same bore and etroke because of the action of the
head piston.

18. In order to compare this new design with
existing engines of about the same size, data has been selected
from descriptions of three of the best German designs used

in their airplanes between 1914 and 1918. From this data, a
card is constructed and constants derived which may be used

in constructing acard for the new type. The analysis of the

three types is presented first.
ANALYSIS OF THREE GERMAN AVIATION ENGINES.

19. The three engines to be examined are the Benz
six cylinder 160 H.P., and the Benz six cylinder 230 H.P., and
the Mercedes six cylinder 260 H.P. From the general data and
test results of these engines as published in "Automotive
Industries” for January 24, 1918; February 14 and 28 and March

7, 1918; and September 6 and 13 ami October 4, 1917 respectively,

the following items are selected.

Beng Beng Mercedes "Av. Engine*
1. Rating-H.P. ~ 160 230 — ~ B60. 3 °
3. Stroke-inches 7.09 7.48 7.09

4. Brake mean effective
pressure 103.4 113 107.5 107.96

5. Fuel c onsumption lbs.
per H.P.-HR 92.5 135 140.63 122.71
13
7. Compression ratio 4.5 5200 4.94 4.813

8. Piston displacement
for one cylinder-

cu. inches 146.05 191.4 220.82 186.1
9. Mechanical efficiency ~875 «85 ° 85 858
20. Since these engines operate with a free exhaust

and are of a rather refined design, the following assumptions
are believed reasonable.

l. Pressure within cylinder at the end of the exhaust
stroke ~- 15.5 lbs. per square inch absolute.

2. Temperature of the gases remaining in the cylimier
at the em of the exhaust stroke - 740° F. absolute.

3. Temperature of the air entering the carburetor -
517° F. absolute.

4. Ratio by weight of air to gas 14 to l.

5. Atmospheric pressure - 14.7 lbs. per square inch absolute.

el. With an air-gas ratio of 14 to l, the average

engine as shown in paragraph 19, would require 122.71 x 14 =

28.62 lbs of air per minute. °°

226 This air would occupy 28.62 = 376.7 cu. ft. at
6.0%6

14.7 lbs per square inch and 62° P., the factor (0.076) being the
weight of one cu. ft. of air under these conditions.
23- The volume of the fuel vapor used in one minute
under these conditions is 122.71 x 4.2 = 8.58 cu. ft., the
| ~ 6

factor (4.2) being the value given by A. M. Levin on page 125
of his book on the Modern Gas Engine (Wiley and Son, 1912)
for the volume of one pound of gasoline vapor.
24. The total volume of air-gas mixture drawn
into the cylinder of the “average engine" per minute is

376.7 + 8-58 = 385.28 cu. ft.
14
25. The piston displacement of the "average engine"

per minute at 1400 R.P.M. is 3 x 1400 x 186.1 = 452 ou. ft.
ies

266 The volumetric efficiency of this engine is then
By = 385.28 = - 853

~ “455 |

27 In order to construct a complete indicator

card from which constants may be derived for this "average

engine”, it is necessary to find’ $ressure and temperature of the

cylinder contents at the beginning of the compression stroke.
The method used by J. R. Du Priest on pages 385 to 387

inclusive of the Trans, A.S.M.E., Vol. 39, to find the pressure

works out as follows for this case.

28. The clearance volume of the “average engine"

is 186.1 - 48.85 cu. in. This is shown on Fig. 2. This

4.613 -1

volume is filled with exhaust gases at a pressure of 15.5 lbs.

per square inch and a temperature of 740° F. absolute.

296 These gases, whem expanded isothermically,

occupy a x 48.85 = 51.5 cm. in. at 14.7 lbs. per square inch.

30. , The volume of air-gas mixture drawn in is (from

paragraphs 19 and 26) .853 x 186.1 = 158.8 ou. in.

Sl. The total volume of exhaust gases and air-gas

mixture at 14.7 lbs. per square inch to be expanded to occupy

the whole cylinder volume ig 51.5 # 158.8 = 210.3 cu. in.

32. The volume which these gases occupy at some

lower pressure is the sum of the piston displacement and the

Clearance volume, and is 186.1 #4 48.85 - 234.95 cu. in.

335-6 Then P,, the pressure at the beginning of the

compression stroke ia P, = 14.7 /210.3 )9°8 - 13,48 lbs. per

234.95)
square inch absolute. The exponent (0.85) was arrived at as a

result of experiment by Mr. Du Priest as shown on pages 389-390
15
Trans. A.S.M.E., Vol. 39.
34. In finding the temperature of the cylinder
contents before compression starts, it is assumed that the
specific heat of the exhaust gases is the same as that of
the new aid-gas mixture in order to simplify the computations
and because the error introduced is negligable in the final
results. The temperature of the air as it reaches the
carburetor is 57° P. (517° F. absolute). There is a drop of
approximately 40° when this air-gas mixture is expanded to the
pressure within the cylinder and the fuel has been vaporised.
Upon entering the cylinder, the air-gas mixture encounters the
hot cylinder walls with a resulting rise in temperature of 55°.
Thus the temperature of the air-gas mixture before it has
mixed with the hot exhaust gases is 517 - 40 + 55 - 532° F. abs.

35. The weight of the air-gas mixture is
M, = 158.8 x 0.076 = .00698 lbs.
1728
36. The weight of the exhaust gases is

M - 48.85 x 522 x 15.5 x 0.076 = 20016 lbs.
ro" [7eB 40 4.7

37. Then 0.00698 lbs. of air-gas mixture at 532° gf,
are raised to a temperature, T, , by mixing with 0.0016 lbs. of
exhaust gases at 740° BF. abs. With equal specific heats, the
equation becomes (Tg - 532) 0.00698 = (740 - T,) 0.0016 from
Which Ta = 572° F. abs.

38. The above results for P, and Te are checked as
follows. The air-gas mixture occupies 158.8 cu. in. at 14.7 lbs.
and 522° F. At 13.48 lbs. and 572° F, it will occupy

572 x 14.7 x 158.8 = 189.9 cu. in. The exhaust gases remaining

B22 13.48
in the cylinder occupying 48.85 cu. in. at 15.5 lbs. ami 740° BF,
16
At 13.48 lbs. and 572° B. they occupy
572 x 15.5 x 48.85 - 44.5 cu. in. The total volume occupied by
740 13.46
the air-gas ani exhaust gas mixture is 189.9 + 44.5 = 234.4 cu. in.
which corresponds very closely to the volume they are supposed
to occupy - 234.95 cu. in.
9. Having obtained the values of P, and Ta» it is
now possible to construct the compression line as shown in Fig. II,

as well as to compute the temperatures as shown by the dashed

line. The volumes at the various points are;

Va = 234.95 cu. in. VY, = 110.90
Vp = 203.94 Ve = 79.88
Vo = 172.92 V, 3 48-85
Va = 141.92

The corresponding pressures in lbs. per square inch absolute,

using N «1.33 in the equation py = constant, are

Py = 16.20 Pp = 55.6
Po = 20.20 Py = 108.2

The corresponding temperatures in degrees Fahrenheit absolute,

using the equation PV n-1 = constant, are

Ta = 572 To = 742
tT) = 607 T, = 824
1), = 640 Te » 968
Ta = 675
40. The explosion pressure Py, Fig. 2 is derived by

means of Pp. M. Heldt's formula as stated in the Horseless Age,

page 1 of the engineering section, December 15, 1916. It is as

follows: M.E.P. - m.e.p. ie (a-1) (_2_) (2 ~ (1) °°}
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in which

M.E.P. 2 mean pressure during expansion

MeOGeDe = " " " compression
P. =» canpression pressure in lbs. per square inch
& -
- 108.2
a- ratio in which P, is multiplies on explosion

&

r volumetriés compression ratio = 4.813

It is apparent that P, (a - 1) is the rise in pressure due to

the explosion. It remains to derive a value for the mean
effective pressure (M.E.P. - meeep.) for this "average" engine
before the equation can be solved for the value of P, (a - 1).

41. The brake mean effective pressure is 107.96

lbs. per square inch and the mechanical efficiency is 0.858.

From these values M.E.P. - me@ep.e = 107.96 w 125.8 lbs. per sq. in.

~ 858
42. Then P. (a - 1) s 125.8 x 023

1 Ll .
(g—piger) (2-ta yg 1°]

The denominator of the second number becomes; + y(r.-l y. i
813 1.678 ~ 9.42

Then  . (a - 1) s 125.8 x 0.3 x 9.42 w=» 374 lbs. and the explosion

 

pressure Py 2 108.2 x 374 = 482 lbs. per square in. abs.

43. The temperature at explosion is T, = ox x Tp =
482 x 968 ~ 4300° F. abs. é

108.2

44, The pressure during the expansion stroke,

using n =» 1.33 in equation PV = constant, were then calculated
and plotted as shown by the dfted line in Fig. 2. Upon checking
the area, and the m.e.p., it was found necessary to plot this
line from some higher point as Pyl. In this case, it was found
that Px should be 10 per cent greater than Px. Allowance is

made for the increase in pressure due to ignition occurring early,

for a rounding of the curve at its highest point and for the drop
18
in pressure when the exhaust opens.

Pyl = 1.10 Py = 530 lbs. per square in. abs. The curve
resulting is shown in the full line Fig. 2. The difference in
area of the large and small loops is 7.565 sqare inches. The
card ig 6" long and the pressure scale is 1” z 100 lbs., so the
Ms OeD. figures out to 126.08 lbs. which is very close to the

value desired. THE pressures at the various points are-

P+ = 530 Py # 99.0
py? - 278 P,t = 7906
P,” = 182 Pa) = 66.0
Pa = 130
and the temperatures are
Tx = 4300 To! 2 2838
Te) = 3630 T > ~ 2700
To. = 3242 Ta! = 2578
t,) = 30%
46, From paragraphs 35 and 36, it is apparent that

the total weight of the cylinder contents at the time of
explosion is M a 00698 # -0016 ws .00858 lbs. From paragraphs
39 ani 43, it is seen that the temperature rise is

Ty, - Te z 4300 - 968 2 3332. The heating value of the fuel

in each charge ig Q 2 122.71 x 19000 =z 9.2 B.T.U.
60 4200

122.71 = weight of gasoline comsumed per hour
19000 «» B.T.U. per lb. of fuel

4200 number of inlet strokes per minute.

The above values may be substituted in the equation
Q = Mcy (Tx - Te), ami a value deduced for oy. hen

Cy a 9.2 = e322
 
19

46. The writer is well aware that a card of this
type igs not usually obtained from an engine operating at 1400
R.P.M. It is probable that the explosion line would be inclined
toward the right instead of being vertical, and that the
pressure at the time the exhaust valve opens would be higher
than that shown. However, this more or less ideal card has been
made to fit test results by using a correction fadpr, so that
@ value Hr the specific heat may be derived. It should be
noted that the figure for the specific heat is approximately
double that accepted for air. This agrees with observed results,
where about one-half the expected rise in preqsure has been
obtained if the specific heat for air has been used in computing
the expected pressure. By applying the factors as derived for
this average engine to another engine operating under similar
conditions, the performance ofthe latter may be quite accurately
prophesied. In working out the data for this engine the
following data will be used.

l. Pressure within cylinder at the em of the exhaust
stroke = 15.5 lbs. per square inch absolute

2. Temperature of above gases = 740° F. abs.

3. Temperature of air entering carburetor - 517° B. abs.

4. Atmospheric pressure 14.7 lbs.
14 to l.

5. Ratio of air to gas by weight -
6. Je Re Du Priest's exponent = 0.85
7. Weight of air at 62° and 14.7 lbs. per square inch =
0.076 lbs. per cu. ft.
8. Factar with which to multiply the explosion pressure

to find point from which to plot the expansion line » Px! » 1.10
£0
9. Specific heat of cylinder contents at time of
explosion = Gy = .522

10. Pressure in cylinder at beginning of compression
atroke = 135.5 lbs. per sqtare inch absolute —

ll. Drop in temperature of air-gas mixture due to drop
in pressure and vaporizing of fuel = 40° F. .

12. Increase in temperature of air-gas mixture due. to
coming in contact with cylinder walls = 50° F. Thies is slightly
less than that used in the analysis of the "average engine."

It is deemed proper because the surface with which the charge
comes in contact is actually less per unit of weight in this
new engine than in the “average engine™ and becanse there is
more air-gas mixture entering the cylimier to be heated by such
contact. Using as accurate comparative figures as possible, it
was found that this rise in temperature would be 46° F. The

writer has compromised on 50° F. as an acceptable figure.

The Jacklin Engine
47. The arrangement of the two pistons and sleeves
in the cylinder of this engine is shown in Fig. 4 (large print
in pocket on back cover), while the displacement curves through
720° of crank shaft rotation for the various parts are shown in
Fig. 3. The operation of the sleeves and auxiliary or head
piston in timing the valves and in varying the clearance volume
is clearly shown. By observing the intersections of the curved
lines with the vertical lines representing the degrees of
rotation of the crank, the position of the parts at any instant
may be accurately found. All necessary design data as to

centers, etc., is noted on this drawing so that no trouble should
el
be experienced in constructing a lay-out.
48. It is seen that the clearance volume is very
small at the end of the exhaust stroke, as represented by V,.
The distance between the faces of the two pistons at this time
is 1/16", which is ample to care for all differences of expansion
of the various parts due to heating. The volume between the
pistons at this time is V.. "1 x area of 5" circle + the volume
X which is a shallow ring s1 x 19.655 ¢ .43 » 1.86 cu. in.
49. The volume of these gases expanded isothermically

to 14.7 lbs. per square inch is 15.6 x 1.86 a» 1.96 cu. in.

50. Since such a small quantity of hot gases would

have a negligable effect in heating the new charge, these gases

should also be considered as cooled to the same temperature as

the cylinder contents at the emi of the suction stroke, which is

T, - 517 - 40 + 50 = 527° F. abs. so that the exhast gases occupy

527 x 1.96 = 1.396 cu. in. at 14.7 lbs. per square inch ami 527°

740 .

F. abs.

5l. The volume between the pistons at the em of

the suction stroke ie V, = 7 7/8 x area of 5" circle - .43 cu. in.

- 7.875 x 19.635 ¢ .43 » 155.26 cu. in.

52. With (Y) the volume of newair-gas mixture

drawn into the cylinder as outlined in paragraphs 27 to 33

inclusive, the following equation is true 13.5 = 14.7 (¥_ + 1.396)9-95
“155.26

from which Y = 139.6 cu. in. air-gas mixture at atmospheric

pressure and temperature.

53. On the basis of “suction displacement" the
volumetric efficiency of the Jacklin engine is Ey = 139.6 =-91
155.26 - 1.86

54. A compurison of the volumetric efficiency
Le
of the Jacklin engine and a conventional engine of the same
bore and stroke (in the case of the Jacklin, the bore and stroke
of the main piston) is interesting at this point. Witha
5 x 7 cylinder, the piston displacement is 137.45 cu. in. Then,
the volumetric efficiency of the Jacklin engine on this basis is
zt
100(1.016 eee) 19.1 per cent more aig-gas mixture than a

conventional engine or, in this case, the “average engine", since

s 1359.6 = 1.016. Therefore, this engine will handle
137.45

853 is the volumetric efficiency of the “average engine." Since,
in general, the power developed depends on the amount of fuel
consumed in the engine, it seems to be reasonable to expect

& large increase in power when this design is used.

55. It remains to figure out an indicator card to
complete the analysis and comparison. The volume in cubic inches
between the pistons at the various points in the stroke of the
‘main piston are (referring to Mig. 3).

Vy = 155.26

Ve s 7.1875 x 19.635 + .43 » 141.43

124.33

105.93

Ve = 6.3125 x 19.635 + .432

a8

V4 = 5.3750 x 19.635 + 043

Va 2 4.4380 x 19.635 + .43 w 87.53
Vo = 3.5000 x 19.635 + .45 w 69.13
Vz = 2.5313 x 19.635 + .43 w 650.13
Vo = 1.5625 x 19.635 + .43 = 30.77
The corresponding pressures in lbs. per square inch. abs. at these

points (using n = 1.33 in the equation PV" = constant) will be

Py = 15.5 P, = £2.8 Py z 61.1
23
The temperature at the end of the compresaion stroke will be
T. = Ty x(Vn)9°5S | 527 x, 155.26 ,0-33 - 901° B. abs.
0 TT! ( ayy)

This temperature is lower than the temperature Tos paragraph

359, so that there is no likelihood of pre-ignition occurring

in spite of the fact that the compression pressure has been

increased about 10 lbs. per square inch.

56, At 1525 R.P.M. this engine would use

4575 x 139.6 = 370 cu. ft. of air-gas mixture at 62° F. and 14.7

lbs. per square in. absolute. The factor 4575 is the number of

suction strokes in a six cylinder four-stroke engine at 15265

R.P.M. Then the weight of the air-gas mixture (assuming the

fuel vapor weighs the same as the air which is reasonably

accurate as the vapor occupies approximately but 2 per cent of

the volume) is 370 x 0.076 = 28.15 lbs.

57. When a 14 to 1 ratio of air to gas is used,

this mixture will contain 28.15 e 1.875 lbs. of fuel, the
i4-1-

heating value of which is 1.875 x 19000 = 35,610 B.T.U.

58. The total weight of the cylinder contents is

28.15 lbs. plus the weight of the spent gases left in the

cylinder at the end of the exhaust stroke. These weigh

4575 x 1.396 x 0.076 - -281 lbs. so the total weight of the

cylinder contents becomes M = 28.15 + .281 = 28.43 lbs.

59. All values except the explosion temperature

(p07) in the equation Q = Mcy ( Tot - T,) are now available,

so that 25,610 = 28.43 x .322 (7,1 ~ 901) from which

t+ = 4916° F. abs.

60. Then the explosion pressure P,? =ToX x Po =

4916 x 117.8 = 641 lbs. per square inch absolute.
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A:

6le Po” the point from which the expansion line

showld be drawn is P," 21.10 x Py’ = 1.10 x 641 =» 704 lbs.
pressures

626 Whe/in lbs. per square iuch abs. during the

expansion stroke at the various points will be

Po" = 704 Py" = 136.2
Pi" = 367 Po" = 112.7
P,." = £39.3 Pe" = 935.1
Ps" = 175.3 Py" = 88.8
63. Fig. 5 shows the compression am expansion

curves as plotted from paragraphs 55 and 62, with allowances

made for the rise in pressure at the time ignition occurs and for
the drop in pressure after the exhaust valve opens. The
difference in area of the upper and lower loops is 11.33 square
inches. The mean effective pressure is then 161.5 lbs. per

square inch.

64. Since the head piston moves outward and

inward #" during the compression and expansion strokes,
respectively, the effective compression and expansion stroke is

7 - & = 64 inohes.

656 Assuming the mechanical efficiency of this

engine at 0.80 instead of .853, which is considered a liberal
allowance for the increased power necessary to operate the valves,
the brake mean effective pressure becomes B.M.E.P. = .80 x 161.5 ss

129.2 lbs. per aquare inch.

 

66. The brake horsepower to be expected from this
engine is then B.H.P..129.2 x 6.5 © 19.635 x 1625 x 3 = 190.7.

2 x 55000
67. Using the B.M.E.P. of the “average engine," the

brake horsepower of a conventional six-cylinder 5 x 7 engine at
25
1525 R.P.M. is B.H.P. = 107.96 x 7 x 19.635 x 1525 x 3 = 171

12 X FB00S0
68. The increase in horsepower by the use of the
Jacklin design is 190.7 =~ 171 = 19.7 H.P. The per cent
increase is 100( 19-7) = 10.32%.
69. A final comparison is not possible unless a
complete design is worked out so that the weight may be
estimated. Figs. 6, 7 and 8 show three views of a5 x 7, six-
cylinder engine according to this design. In the following
paragraphs, the various features of the design are described,
followed by a table showing the specifications of this engine
arranged parallel to those of the three engines referred to in
paragraph 19.
70. The piston is an aluminum casting of a very
heavy section as is apparent in Figs. 4, 6 and 7. The center of
the head is relatively thin and becomes thicker toward the edge,
from which point the metal is tapered until at the lower edge of
the skirt it is very thin. This type was decided upon because
of the advantages it presents. The heat from the piston head is
conducted down slong the skirt and is communicated to the
surrounding sleeve. Also no ribs are necessary and a cleaner
casting should result. The rings are placed a considerable
distance from the head in order that they may not expand into
the parts in the inner sleeve. They are of the three piece type.
The bearing for the piston pin is the aluminum casting itself.
This is a good bearing material, and in addition the weight of
bushings is saved. The head piston is also an aluminum casting.
The piston pin bearings are exactly the same as those for the
Main piston. Since there is a relatively long connecting rod

(over 7 inches), while the stroke is only 14+ inches, it is
 
£6
permissible to make this piston very short as has been done.
The wide junk ring is pinned so that its joint comes between
the two ports in the inner sleeve. The triple ring just above the
junk ring should prevent all possibilities of leak at that point.
As is apparent in Figs. 6 and 7, the head piston is water cooled,
it being cast with a double head. This is deemed necessary
because the spark plugs are carried in the head piston. The
water connections and spark plug connections are very simple
and should not give any difficulty because of the shortness
of the stroke. The weight of the main piston with rings and
pin is 4.5 lbs; while that of the head piston is 4.5 lbs.
71. The piston pins are 1 3/8 inches in diameter
and are held in the upper end of the connecting rod. They are
taper reamed sothat they should be very stiff and amply strong.
The connecting rods for the main engine piston and the head
piston are exactly alike except as to length. That for the
engine piston is 12 inches between centers, while the one for
the head piston is 7 9/32 inches. Both are machined all over
from the rough. They are 1 3/8 inches outside diameter and
11/8 inches inside diameter. This gives plenty of strength
to resist the estimated high explosion pressure of 640 lbs.
per square inch, am in addition should not give out due to
Whipping because of the large diameter. The bearings in the
piston are 1 3/8 inches diameter by a total of 3 inches length.
The bearing at the main ends is 21/2 inches diameter by 2 3/4
inches length. The weight of the piston pin is 1.2 lbs. while
that of the long connecting rod is 5.3 lbs., and the short
rod weighs 4.6 lbs.

Ta6 The crank shaft is of the six throw type with
 

27
seven main journals; the front 6 inches long, all intermediate

bearings 1 3/4 inches long ami the rear 1 3/4 inches. The crank
pins are 2 3/4 inches long as already explained. Both main
journals and crank pins are 2 1/2 inches in diameter. fThe first
two crank cheeks next to the propeller are one onch think. All
others are 15/16. The width of the cheek in all cases is

3 inches. The main journals am crank pins are drilled out
11/2 inches. The weight of the crank shaft complete is

87 pounds, Owing to the small throw of the eccentric shaft it is
impossible to drill out either the main journal or the crank

pin as is done with the crank shaft, However a 1/2 inch oil
hole is drilled from end to end. COheeks of the eccentric shaft
are made use of as eccentrics for operating the sleeve valves.
The cheeks are lightened by drilling as shown in Pig. 5. In
addition to these holes two 5/8 inch holes are drilled into the
crank pins from an angle as shown in Pig. 7. Altogether this
lightening is considerable. The main journals and crank pin
journals of this shaft are exactly the same size as those for
the crank shaft. The weight of the eccentric shaft is 61.3 journals.
73-6 In order to make the sleeves as light as
possible it was decided to use steel tubing. The inner sleeve
of course must be the heavier am is made 3/32 inch thick, while
the outer sleeve is only 1/16 inch thick. The upper end to which
the pin is attached is only a segment of a circle and is made
1/4 inch thick. The pin is to be welded to the sleeve. A slot
into which the guide pins fit serve to prevent the sleeves from
rotating, which tendency they are liable to have owing to the
shortness of the rods--only slightly over 4 inches for a atroke

of 3 inches. It ig intended to finish the pin by drilling and

 
28
hollow milling. This construction is relatively light and
should be as cheap as any other which might be adopted. The
inner sleeve weighs 6.7 lbs. while the outer one weighs only
4.3 lbs. As already stated, the rods for operating these
sleeves are very short. They are bronze or brass castings
and are not fitted with other bearing material or with bushings. |
That for the outer sleeve wighs 1.94 lbs., while that for the
inner sleeve which is slightly longer weighs 2.035 lbs.

74. The cylinders and upper half of the crank case
are cast en-bloc of aluminum with large openings on both sides
of the water space and also on both sides of the upper case,
Since the cares can be located very accurately a thickness of
5/32 inch for the cylinder barrels as finished is deemed
sufficient. The crank case casting is well ribbed and should
be very rigid whem assembled. The explosion pressure is
communicated through the two pistons to the two connecting rods,
and from there to the two shafts. Long through bolts are
provided between the bearing caps for the main journals of the
crank shaft and eccentric shaft, so that mone of the direct
stress from this source is communicated to the cylinder. These
through bolts pass through the water jacket. Water leaks are
provided against by the use of gaskets and half nuts at the
upper and lower ends of the jackets. The weight of cylinders
and upper crank case complete with all bolts, covers, etc., is
129.5 lbs.

75. After comparative layouts, it was decided to
use the straight spur gear drive for the eccentric shaft as
about 15 H.P. must be transmitted to operate the head pistons.
Then, too, the two magnetos and the water pump can be mounted
29
on the gear housing with a minimum required weight for making
connections.

766 All pumps are driven by an eccentric at the rear
end of the crank shaft. Two oil cylinders are provided and one
air cylinder. Oil is drawn from the bottom of the crank case
and pumped to the tank by one pump. The other pump serves to
take this cool oil and pump it under pressure to the main
journals whence it is distributed to the crank pins, the cylinders
being lubricated by the oil thrown from the rotating parts.
Excess oil goes through the vertical pipe in the gear housing

to above the top gear where a pressure regulator and pulsation
damper are provided. The overflow is directed onto the gear
train, from whence it flows to the lower crank case and is
returned to the tank. The air pump supplies air to maintain
pressure on the gasoline tank which may be mounted anywhere in
the fusilage.
SPECIFICATIONS
Jacklin Bens
190 H.P. 160 _H.P.
Cycle of Operation Modified 4estroke
4~-stro ke
No. of Cylinders 6 6
Bore 5 5.12
Stroke 9 7.09
Stroke-Bor e-Ratio 1.400 1.385
Area of Piston
sq. in. 19.635 20.60
Total Piston area
sq. in. 117.81 123-60
Piston displacement
of one cylinder 137.45 146.05
Total Piston displ. 824.67 876.30
Vol of clearance
space 30.77 41.65
Comp ression ratio
Normal B.H.P. 190 160
" speed R.P.M. 1525 1400
Piston speed ft./
min. 1780 1655
Brake mean effective
pressure 129.2 103.4
(est) (est)
Mech. efficiency 80% 87.50%
Indicated mean
pressure 161.5 117.7
Cu. in. of Piston |
displ. per B.H.P. 4.34 5.48
Sqe in. of Piston
area per BoHeP. e620 0773
H.P./ou.ft. of
Piston displ. 374 515.5

Beng
200 _ H.P.

ditto
6
5.71
7,48
1.31-1

20.59
153.57

191.386
1148.32

48.66

5=1
230
1400

1744

113

85%

133
4.99
«667

546.3

Mercedes
£260 HP.

ditto
6
63
7.09
1.125-1

31.164
186.984

220.82
1324.92

56.12

494-1
252
1400

1655

107.5

85%

126.5
5.25
074

529.14
H.P./sq.ft. of

Piston area 279.5
Direction of

rotation of crank

& propellor Clock
No. of Carburetors 2
Type Zenith
Mixture control an to
Ignition

No. of units ZL
Mag. or Batt. Mag.

Spark plugs per cyl. 2

ww w

position head

Cylinder numbering
from propeller

Firing order from 1-5-3
6 4

propeller -2-
Timing z0°
Mag. speed 11/2

eng.
Type of Cylinder I-head
Material al
How made cast
en-bl oc
Type of valves Slide
Loc ation opposite
sides
Valve Timing
Inlet opens 0°

Inlet closes 45°

186.4

anti-
clock

Benz
2 jet

auto

Mag e
2

sides

1L=2-3
4-5-6

1-5-3
6-2-4

32° early

C.l

cast &
welded

poppett
head

60°

215.9

anti -
clock

Benz
2 jet

auto

mag.
2

sides

1-2-3
4-5-6

1-5-3
6-2-4

30° early

2/3
eng.

Cel

cast &
welded

poppett
head

o1
194.6

anti-
clock

1

Mercedes
2 jet

auto

Individual
Steel
welded

poppett
head

1°
49.3°
Exhaust opens
Exhaust closes

Valve Detail
Inlet Valves

Port dia

Lift

Width of opening
Max. height

Mean height of
opening

Mean area

Max. area

Exhaust Valves
Port dia

Lift

Width of opening
Max. height "
Mean height "
Mean area

Max. area

57°

5.75
©875

465
2.68
5.04

5.75
«6875
«406

254

3.96

Gas Velocity - ft./sec.

Inlet valve
Exhaust valve

142
150

Piston Pin Bearings

Dia
Length

Location

3

piston

60°
18°

22420
1433

5029

20420
0433

3229

172.8
172.8

1.18
29]

in rod

60°
20°

2.98
0465

5.96

2206
0443

5.68

124
130

1.52
5.57

in rod

32
50.6°
17.6"

22193
~402

5.44

2-193
~402

5.4

158
145

1.667
5.78

in rod
Material

Fastening
Area in sq.in.
Ratio of piston

Bosses
Al.

in rod

4.125

area to projected

area

Piston Type

Material
Total

4.76 to l

Al

Distance from center
of pin to top of

piston

2-500"

Length over all 5.00"

Piston Rings - no

Material

Type
Width

Crank Shaft

Type.”
\

No. of bearings

Dia of main
bearings

Dia of crank

pin

Length of front

" of inter-

mediate

2
Cel
3 piece
5-16"

hollow

2500

2-500
3-000

1.750

Length of rear 1.750

No. of bolts ea. big.

Front

TntarmaAiata

4

2

Phosphor
Bronzg

in rod

32435

6 to l

Col
5.00

C.1
plain
©5175"

hollow
7

22165

22165
2-756

2.2047
1.868

Phosphor
Bronzg

in rod

5.38

Col
7262

Cel
plain:
° 3175"

246

2258
3.14

20142
2018

35

Phosphor
Brong

in rod

6.3

5.04

C.l
11.00

Cel
plain
e199"

22519

22519
4.094

2.019
2519
Dia hole in main 1.500

Dia hole in
crank pins 1.500

Cheek thickness
next to propeller 1"

Cheek thickness of
balance of crank 15/16"

Width of crank webs 3"
Weight of crank 87
Cylinder centers 6.500

Dia of drilled
oil holes in shaft (4)-4"

Connecting Rod - No 6

Type Tubular

1.1024

1.1024

0.945

2.598
72.70
6.69

01575

6
Tubular

Length c toc 12% & 79/32 12.37

No. of bolts-

lower end 4
Dia 0375
Weight of upper

end 1.37
Weight of lower

end 3.93

Total 5.35 & 4.6
Dia 1.375
Bored out 1.125

Connecting Rod Bearing
Dia 2.500
Length (actual) 2.750
Effective length 2.500
Area 6625

Ratio piston area
to projected area 3.14

4.
©5937

1.8

302
5.00

2-165

£0244
4,86

4.18

1.072

1.072

109.25

6
Tubular

24 502.

4# 1l2o0z.
7# loz.
1.429
1.190

2238

28
6.65

3.82

34
1.59

1.39

1 1/18

139 .5

6
H section

12.95

oH L086

4# 1402.
l#

2519
5.175
2925
7.36

4.3
35

General Analysis of Weights and Measurements.

 

Jacklin Beng Mercedes Beng Mercedes
190 HP. 160 HP. 160 HP. 250 HP. 260 HP,
Cylinders 129.50 168.00 115.5 265.50 205.70
Base chamber (top) ° 100.00 118.59 103.25
" " (bottom) 19.75 65.75 73.00 110.25
Crank shaft complete 87.00 72.70 70.00 109.25 139 .50
Propeller hub " 12.00 12.00 13,00 19.32
" bearing 3.75 3.81 2.56
Camshaft complete 61,30 10.50 10.50 16.50
Pistons complete 54.60 30.00 41.10 45.72 64.50
Connecting rods 59 .40 30 00 30 .00 42.78 42.00
Valves complete 66.00 11.91 13.68 27.36 19.00
" operating gear 23.82 14.20 5.94 38 .4 15,00
Inlet manifolds 1.50 8.125 11.75 5.25
Carburetors complete 8.00 725 12.73 13.08
Magnetos complete 27.00 29.00 30.00 21.50
Ignition wiring &
tubes 4.75 4.35 4.00 4.24
Water pumps complete 5.83 4.81 6.87 8.75
" pipes 17.00 3.00 19.00
Oil pumps complete 5.00 4.556 12.00 2.89
Air " " 1.50 1.125 1.50 }
Fuel pumps 6.75
Exhaust Manifold 10.00 15.00 26400
Miscellaneous 15.00 7.53
Total 612.7 591.5 618. 848.32 936.00
Beng
160 H.P.

Mercedes
160 H.P.

56

Jacklin
190 H.P.
Weight per B.H.P. a3
Weight of fuel per
hour 112.5
Weight of oil per
hour 6.00

Total weight of fuel
and oil per hour 118.5

Grogs weight of engine
in running order
less fuel and oil.
Cooling system at
e65 lbs. per. B.H.P. 726.5

Weight per B.H.P.
ditto 32822

Gross weight of
engine in running
order with fuel
and oil for six
hours. Tankage
reckoned at 10%
of weight of fuel
amd oil 1508.6

Weight per B.H.P
with fuel and oil

for six hours. 7.93
Overall length 64
" width 18
" height 46
Total cubic space
cu. in. 53,000
Cubic space

necessary per
B.HeP. - cu. in. 279

695 .5

4.35

1277.5

7.98
61.7
25.0
41.2

63,510

397

722

4.51

1304

8.15

Beng Mercedes
250 H.P. 260 H.P.
3.68 3-71
135 140 .63
5.06 5.31
140 .06 145.94
996 1098
4.43 4.36
1920.89 2061
8.35 8.18
76.8 77.5"
2328 26.6
53.5 46.0"
97,800 94,900
425 365
37

77. The figures from the table showing the
compariosn of weights, etc., show that this engine would have
an advantage over existing types as regards weight, and that it
has a greater advantage as regards the space necessary for a
given H.P. The writer purposely selected the most unfavorable
conditions for meking the comparison by using the aviation type,
as there ig no question but that these engines have received
more real thought and development than any other gas engines.
The weight of this engine could be further reduced by lightening
the vital parts, such as the crank shaft, eccentric shaft and
connecting rods, and by hand finishing the cylinder and crank
case castings to thinner sections than shown, to between 560
and 575 pounds. This would mean one horse-power to each 2.9
to 3.2 pounds of engine at 1525 R.P.M.
786 Conclusi ons; -

a. Even when the vital parts are not designed with
the lowest possible factor of safety, there is an indicated
reduction in the "weight of engine per horsepower developed.”

be With the parts trimmed down and with a good deal
of hand finishing, there is an indicated reduction of about
One-half pound in the weight of the engine per horsepower
developed.

c. A muffler can be used with this engine with less
effect on the power developed than in present engines, since
& very small quantity of exhaust gases are left in the cylinder.
This might be a great advantage for certain purposes in aviation,
and is absolutely necessary in motor cars. The effect of back
pressure on the distribution and induction of the new charge is

entirely removed, so that all cylinders can easily procure
58.

@ proper charge.

d. No springs, valve rockers, etc., are present to
give trouble. The valves are opened and closed positively
at the proper time.

e. When used in a motor car, this engine should
Operate on low throttle better than present types, since
practically all the charge in the cylinder is inflammable
and ignition would always occur.

£. The shape of the combustion chamber is, favorablp
to the obtaining ‘higher thermal efficiency than that used in
most poppett valve engines.

ge When lower compression pressures are used as
required when the engine is to be operated on present grades
of gasoline or kerosene, the gain in power would be greater
than that shown in this case. The stroke of the head piston
would be greater than when higher compression ratios are used.
Therefore the head piston would have a greater effect in
increasing the amount of fuel handled. The volumetric
efficiency of this engine would be no lower for a low
compression engine for a higher compression type as in the case
already investigated. Since few automobile engines have a
volumetric efficiency of over 75 per cent, the use of this type
engine may result in a 33 per cent or even greater increase
in power.

h. If more power were not desired, a greater drop
in pressure dmring the inlet stroke might be used. This would
assist in vaporizing the fuel.

i. A considerable decrease in the space required per

horsepower developed is indicated. The fusilage of an
39

: 4 lw
te Bie 0S, fa?

= Stee B

airplane or the engine compartment (nooa) of a motor car may
be made smaller if this engine is used. This would result
in a reduction of the weight of the vehicle am should be
credited to the engine.

je. All in all, this investigation shows that this type
has attractive possibilities. It has prompted the writer to
construct an engine of this type with a few variations from
the construction shown to increase the thermal efficiency and
to adapt it to motor car use. This engine is expected to be

in operation in the near future.
 
 
ae

STATE UNIV

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