AN INVESTIGATION OF THE HEAT REJECTION TO THE
ENGINE OIL IN AN INTERNAL COMBUSTION ENGINE
by
Lester Simon
and
William A. Jack
SUBMITTED IN PARTIAL FULFILLMENT OF THE
REQUIREMENTS
FOR THE DEGREE OF
BACHELOR OF SCIENCE
in
MECHANICAL ENGINEERING
from the
MASSACHUSETTS INSTITUTE OF TECHNOLOGY
1944
Acceptance:
Instructor in Charge of Thesis,_
t'
414 Beacon Street
Boston 15, Massachusetts
February 10, 1944
Professor G. W. Swett
Secretary of the Faculty
Massachusetts Institute of Technology
Dear Sir:
In accordance with the requirements
for the
degree of Bachelor of Science, we herewith respectfully
submit a thesis entitled "An Investigation
of the.Heat
Rejection to the Engine Oil in an Internal Combustion
Engine".
Sincerely yours,
Lester
Stimon
William A. JarkB
ACKNOWLEDGMENT
The authors wish to express their appreciation
to Professor A. R. Rogowski for his advice and interest
in conducting this investigation
and also to the Sloan
Laboratory personnel for their valuable assistance.
TABLE OF CONTENTS
Page
Title
Introduction
1
Previous Investigations
2
Body of Thesis
Summary of Results
3
Apparatus
5
Method of Testing
10
Direct Heat Losses
14
Frietion Heat
18
Total Heat Rejection
24
Future Investigatgations
27
Appendix
Bibliography
29
Schematic Diagram
30
Photographs
31
Calculations
33
Graphs
36
Data
51
LIST OF GRAPHS
Fig. 7
Effect of Varying Oil Pressure on Total Heat
ReJection
Fig. 8
Effect of Varying Water Jacket Temperature on
Fig. 9
Effect of Varying Oil Temperature on Total
Heat Rejected
Fig. 10
Logarithmic Plot of Heat Rejection vs. Indicated Horsepower for Different Speeds
Fig. 11
Logarithmic Plot of Heat Rejection vs. Indicated Horsepower for Different Speeds
Fig. 12
Friction Horsepower vs. Engine Speed
Fig. 13
Friction Heat Rejected vs. Friction Horsepower
Fig. 14
Logarithmic Plot of Heat Rejection vs. Friction Horsepower
Fig. 15
Total Heat Rejected vs. Indicated Mean Effective Pressure
Fig. 16
Total Heat Rejected vs. Indicated Horsepower
Fig. 17
Total Heat Rejected vs. Brake Horsepower
Fig. 18
Total Heat Rejected vs. Brake Mean Effective
Pressure
Fig. 19
Total Heat Rejection vs. Engine Speed
Fig. 20
Total Heat Rejected vs. Engine Speed
Total Heat Rejection
INTRODU)TJION
-
-
INTRODUCTION
The purpose of this investigation was to determine and analyze the nature of the heat rejection to the
engine oil under various operating conditions.
This in-
cluded a verification of previous workl* and-a separation
of the components of total heat rejected.
the contributing
By analyzing
effects of each component the general
nature of the total heat rejection was determined to be a
function of only two variables,
the friction and the in-
dicated horsepower.
* References are to bibliography, page 29.
PREVIOUS INVESTIGATIONS
An investigation of the literature available
showed there has been little work done on the problem of
heat rejection
to the engine oil and the closely related
question of engine friction.
Kaneb and Hoeyl found
that the total heat rejection to the oil showedan increase which was linear with speed at constant IMEP. Increasing load at a constant speed showed a slight increase
in heat transfer at low loads while at higher loads the
transfer was much greater.
Over the operating range
the heat transfer increased slightly less than fifty per
cent as the load was quadrupled.
As these tests were
run at only two speeds, 1800 and 2000 RPM, definite conclusions could not be drawn although trends were estab-
li shed.
The object
of an investigation
by McLeod2 was
to determine the degree of error in the motoring test
and to establish a correlation between it and the true
engine friction.
This aspect will be discussed later
in conjunction with the motoring tests made in this inves-
tigation(see page 20 ).
3
SUTMIARY OF RESULTS
It was found that the heat rejected to the
engine oil was some function of the indicated and the
Empirical equations were developed
friction horsepower.
for the heat rejected due to friction and heat rejected
These separate relationships
from the cylinder gases.
were combined to obtain an empirical relationship for
the total heat rejected to the engine oil.
All are given
below: first in general form, then with the constants
evaluated for the particular engine tested.
All rela-
tionships for heat transfer are in Btu/minute.
General:
nl
Qf = EK(FHP)
(friction heat rejected to oil)
O
K(IRHP)
(heat rejected to il from
cylinder gases)
K8(
- FHP)
+ K5 ( I
)n2
(total heat rejected to oil)
Qt
=
Plymouth Engine Tested:
1
= 1.4(FHP)
Q
5
Qr = 0.82(IP)
t
l4(FP)1.5
+ O.2(IHP) 1.1
The observed data showed that the engine speed
exerts much influence on the heat rejected to the oil.
-4-
The heat rejected to the oil increased threefold over the
range of speed tested (1200 to 2400 RPM).
It would,
therefore, be necessary to use large oil coolers with
large high-speed engines.
The change in heat rejected over the range of
output tested (3 to 46 BHP) at constant speed was small
compared to the variation found for the speed range, and
it remained approximately
the same at all speeds.
At
high engine speeds, load or output had comparatively
little effect on the heat rejected to the oil.
At low
speeds, load became more influential because of the small
heat transfer involved.
It was also made clear that the oil is an im-
portant factor in absorbing the heat rejected from the engine.
The oil system absorbed nearly a quarter the total
heat rejected from the engine --
the remainder was ab-
sorbed by the cooling water system or cooling air surfaces.
APPAIRATUS
The tests
represented were performed on a six-
cylinder
1936 Plymouth engine having the following specifications:
Bore
Stroke
-
-
-
-
-
-
-
- - 3-1/8
inches
- - 4-.3/8 inches
Piston Displacement - - 201.3 cubic inches
Compression
Ratio - - - 6.70
RatedOutput
- - - - - 82 brake horsepower
at 3600 RPM
Load for firing and power for motoring was supplied by a dynamometer having the following specifications:
Current
Voltage
Output
-
-
The engine
necessary alterations
-
-
-
-
-
-
-
117
250
100
amperes
volts
horsepower
was of standard manufacture, but
were made to facilitate
testing.
A description of the engine set-up previously used for
similar tests on this engine is given below.1
Old Set-up
The crankcase panwas lagged with asbestos to
prevent appreciable.heat
transfer from the oil to the at-
mo sphere.
The engine oil pump was disconnected and replaced by an external electric motor-driven gear pump.
The oil was pumped from an external reservoir to the en-
- 6
-
gine through the port which originally contained the pressure relief valve in the lubricating
system of the engine.
The oil circulated through the engine and returned to the
external reservoir by gravity.
A long pipe was hinged
to a stand beneath the engine.
This pipe could be swung
so that it bypassed the gravity oil flow from the engine
and emptied it into a weighing pan placed on a platform
the rate of
balance. By meansof this weighing system,
oil flow could be determined.
Several heat exchangers using steam as the heating element were installed in the external oil feed line.
Cooling water was also piped to these heat exchangers in
order to effect a more sensitive regulation of the oil in-
let temperature.
Thermometers were placed at the inlet
and the outlet oil ports.
These thermometers were located
as near the ports as possible in order to obtain accurate
readings of the inlet and outlet temperatures.
An oil pressure gage connected to the inlet oil
port was used for the purpose of determining the relative
rate of oil flow through the engine.
To keep the flow
nearly constant, a gate valve was installed in the external oil line before the engine inlet port.
The engine exhaust line was attached to the lab-
oratory trench pump system, and the air intake system was
left unchanged from the original engine.
- 7
Cooling water was pumped to the engine Jacket
and to a cooler surrounding the exhaust pipe.
A thermo-
couple was inserted in the Jacket cooling system, and a
gate valve was used to regulate cooling water flow and the
resultant Jacket temperature.
A surge tank was added to
this cooling system to maintain a fairly constant rate of
flow and at the same time to keep the cooling water tem-
perature more stable.
New Set-up
To carry out the test routine decided upon, it
was found necessary to make several changes and additions
to the original arrangement as described above.
set-up used is shown diagrammatically
and in Figs. 5 and 6, pages 3
The new
in Fig. 4, page
and 32
0 ,
, respectively.
First, the original asbestos layer was removed
from the crankcase and replaced by a fresh and slightly
thicker application of asbestos.
Chicken wire,served to
bind the asbestos coating to the pan.
These precautions
were taken to prevent damageto the insulation from exces-
sive vibration.
Next, the oil-weighing system was found to be
unsatisfactory.
The bypass pipe to the weighing pan was
a source of some error in determining rate of oil flow.
This rate of flow was measured from the end of the bypass
giving not the true rate of flow through the engine, but
the rate of flow through the engine as modified by the
-
8
-
time lag produced by the friction in the long pipe.
More-
over, the process of emptying the contents of the weighing
pan into the reservoir after each run was an unnecessary
difficulty.
These disadvantages in measuring the rate of oil
flow were eliminated by rearranging the weighing system so
that oil from the engine outlet port flowed by gravity
tank.
directly into a five-gallon
From this intermediate
tank the oil flowed by gravity through a short pipe to
another five-gallon tank placed on a platform balance.
A two-way valve was attached to the end of this short pipe
so that flow to the weighing
tank could be stopped quickly.
The suction side of the external oil pump and its bypass
line also were led from this weighing tank.
Now, the oil flow to the engine could be measured accurately
by determining the time necessary for a
given weight of oil to be pumped out of the weighing
and into the engine.
tank
Once equilibrium had been estab-
lished, the rate of flow measured represented the rate of
flow of oil to the engine itself.
The oil flowed contin-
uously; no time had to be taken to remove the contents of
one tank and replace the contents of the other.
During the trial runs, it was found that the oil
pressure in the engine became very low at high engine
speeds.
To compensate for this drop in pressure (which
also meant a corresponding
drop in the rate of oil flow),
9
a gate valve was inserted in the pump bypass.
Shutting
this valve raised the flow and pressure from the discharge
end of the pump.
To maintain approximately constant back pressure,
the exhaust was disconnected from the laboratory line.
This prevented
a change in exhaustconditions
in case the
laboratory line was being evacuated.
A steam line was connected to the water jacket
cooling system so that jacket temperatures similar to
those during firing could be maintained for the motoring
runs.
An exhaust gas analyzer was used to determine
the fuel-air ratio supplied to the engine.
A strobotac
was used to check the accuracy of the cable-driven tachometer.
Both of these instruments were soon abandoned in
view of the slight effect produced by a change in fuel-air
ratio and the difficulties encountered from use of the
strobotac as compared to the accuracy possible with the
other measurements in the tests.
The rest of the original set-up was left intact.
It was considered adequate for the accuracy desired and
proved to be so in the ensuing tests.
_ 10 -
METHOD OF TESTING
Preliminary Tests
In this research there was a considerable number of variables to deal with.
The dependent variable
in all tests was someportion of the heat rejected to
the
oil, i.e., total heat rejected, friction heat, or that
portion rejected from the hot engine parts.
The independ-
ent variables were oil pressure, water Jacket temperature,
oil inlet and outlet temperatures, spark advance, fuelair ratio. speed, and load.
sidered to be constant.
All other factors were con-
For best results only one inde-
pendent variable should be varied, with all others remaining constant; but, to accomplish the purpose of this investigation in the limited time available, it was decided
to determine the relative importance of each variable on
the heat rejection to the oil.
Runs were taken to find
the allowable operating range for each variable without
impairing the accuracy of the heat transfer measurements.
Greater accuracy than that of the temperature readings
was unwarranted.
In each of the following tests, all
other variables were held rigidly constant.
Over the total pressure
range, 40 to
O80psig.,
it was found that the greatest change in heat rejected
amounted to about ten per cent from the mean value, 65 psig.
which had been selected for use in the testing.
It was
- 11
therefore decided that a variation between 60 and 70 psig.
would be permissable as this allows a percentage error of
not more than five per cent. (See Fig. 7)
Water jacket temperature changes have some effect on heat rejection.
Fig.
over the range tested to be 14%.
shows the maximum change
With reference to this
curve, it was considered adequate to operate within a temperature range of 155 degrees to 170 degrees, Fahrenheit.
The maximum possible variation in total heat rejected over
this range then becomes about three per cent.
The oil inlet and outlet temperature have a
marked effect on the total heat transfer.
Fig. 9b indi-
cates that as much as 74% change in heat rejected occurs
for the small outlet temperature range of 185 to 195 degrees Fahrenheit.
Heat rejected is almost as sensitive
to inlet temperature changes.
Fig. 9a shows the quan-
titative changes in heat rejection with respect to temperature while Fig. 9b gives the percentage variation.
From this preliminary test, it was decided to keep the
mean temperature of the oil at a relatively constant value.
These readings of temperature introduced the controlling
error of all the results.
With the range selected the
maximum error introduced was under fifteen per cent.
The change in the heat rejection with change in
spark advance was never more than three per cent over a
wide range.
(See data sheet Number 1, test Number 5.)
-
For the first
12
-
tests, a check on the fuel-air
ratio indicated no appreciable change during operation.
The fuel-air ratio was then assumed to be constant at
0 .0785.
A strobotac was used during the first few tests
but it was found tna1 for the accuracy desired, speed
could be maintained within the limiting range by merely
using the RPM dial indicator.
Main Tests
With limitations necessary for sufficient accuracy in view, the actual testing was begun.
A series
of constant speed tests was-performed with the brake
Allowing twenty to thirty minutes
horsepower varying.
for the establishment of equilibrium for each run, measurements of brake load, Jacket temperature, oil pressure,
oil inlet and outlet temperature, speed, and rate of oil
flow were then taken.
To reduce the number of tests,
constant brake horsepower settings were taken during the
speed runs by calculating the correct brake load in advance.
The speed range was 1200 to 2400 RPM with incre-
ments of 200 RPM.
The brake horsepower range was from
3.2 to 46.
Motoring runs were obtained separately from the
corresponding firing runs.
Methods of maintaining similar
operating characteristics will be explained below.
For
-
13 -
the motoring runs, measurement of speed, friction (brake)
load, jacket temperature, oil pressure, oil inlet and outlet temperature, and rate of oil flow were taken as before.
The speed and load ranges were limited by the
characteristics of the cradle dynamometer.
Measurement
s
Brake
Engine speed was read on a tachometer.
and friction loads were read from the balance of the dynamometer.
Jacket temperatures were obtained from a dial
gage connected to a thermocouple in the cooling system.
Oil pressure readings were taken from a dial gage connected to the inlet
let port.
side of the oil system at the engine in-
The oil inlet temperature was measured by a
mercury thermometer placed in a well at the engine inlet
port and the oil outlet temperature was measured by a
mercury thermometer at the exit passage from the engine.
A running balance was taken
or the rate of oil flow.
First, the weighing tank was allowed to fill up.
Then
the two-way valve in the line leading from the intermediate tank was closed.
The usual weight of oil used dur-
ing the running balance was ten pounds.
All measurements were an average of at least
three runs after equilibrium had been established.
Par-
ticular attention was paid to the reading of the oil temperatures as they were the major source of error.
-
14
-
DIRECT HEAT LOSSES
General Theory
Heat transfer within the internal.combustion
engine may be divided into two processes.
One is the
heat given off to the engine parts by the working fluid,
and the other is the heat produced by friction of the
moving parts.
It has been found by dimensional analysis4 of
the variables involved that the heat transfer from the
working fluid may be represented by
S. = ATAC(es)n(n -
(1)
where AT = mean temperature difference between
the gas and the cylinder walls
A = exposed area
Cp = mean specific heat of the gas
e = mean gas density
S = mean piston speed
B= mean gas viscosity
n = exponent determined by experiment
Hence, for a given engine operating with approximately the same outside cylinder temperature (AT is constant), the heat transfer from the working fluid is seen
to be some function
density.
of piston speed (or RPM) and gas
-
r( e s)n or
15
-
Qra( e x RPM)n
(2)
Now, the indicated mean effective pressure of an
engine may be represented
by
IMEP =-Je (F x Ex
where
1 t) x
(3)
14 4
J = mechanical equivalent of heat
(a constant)
e
= density of gases in cylinder
F = fuel-air ratio
heating value of fuel
E
It
=
indicated thermal efficiency
The heating value of the fuel, fuel-air ratio,
and the indicated thermal efficiency were essentially
Therefore IEP
constant for the tests performed.
depends
only on the cylinder density of the working fluid.
x
IMEPaL
From which equation (2) becomes
Qr
=
(4)
K1 (INMEPx RPM)na
Since indicated horsepower is directly proportional to IEP
at constant speed, equation (4) may be
written
Qr =
2
(5)
(IBP)n
Either equation (4)
or (5)
is suitable for de-
termining the heat rejection from the working fluid.
Equation (1) was developed by first obtaining
an expression for the heat transferred from an air-cooled
cylinder to the air and then by making suitable assumptions4
JL relationship was obtained for the flow of
-
16 -
heat from the hot gases to the cylinder walls.
As the
flow of the gases in the cylinder is turbulent,
the
ponent
ex-
should be about 0.8.
n
Discussion of Results:
Part of the heat dissipated in an internal combustion engine goes to the cooling water, while the rest
of it is removed by the oil and crankcase.
Judge3 showed
that of the total heat dissipated by an engine, approximately fifty-five per cent was given up to the cooling
waster and approximately forty-five per cent to the engine
oil and crankcase.
The results obtained in this investi-
gation were much lower.
Lubricating oil vaporizing from
the underside of the piston and the cylinder walls counted
for some of this transfer.
Direct conduction to the
crankcase from which the heat is transferred to both
the oil inside and the air outside4
remainder.
accounted for the
Since the crankcase was lagged with insula-
tion, all the heat transferred from it was assumed to go
to the lubricating oil, none to the atmosphere.
It was shown earlier that the total heat reJection is an exponential function of IHP and speed (equation 5) and that in all probability the value of the exponent should be less than unity.
By reference to the
data obtained during this investigation, the general verification of the theory developed above was found.
-
17
-
By taking the test data for all runs made and plotting the heat rejected from the hot engine parts Qr --
deter-
mined by subtracting the motoring heat rejection Qf from the
total rejection QT --
as a function of indicated horsepower
for the range of speed used on log-log graph paper, Fig. 10 was
obtained. Except for deviations at low loads all the curves had
approximately the same slope.
By replotting the same points
for Fig. 11 and weighing the scatter of the data, the general
slope, n2 , was obtained. K5 was then determined (see Appendix).
ri
2 had a value of approximately one.
With the development of
Eq.(5) it was stated that the usual value of n2 would be around
0.8.
The reason for the higher value that was found in this
investigation
may be due to the change in the direction or pro-
portion of heat transfer in cylinders as conditions change.
The oil splashed onto the cylinder walls and the change in pis-
ton ring action may be the chief influence. The piston may
transfer less heat through the rings and more to the oil at
higher speeds. As the speed increases, there is probably a
change in the type of lubrication with a corresponding change
in resistance to heat transfer in the cylinders. As a result
of this the piston may ride up on the oil film and a larger percentage of heat will be rejected to the oil.
The resulting empirical formula developed for the
heat rejected to the oil due to cylinder gases became for this
particular engine
Qr =
. 82 (IP)
Sample calculations
Btu/minute
in the Appendix
(5a)
show that this
equation is within the accuracy of the observed data.
- 18-
FRICTION
HEAT
General Theory
So far, only a portion of the heat rejected
to the oil has been discussed.
The remaining portion
due to friction also can be explained in a manner similar to that of rejection from the hot gases.
First, it
is necessary to separate the friction into two types -coulomb and viscous friction.
Coulomb friction is independent of the relative velocity of the sliding surfaces, but viscous fric-
tion will
increase
as some function
of the speed. The
coefficient of viscous friction is a function of RPM.
f=0 (i)
where
0 = some function
= viscosity of lubricant
N=
RPM
P = unit
Professors E.
bearing pressure
S. Taylor and C. F
Taylor6 assume
that
6.
the coefficient of friction is nearly proportional to
Friction = fPa# N
and
Qf = K3 (Friction x N)
or
Qf = 4(LCN x N)
For a given lubricating oil at a given condition
is a constant, and
Qf = K5 N 2 (For viscous friction)
(6)
19
-
Adding the effect of constant coulomb friction
to that of the varying viscous friction should give the
heat rejected
to the oil due to friction.
However, no
method of testing has yet been devised to separate coulomb
and viscous friction.
In addition there exist regions of
partial lubrication which do not obey either of the above
laws.
Consequently the results of this research contain
only heat losses due to total friction.
Unfortunately,
the theoretical formulae developed for determining the
coulomb and viscous friction separately cannot be used,
but they do give an insight into the nature of the resulting combined friction.
There should be some intermed-
iate formulation for total friction heat rejected to the
oil lying between those for coulomb and viscous friction,
but closer to viscous which is the determining factor
since it increases with the square of the speed.
This relationship for friction heat rejected
should depend on speed (as is obvious from equation (6)
for viscous friction).
The exponent which the speed is
to should be less than 2, the value for viscous friction
alone.
Discussion of Results
The friction heat rejected to the oil and to
the cooling water should be the friction horsepower of
the engine.
It is impossible
to separate
the amount of
friction heat loss to the oil from that to the water
Jackets.
The proportion to each varies under different
conditions.
Since coulomb friction (mostly piston fric-
tion) varies linearly with speed, and viscous friction
(bearings)
varies with the square of the speed, it fol-
lows that a greater proportion of the heat will be rejected to the oil at higher
speeds.
The part of the
friction heat rejected to the oil may be expressed by
Qf
=
where
X(FP )n
X is
some function of speed and design.
From a comparison with the heat rejected due
to cylinder gases, the following equation for heat reJected due to friction will be assumed.
Qf = Kg(Fp)na
(for combined friction)
(8)
For the purpose of finding the constants in
this relationship, the results of the motoring test were
used.
The plot of FP
vs. RPM (Fig. 12) indicates a
linear proportionality between the two variables.
This
corroborates the analogy between equations for heat reJected due to friction and that due to the cylinder gases.
This also
ustifies the assumption that total friction
heat rejected is some function of FHP.
The plot of the friction heat rejected as a
function of friction horsepower (Fig. 13) shows that the
function is an exponential.
This may be explained by
the fact that at higher FHP's there is a greater per cent
of the heat rejected to the oil than to the cooling water.
.
21
-
Heat rejected to the oil comes mostly from the bearings,
and bearing friction power loss increases with the square
of the engine speed.
Heat rejected to the cooling water
comes mostly from the cylinders where friction power loss
increases linearly with the speed (because of partial lubrication of the cylinder walls).
It follows that the
curve rises slowly at low friction horsepower and more
Thus, the exponential
rapidly at higher values of FHP.
shape of the curve is Justified.
From the logarithmic plot of friction heat rejected vs. FHP (Fig. 14), the exponential, n
to be 1.5.
values
for
By dividing
values
for
Qf by corresponding
FHP raised to the exponent, 1.5,
value for K8 was determined.
, is found
an average
(See Appendix, page 3+ ).
Thus the empirical formula developed for the heat rejected
to the oil due to friction becomes for this particular
engine
Q - 1.4( FHP)1 ,5
Sample
calculations
(ga)
in the Appendix show that
this equation is within reasonable accuracy when compared
with the observed-data.
The errors resulting from the usual method of
measuring friction are explained in a paper by McLeod2 .
From his experiments it was found that there are many
errors tending to make the friction
ing
method
measured by the motor-
deviate from the true firing friction. McLeod
further
statedthat these errors were not all in the same
direction, but tended to compensate each other making the
final deviation small as is shown below.2
40
02
30
Pa
0
X
20
Firing
---
14
rk
4
Motoring
0
900
1200
1500
RPM
Fig. 1
Performing the motoring tests all together
under controlled conditions similar to those present during the firing tests obviated the fime effect on the
friction measurements.
This time effect arises from
the change in conditions which takes place during the
delay when a motoring test is taken following each firing test without the operating conditions being controlled.
It is presented by McLeod2 and reproduced below.
The close correlation between the observed data
and the calculations obtained by use of the formula derived indicates that the friction measurements were satisfactory for the accuracy desired and that the relationship
0
co
33
32---
0
31
I
30 P,0
;E
29
2827
45
75
105
136
165
Time in Seconds
Fig.
2
betweenheat rejected to the oil and friction horsepower
is valid.
TOTAL
EAT REJECTION
In the preceding sections empirical relationships have been developed separately for the two componBy a com-
ents of the heat rejected to the engine oil.
bination of the two results the total heat rejection may
be found for any speed and load as long as the Water
Jacket and oil temperatures are within the prescribed
limits, which cover the normal operating range.
Qt
=
Qf +
r
Qt = $(FHP)n;
t = 1.(FHP)'
(9)
Btu/minute
+ K 5 (IiP)n
5
ll
'
+ O.2(IHP)
(10)
(11)
"
Equation (11) checks the experimental data obtained during the tests within
the limitations of the
accuracy of about ten per cent.
An inspection of Figs. 15 to 20 shows the general nature of the total heat rejection over varying
ranges of speed and load.
The following comments on
these curves are offered in explanation:
Increased IIEP, Fig. 15, at constant speeds a
slight increase in total heat rejection to the oil.
Ex-
cept for the run at 1200 RPM all the curves have approximately the same slope.
The run at 1200 RPM will not
be discussed as it obviously contains errors.
Previously
it was stated that at constant speed, the mechanical friction would be constant but this may not be the case and
25
-
in all probability the increase in heat rejection was due
to the increased piston friction.
Under normal operating
conditions, the friction of the Journal bearings is independent of the bearing pressure and therefore does not
vary with varying cylinder pressures.
A very large part
of the mechanical losses of an engine arise from friction
of the pistons and rings.
This is due to the poor lubri-
cation conditions combined with high relative velocities
and large surfaces in contact.
From research done by
M.P.Taylor 6 , the effect of cylinder pressure on mechanical
The nature of the relationship
friction was determined.
is shown in Fig. 3 below.
28
4
4
;g 24
o(
40
12
fL
8
20
40
60
80
100
120
140
Gas Pressure on Pistons in Lbs. Per Sq. In.
Fig. 3
-26-
One of the chief causes for the shape of the above curve
is that the cylinder pressure is communicatedto the
spaces behind the piston rings, thereby increasing the
This
pressure of the rings against the cylinder walls.
is
relationship
in all probability the main cause of the
increase in total heat rejection with increasing load.
The effect of increasing engine speed at constant IMEP's on total heat rejection, Fig. 19, is practically linear.
This checks the result obtained by
Kaneb and Hoey1 and has been discussed previously.
Fig. 20 is of more practical value in determining the size of the oil cooling system needed.
The rela-
tionship between heat rejection and engine speed at constant BHP is in the form of an exponential which follows
from the reasoning on pages 14 and 15.
When running at low loads at speeds above 1600
RPM, the total heat transfer to the oil began to decrease
to a minimum and then increased as more load was applied.
The explanation for this has not been ascertained.
From comparison of the plotted results it is
obvious that speed exerts a major influence on heat reJected to the engine oil.
ant.
Load is relatively unimport-
- 27
-
FUTURE INVESTIGATIONS
If further investigation is to be done along
this line, more accurate temperature control is an absolute necessity.
As was shown earlier, the most serious
error in all runs was in the temperature reading of the
oil.
A one-degree error in reading amounted to approx-
imately a ten-per cent error in the heat rejection calculations.
Also, time should be allowed for stable oper-
ation.
The limitations of load and speed were set by
the cradle dynamometer,
so if higlher speeds are to be in-
vestigated this will necessitate a change in the set-up.
For further confirmation of the nature of the
heat rejection, it would be necessary to operate the engine under different water
acket and oil temperatures.
A more complete investigation of heat rejected
at low loads should be made to determine the reason for
the minimum in the power curves at constant speeds.
Until more research is done upon piston and
ring friction, it will be impossible to determine more
exact relationships.
The empirical formulae for determining the heat
transfer to the oil should be checked on other engines
before being accepted.
modified
Also, these formulae should be
so that some relationship
for geometrically
sim-
-
28
-
ilar engines may be obtained.
It would then be possible
to predict the heat rejected to the oil for various sizes
of
similarly designed engines.
This would be advantageous
in design work.
4
APPENDIX
- 29
-
BIBLIOGRAPHY
1.
A Study of the Heat Transfer to the Engine Oil, J.F.
Hoey Jr. and W.
aneb, M.I.T. Thesis for S.B., Course
II, 1943
2.
The Measurement of Engine Friction, M.K.McLeod, paper
given before S.A.E. Annual Meeting, January, 1937
3.
Aircraft Engines, A. W. Judge, D. van Nostrand Co.,
Inc., 19 40
4.
The Internal Combustion Engine, C. F. Taylor and
E. S. Taylor, International Textbook Company, 1938
5.
Heat Transmission, W. H. McAdams, McGraw-Hill, 1942
6.
The Effect of GasPressure on Piston Friction,
M. P. Taylor,
1936
S.A.E.
Journal,
Volume 38, No. 5, May
- 30
r4
0
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31
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...
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--- -----
_-_._ _ ______.
Fig. 5.
View of Engine Showing Control Panel,
Asbestos Insulated Crankcase, External
Oil System ( Pump, Entrance Thermometer, Heaters, Pressure Valve )
- 32 -
··
1·=-·-
.
Fig. 6.
View of Engine Showing Flow Measuring
System, Oil Pump Suction Line And ByPasts,Exit Thermometer
- 33
CALCULAT ION S
Heat
,~~~a
, Transfer
where
Q_ w GpBT
_ (10) (o.5) (10)
p
specific heat of
oil at constant
pres sure
W= rate of oil flow
= 50 BTTU/min.
in lbs./min.
6T
oil temperature
difference, OF,
Brake Horsepower
Bhp- Brake load x RPM
4000
= (87.5)(1800)
= 39.4
Friction Horsepower
Fhp= Fricticm load x RPM
.
4000-...
- 26.6)(1800)
= 12.0'
Indicated Horsepower
Bhp
Ihp= Fhp
= 12 + 39.4
= 51.4
Constant
Mean Effective Pressure
~
...._
....
Hp
__,
Brake load x RPM
-
o00
where Dynamometer Constant=
4000
-
34
-
Hp= MEP x Piston Displacement x RPM
-33000 x
MEP= 0.971 x Brake load
Piston Displacement- 201 in3.
Equation (5a)
n2
Qur K 5 (Ihp)
n - 1.1 (fromFig. 11)
21.2
25.0
34.0
40.0
45.0
50.0
57.0
1.1
Ihp
(Ihp)
20.5
23.4
30.9
35.3
37.4
39.3
43.6
27.7
32.0
=1.1
K5 3Qr(Ihp)
1.1
0.77
0.78
44.0
0.78
50.0
0.80
54.0
0,83
57.0
0.87
64.0
0.89
( K5)ave 0.82
%= 0.82(Ihp)
Equation (8a)
Q = K8 (pP)l
nl =1.5 (from Fig. 14)
Speed
,Qf
1200
1400
1600
1800
2000
2200
2400
30;5
Fhp
6.9
8.5
-35.2
44.0 10.2
57.5 12.0
77.0 13.9
91.0 16,2
120
18.4
1.5
(Fhp)
18.0
25.5
32.5
41.6
520
65.0
78.5
1.5
K8
Q (P)
1.7
1.4
1.4
1,4
1.5
1.4
1.6
(K)ave= 1.4
1.5
1.4(Fhp)
Qf
Equation
(11)
1.1
1.5
t= l.4(Fhp)4- 0.82(Ihp)
Tabulation of total heat rejection as calculated from
equation (11) and test results with percentage error.
Test
t
error
52.0
51.0
+1.96
75.5
59.2
84.1
57.8
10.2
82.0
85.6
-3.5
21.5
12.0
82.1
81.5
+0.74
2000
17.1
13.9
91.5
103.5
-11.6
2000
35.3
13.9
114.3
117.0
-2.3
2200
41,6
16.2
148
150
-1.3
2400
64.,3
18.4
191.4
191
+0,2
2400
43.8
18.4
163.5
170
-3.8
Formula
Ihp
Fhp
1200
23.9
6.9
1400
1400
34.0
21.1
8.6
8.7
1600
31.6
1800
Speed
-10.2
+2,4
-3IG.
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