-------
-2-
Table of Contents
I. Introduction
II. Summary
Page
3
5
III. Conclusions
8
IV.
A. Volumetric Efficiency Losses
B. Prevention of Volumetric Efficiency Losses
C. Optimum Fuel Delivery
Discussion
8
8
8
9
A. Adiabatic Temperature Drop vs.
Vapor Pressure
B. Heat Transfer
References
Appendix I
Appendix II
Appendix III
Appendix IV -
Appendix V
Appendix VI
Appendix VII
Appendix VIII
Partial Pressure Calculation
Temperature Drop With
Vaporization
Volumetric Efficiency Change
Due to Adiabatic Temperature Drop
Observed Volumetric Efficiency
Loss
Heat Transfer Calculations
Changes in Pumping Work due to
Temperature Changes
EPA Memo, Results of Fumigation
Testing
Test Data
11
16
22
23
25
35
38
39
44
48
63
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-3-
Methanol Vaporization: Effects on Volumetric
Efficiency and on Determination of Optimum
Fuel Delivery System
I. Introduction
The Office of Mobile Sources within the Environmental
Protection Agency has studied and evaluated alternative
transportation fuels since its formation in 1970. EPA's
responsibilities under the Clean Air Act also have
necessitated a significant regulatory role dealing with
transportation fuels. In particular, Section 211 of the
Clean Air Act requires EPA to play a key role in the
introduction of new fuels and fuel additives. Perhaps most
visible was EPA's role in the introduction of unleaded
gasoline to permit the use of catalytic converters on 1975
and later model year automobiles. More recently EPA has
responded to a growing interest in the use of oxygenates (in
particular methanol) for use in motor vehicles and for
blending with gasoline.
As part of this response to the interest in methanol, a
program was undertaken to grade several fuel utilization
concepts. One of these concepts that has been evaluated has
been loosely labeled fumigation. Normally, the term
fumigation is used in the context of diesel engines when all
or a majority of the fuel is introduced into the intake
track, as opposed to injecting the fuel into the cylinder (or
pre-chamber). Generally, fumigated diesel engines are
unthrottled engines and rely on the fuel delivery system to
regulate the load. In our testing with a throttled spark
ignition (SI) engine converted to methanol use, only a
portion (20-33%) of the fuel was injected or "fumigated" into
the intake track. Two fumigation locations (see figure 1)
were used - one significantly upstream from the throttle
valve, and the other slightly downstream of the throttle.
Both locations were upstream from the normal port injectors
which injected the majority of the fuel (for more details see
reference 1).
Testing has shown this upstream fumigation of methanol to
cause a decrease in volumetric efficiency(1)*. Volumetric
efficiency is usually defined as a measure of the mass of
air/fuel charge which is successfully taken into the cylinder
on each intake stroke. It is an "efficiency" with respect to
utilization of the cylinder space, but only indirectly with
*Numbersin parentheses designate References at end of paper.
Reference (1) is reprinted in Appendix VII.
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-4-
AIR FILTER
POPT INJECTORS
\
THROTTLE
VALVE
FUMIGATION
"INJECTORS
(LOGAT ION 2)
Figure 1
FUMIGATION
TEST SET-UP
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-5-
respect to energy or fuel utilization. A loss in volumetric
efficiency would have as its most direct consequence a
reduction in power for a given throttle position and engine
speed, including a reduction in maximum power. Pumping
losses would increase, causing a small reduction in overall
engine efficiency. The observed loss in volumetric efficiency
was opposite of the commonly held assumption that vaporizing
the fuel would cool the intake air and increase the
volumetric efficiency. Thus, a small mathematical exercise
was undertaken to attempt to quantify the physical processes
causing the real effects to be opposite of the anticipated
ones. Although the results of this exercise did not uncover
new phenomena governing engine operation (as a review of
Reference 2 will indicate), it did reinforce the importance
of existing but sometimes over-looked phenomena on engine
operation. Consideration of these phenomena may be more
important for a methanol engine than for a gasoline engine.
II. Summary
EPA tests utilizing fumigation of methanol caused a decrease
in volumetric efficiency. The apparent cause of this
reversal in the actual versus the expected result was
primarily due to heat transfer effects from the walls of the
intake passage. A secondary issue deals with the necessity
for the combination of the local partial pressure of the fuel
and the stagnation temperature to be on the vapor side of the
vapor pressure curve for the fuel in question in order to
allow vaporization to take place. The partial pressure is
the key to this secondary effect and is influenced directly
by the local equivalence ratio.* In the cases discussed,
this secondary effect had only a minor influence on the
results, primarily because the effective equivalence ratio
was very lean at the point of fumigation (i.e., the
fumigation injectors received only part of the total engine
fuel flow resulting in a much lower partial pressure of the
fuel in the local area). If, however, the local equivalence
ratio at the fumigation injector had approached a
stoichiometric mixture or richer, the partial pressure/
temperature effect would prevent a significant portion of the
fuel from vaporizing. In this case (rich condition), the
intake air temperature depression due to vaporization would
be limited to the level allowed by the vapor pressure curve
of the substance.
As stated, our experiments were not substantially effected by
the secondary effect because our local air/fuel ratios were
* Note: the Greek word "phi" is used to designate
equivalence ratio throughout the report.
-------
—6—
very lean. Under these conditions, all or nearly all of the
fuel was evaporated. On the other hand, even though the lean
mixture allowed vaporization, the temperature depression
available from this vaporization was limited because of the
small mass of fuel involved. Under these conditions, normal
component temperatures of the intake air duct-work appeared
to be sufficient to cause the initially cooled mixture of
fuel and air to heat back up to the original temperature (due
to the temperature differential between the wall and the
newly cooled mixture). Since the air returned to its
original specific volume (at its original temperature), and
the vaporized fuel had a much larger specific volume than the
liquid fuel, the volume of the mixture (air plus gaseous
fuel) was greater than the volume of intake air with no
fuel. Because the volume flow rate (Q) of the engine would
remain the same for the same manifold vacuum,* an intake
charge of this gaseous fuel-air mixture would contain less
air than an intake charge of air without the fuel.
This decrease in air flow when using a fuel-air mixture will
manifest itself in a low volumetric efficiency (VE). In our
case, if heat transfer is sufficient to reheat the mixture to
its original temperature, then the theoretical loss in
volumetric efficiency (for phi = 0.8) is around 2.3 percent
for location 1 and around 3.8 percent for location 2 (see
Table 1). The average loss obtained from an analysis of the
test results is 2.8 percent for location 1 and 4.1 percent
for location 2.
Table 1
Volumetric Efficiency
Location 1 Location 2
Calculated loss
(Including Heat Transfer) 2.3% 3.8%
Measured loss (average) 2.8% 4.1%
*For a complete derivation, see Appendix III. It works out
that the total mass flow and total gaseous volume flow vary
only slightly with changes in density for a constant pressure
drop. What changes is the portion of the total that is fuel,
the remainder is the change in inlet air flow. This can
readily be seen by comparing the component volumes of a
stoichiometric mixture per pound of fuel in Tables AIID-1 to
3 in Appendix II (air component =Mvl, fuel component = Mxv2).
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-7-
Recognizing that the average test results mask some data
scatter, we are left with the fact that fumigating a fraction
of the total fuel flow caused a loss in volumetric
efficiency. The theory suggests that heat transfer which
reheats the charge to the original temperature and expands
the fuel volume can explain a majority, if not all, of the
difference between the expected results and the actual
results. Any differences not explained by heat transfer are
probably due to experimental error, or some other phenomonea
not yet fully understood.
In principle, it should be possible to reduce the heat
transfer down stream of the fumigation point. This might be
accomplished, for example, by insulating the inner surface of
the intake track. However, the temperature drop would be
limited by the dew point of the mixture strength. For
example a carbureted system at stoichiometric conditions and
15 in. Hg MAP would require a mixture temperature above 50°F
to maintain 100% vapor. (Note: At this temperature, the
effect of fuel expansion would cause a theoretical loss in
volumetric efficiency of slightly greater than 6%.) On the
other hand, if one were willing to tolerate a portion of the
fuel in liquid form, there could be an equilibrium point
where the mix temperature is depressed sufficiently to allow
an improvement in volumetric efficiency. However,
distribution of the liquid fuel could become a problem. If
partial fumigation with an insulated manifold was used in
combination with port injection for cylinder distribution, a
small increase in volumetric efficiency might be possible
(our testing suggests 2 to 4 percent - see table 3). This
increase might only be available under certain operating
conditions due to changes in the equilibrium condition of the
mix during different operating regimes. Therefore, while it
might be possible to obtain some benefit with partial
fumigation, it would require careful design to insure that
such potential improvements would occur at the design
operating point, since it seems possible that the off-design
points would incur losses in volumetric efficiency.
The results of this experimental work and theoretical
analysis allowed us to make the following generalizations and
conclusions.
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-8-
II. Conclusions
A. Volumetric Efficiency Losses
A loss in volumetric efficiency when mixing methanol with air
in the inlet system (compared to air only) is primarily a
result of heat transfer effects warming the mixture.
B. Prevention of Volumetric Efficiency Losses
By awareness and proper manipulation of the following five
factors a potential loss in volumetric efficiency can most
likely be avoided.
1. The partial pressure of the fuel is governed by the
mixture strength (f/a), the local manifold air pressure
at the time of vaporization, and the portion of the fuel
permitted to be vaporized by the partial pressure
characteristics at the local temperature.
2. The temperature of the mixture after the heat of
vaporization is included is relative to the degree of
vaporization, the partial pressure of the fuel, and the
amount of the fuel. The temperature of the mixture is
also affected to some degree by the initial stagnation
air temperature.
3. The heat transfer to the mixture is governed by the
difference between the wall temperature and the mixture
temperature, but is also influenced by the amount of
liquid fuel coming in contact with the wall. The effects
of the liquid fuel contact may be substantial.
4. The length of the intake passage from the point of
fuel introduction to the cylinder over which the heat
transfer can take place is of concern.
5. The time during which heat transfer can take place
also affects the results. The time can be influenced by
engine speed, load (i.e., throttle position), geometry of
the intake tract (i.e., velocity in the system), or,
distance (i.e., length of tract).
C. Optimum Fuel Delivery
From these factors, it appears that the best location to add
fuel when considering volumetric efficiency (VE) with
methanol is directly into the cylinder when the intake valve
is open. Direct injection would probably be the best(2), but
port injection near the intake valve would be expected to be
a better choice on a cost/benefit basis than direct
-------
-9-
injection. Port injection, of course, would be expected to
be much better than carburetion for VE. Because of potential
side benefits of improved driveability (less manifold
wetting, etc.) and possibly better emission control, port
injection should also be better than carburetion on a
cost/benefit basis. If carburetion is used, however, more
effort should probably be directed at developing a more
uniform wet-flow cylinder distribution for the methanol
intake manifold than has been found to be necessary for a
gasoline manifold. This analysis of methanol evaporation
versus vapor pressure and temperature suggests that
sufficient manifold temperature will not be available (and
may never be available) under enough conditions to avoid wet
flow mal-distribution problems (e.g. cold start, warm-up,
etc) if a normal gasoline manifold/carburetor combination is
used when using methanol.
Considering the qualitative aspects of this analysis, one
might predict that phased port injection would perform better
than normal port injection. Such a prediction would be based
on the assumption that a majority of the fuel would be
expected to be evaporated within the cylinder when phased
port injection was used. Such evaporation when the intake
valve is open would be expected to improve volumetric
efficiency because the greater volume to surface area within
the cylinder (compared to the intake tract) would tend to
limit the heat transfer to the evaporated mixture. Secondly,
any evaporation after the intake valve is closed would reduce
the work of compression by cooling the mixture. Finally, any
heat transfer that did occur would cool the cylinder wall,
which would tend to reduce the load on the cooling system.
These assumed physical reactions could allow the potentially
cheaper phased port injection system to duplicate the
performance of a more expensive direct cylinder injection
system.
IV. Discussion
The original focus of the fumigation testing was to ascertain
any potential benefits that might be available due to the
high latent heat of methanol. An analysis in Appendix VI
suggests that if we were to cool dry air (i.e. no fuel) down
to the lowest fuel dew-point temperature experienced in
testing, we could reduce the pumping work by roughly 2
percent for a constant volume flow rate or in the case of
constant mass flow rate, 15 percent. We hypothesized that by
vaporizing only a small fraction of the fuel upstream of the
port injectors, the high latent heat in methanol would allow
some improvement in volumetric efficiency, and possibly some
reduction in the pumping work. The test results contradicted
this hypothesis.
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-10-
In order to investigate this problem/ the basic precepts
behind volumetric efficiency were examined. Volumetric
efficiency can be defined as the amount of intake air
ingested (at STP) expressed as a percentage of the swept
volume of the engine. Several factors can influence the
volumetric efficiency at wide open throttle; examples are
camshaft design, valve size, exhaust restricition, intake
path size and length, and resonant effects due to intake
geometry. At part throttle all of the above factors are
modified by the throttle valve. In our experiments all of
the basic factors were held constant since the same engine
was used. Also, all tests were conducted at the same RPM and
the same low level of torque (which required part throttle
operation). The comparisons of volumetric efficiency were
all made at the same manifold vacuum level (i.e., same
throttle position).
Under these conditions, only two basic factors are readily
apparent that could cause the volumetric efficency to
decrease when fumigation was added. One is that the basic
air ingestion characteristics of the engine were changed with
the addition of the fumigated fuel. The other is that the
fumigated fuel displaced some of the air that would normally
be ingested if the fuel were not there. In the first case,
it is hard to imagine how adding fuel to the air stream could
change the basic air ingestion characteristics of the engine
under the moderately throttled condition as occured in our
tests. At wide open throttle, factors such as poor
cylinder-to-cylinder distribution, or a change in the
resonant effects due to the additional fuel mass might affect
the volumetric efficiency, but the power would also be
expected to vary, hence dissolving the comparison at constant
power. The second consideration, the one of fuel volume
replacing air volume is more palatable, but only if the fuel
is vaporized. In liquid form, the volume displacement of the
fuel is so small that the effect on volumetric efficeincy
would be expected to be pratically negligable.
Therefore, if only the vaporized fuel has the potential to
cause the majority of observed effects, then the factors
influencing vaporization must be evaluated to determine if
these factors could have been present in sufficient magnitude
to create the observed effects. The factors affecting
vaporization are temperature and pressure. As will be shown
in the following sections, these factors have a dramatic
influence on the amount of vaporization that takes place.
The first step taken to evaluate the potential volume increase
of the fuel was to identify the expected temperature of the
intake air after a given quantity of methanol was evaporated
(regardless of other influencing factors). The following
sections address the determination of this temperature drop,
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-11-
and how the vapor pressure of the fuel, in concert with heat
transfer effects, influence the final fuel volume.
A. Adiabatic Temperature Drop vs. Vapor Pressure
If heat input is assumed to be equal to zero for the time
being, equation AII-1 (ref. 2) in Appendix II provides an
estimation of the temperature drop of the intake air. The
temperature drop predicted by the equation is governed by the
portion of the fuel evaporated and the local fuel-air ratio.
Since the first evaluation looked at temperature drop with no
heat transfer, the term Q was set at zero for the first
analysis.
(AII-1) AT = [(x) (F) (HLG) + (Q)] / (1-F + xF) (Cp)
In order to determine the amount of fuel that can be
evaporated under any given condition, the vapor pressure
versus temperature characteristics of the compound in
question must be known. Data on these characteristics was
obtained from reference 3 for methanol (see Table 2), and a
curve of that data was constructed (see Figure 2). The next
step in order to use this data would be to determine the
partial pressure of the methanol under the local conditions
which are controlled by the local fuel-air ratio and the
portion of the fuel evaporated (e.g., 25%, 50%, 75%, 100%). A
range of partial pressures can be calculated for each
fuel-air ratio (see Appendix I Section IIIC). Using these
partial pressures, and the curve in Figure 2, a dew point
temperature can be estimated such that if the' mixture
temperature were below this temperature (and at the same
Table 2
Vapor Pressure Curve for Methanol(3)
mm HgA °C
1 -44.0
5 -25.3
10 -16.2
20 - 6.0
40 4-5.0
60 +12.1
100 +21.2
200 +34.8
400 +49.9
760 +64.7
*F = f/a, HLG = heat of vaporization, x = portion evaporated
-------
-12-
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:..—.r—-:. Methanol Vapor Pressure Curve
— : — ..-'—. Figure 2
, — ; 1 1
1 1 .
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reaper
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ature
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-20
+20
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-------
1000
Vapor
Pressure
mm HgA
100 — — - :^=^=^.._-_-^.-.- -
r,
c
, End Point PHI =1.0
_ / .
rr=-:^==r=E= Mcthano-1 Vapor Pressure Curve
zS-i^r:: -'r"::^ Case 2 Dual Plenumb injectors
(33% of Total Fuel Delivered)
^ Phi = 1
0 Phi = (.8)
Figure 3
•/-. -:.~.~
? •
-40
-20
+20
+40
-1-60
-I-80
-HOO
-------
-14-
partial pressure), it would be expected that any methanol
vapor would condense back into liquid droplets until new
equilibrium conditions were satisfied. The same assumed
evaporative percentages used to calculate the range of
partial pressures can also be substituted into the intake air
temperature drop equation (e.g., AII-1), to obtain estimated
adiabetic intake air temperatures (Tl- ATI in Appendix II
Section IIB).
If the estimated intake air temperature from equation AII-1
is warmer than the dew point temperatures (i.e. Tl-ATl
greater than TDEW, Appendix II, Section IIB) then
vaporization could be assumed to take place.
In our case, where we introduced only 20 percent of the fuel
at location 1 (Figure 1) upstream of the normal injection
system (the remainder from the port injectors), all of the
fuel can be evaporated at this upstream location due to
partial pressure effects (Location 1, Appendix II, Section
IIB). However, the richer of the two equivalence ratios
indicates that total vaporization is marginal without
additional heat since the depressed intake air temperature is
effectively the same as the dew point temperature. If we
look at Location 2 (Figure 1) where 33% of the fuel is
introduced upstream of the port injectors (Appendix II,
Section IIB), the leaner of the two equivalence ratios would
have marginal evaporation, while the richer ratio cannot
vaporize all of the fuel from partial pressure effects alone
(only about 78% will evaporate). The visualization of this
effect of incomplete evaporation can be seen by plotting the
calculated intake temperature versus the vapor pressure as in
Figure 3.
The result of this dew-point limiting temperature can also be
seen in the volumetric efficiency changes (CVE) in Table 3
(positive is gain, negative is loss). At location 2, we see
a 4.1 percent improvement in volumetric efficiency with 75
percent of the fuel evaporated. Whereas at 100 percent
evaporated, the volumetric efficiency improvement drops to
only 2.4 percent. This is because the temperature of the mix
must be raised from the 75 percent evaporation dew point to
the dew point temperature necessary to achieve evaporation of
100% of the fuel.
Comparing the calculated volumetric efficiency improvements
in Table 3 to the actual results in Table 4, we see there is
a sizeable difference. At an equivalence ratio of 0.8, the
calculated results for location 1 predict a 2.5 percent gain,
while the actual results show a 2.8 percent loss. The
numbers are a 3.8 percent gain, and a 4.1 percent loss for
location 2. Obviously the results do not match, and further
investigation was in order.
-------
-15-
Phi
,• **
1.0
0.8
Table 3
Volumetric Efficiency Change "Without"
Heat Transfer from the Engine Structure
X
.25
.50
.75
1.00
.25
.50
.75
1.00
CVE* (%)-Location 1
0.7
1.5
2.4
2.9
0.6
1.2
1.8
2.5
CVE* (%)-Location 2
1.2
2.6
4.1
2.4
0.9
2.0
3.1
3.8
*CVE is calculated by Equation AIII - 22 in Appendix III
**Represents overall equivalence ratio, the local equivalence
ratio at location 1 was 20% of th overall ratio, and 33.3% of
the overall ratio at location 2.
Table 4
Average* Volumetric Efficiency
Change (Test Results)
Phi
i **
0.8
CVE (%)-Location 1
-2.8
CVE (%)-Location
-4.1
*See Appendix IV.
**Several different equivalence ratios
Generally they ranged between 0.7 and 0.9.
were
tested,
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-16-
B. Heat Transfer
By noting the probability that only marginal evaporation was
likely to take place under some conditions (from the 20% and
33% fuel split at the fumigation injectors), the physics
seemed to imply that as more fuel was added to the fumigation
injectors, the amount of the fuel that could be vaporized was
reduced. The apparent reason for this was that the
additional fuel flow (either from a larger flow split or a
richer fuel air ratio) caused the partial pressure of the
fuel to increase which in turn raised the dew point
temperature necessary for evaporation. in the extreme, if
all of the fuel were to be introduced at the front end of the
intake manifold (as in a carburetor) the partial pressure of
the fuel would be considerably greater (from 2.8 to 4.5 times
greater*) than the cases analyzed. In order for total
evaporation to occur under these conditions, the final intake
air temperature cquld not be below 45° to 50°F (assuming 15
in. HgA MAP). However, since vaporizing the total fuel flow
would depress the inlet air temperature approximately 300°F,
and assuming an initial inlet air temperature of around 80°F,
a discrepancy of some 265°F exists that must be made up by
heat transfer from some source before the fuel can fully
vaporize. Equation AII-1 indicated that the addition of heat
during evaporation would cause the final intake air
temperature depression to be less than in the absence of heat
addition. Therefore, heat transfer becomes a logical
parameter to be investigated in resolving the difference
between the expected results and the measured results.
Since heat transfer would raise the temperature of the mix,
it would be useful to assume several recovered charge-air
temperatures in order to identify if further investigaton
would be fruitful. Then we can perform the heat transfer
calculations to identify if there could be a sufficient
temperature differential to create the assumed recovery
temperatures. The results in Table 5 speak for themselves.
The volumetric efficiency losses from heat transfer resemble
the actual losses observed (Table 4) much closer than the
calculated gains in Table 3. Also, if we were to carburet
the entire amount of fuel, as opposed to the fractional
fumigation used in these tests, a sizable loss in volumetric
efficiency would occur (assuming sufficient heat transfer to
vaporize all of the fuel).
*Compare the mole fraction (mf 100) for 4 injectors (4 total)
which would be all of the fuel to the mole fraction for 1
injector (5 total) and 2 injectors (6 total) for the two
equivalence ratios in Appendix I Section C.
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-17-
It is interesting to note that the calculated percent loss
for volumetric efficiency remains constant for apparently any
temperature above the original air temperature (assumed to be
80°F for location 1, and 90°F for location 2). The mechanics
of this anomoly (constant VE loss with increasing
temperature) are that the comparisons are made with all
components at the same recovery temperature. Therefore, the
Table 5
Voluemtric Efficiency Change "with"
Heat Transfer from the Engine Structure
Resulting in an Assumed Charge-Air Temperature
Assumed Charge-
Air Temperature
-Partial Fumigation*
80°F
100°F
120°F
80°F
100°F
120°F
-Carbureted*
100°F
i *
Phi
1.0
1.0
1.0
0.8
0.8
0.8
1.0
0.8
CVE (%)
Location 1
-2.9
-2.9
-2.9
-2.3
-2.3
-2.3
•12.8
•10.5
CVE (%)
Location 2
-3.9
-4.7
-4.7
-3.0
-3.8
-3.8
•13.7
•11.4
*Theequivalence ratio for partial fumigation represents the
overall equivalence ratio, the local equivalence ratio at
location 1 is 20% of the overall ratio, and 33.3% of the
overall ratio for location 2. For the carbureted condition
the local and the overall equivalence ratios are the same.
For both the fuel delivery senerios, the pressure at
location 2 is approximately 15 in. Hg of manifold vacuum.
total volume of fuel and air at 120 °F is compared to the
volume of air with no fuel at 120°F, not at 80°F (the
original inlet temperature). The comparison at the elevated
temperatures is made because it is assumed that any heat
transfer that would heat the mixture to some temperature
above the original air temperature would also heat the intake
air in the baseline case (which did not have fuel in the
mixture) to the same temperature. The reason the base case
temperature would not be higher, is that in the base case the
temperature gradient between the duct wall and the intake air
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-18-
is very small. Hence little heat is transfered. However,
when methanol is injected, the intake air is cooled creating
a reasonably large temperature gradient which allows much
more rapid heat transfer. Once the cooled fuel-air mixture
has returned to the base case, then any temperature gradient
that would have affected the base case would also affect the
fuel-air mixture in a similar fashion because the specific
heats of air and methanol vapor are very similar. Hence, the
conclusion that the volumetric efficiency change between air
alone and the fuel-air mixture must be based on the same
recovery temperature.
Because the investigation of heat transfer effects began
after the tests were run, many of the temperature
measurements that would be useful for the heat transfer
calculations were not available. Therefore, many of these
temperatures were assumed based on other factors. The issue
here is to identify if rough heat transfer approximations
with reasonable temperature assumptions can account for any
of the discrepancy between the expected volumetric efficiency
change and the measured change. As it turned out, even with
extremely unsophisticated heat transfer approaches, heat
transfer does account for a majority if not all of the
difference between the expected and the observed.
The first assumption was that the only heat transfer that
occurred, occurred between the intake tract wall and the
vapor phase methanol-air mix. By using this assumption, we
were able to apply equation (13.2 - 25) from reference 4 for
convection in tubular heaters (this equation is repeated as
AV-1. The various boundary assumptions used for the analysis
are listed in Appendix VI.
The first case analyzed was for the single injector upstream
of the inlet to the EFI system (ie. upstream of the air
filter). The total distance from single injector to the
throttle body mounted on the intake air collector was
approximately 80 inches.
Since nearly all of this distance was surrounded by
atmospheric temperature, it was assumed that at least the
initial portion of the inlet system was at room temperature
(approximately 80°F). It was further assumed that the
temperature gradually increased along the length of the duct
to a final temperature of 160°F. This temperature rise is
considered to be an upper limit for estimation purposes for
the following reasons. At about the 40 percent point in the
intake tract, the intake air passes through a large box-like
structure that houses an air filter. In leaving the air
filter box, the intake air passes through a spring loaded air
meter device. The box and air meter are metal and are in the
-------
-19-
vicinity of the exhaust manifold (within 10-12 inches). It
is assumed some radiative heat transfer occurrs to the air
filter/metering system. However, no attempt was made to
correct for any of the potential heat transfer improvements
due to the proximity of the exhaust manifold to the
filter/metering box, to the increased surface area in the air
filter, or to the increased residence time in the air box
other than to assume the above temperature gradient along the
duct. Following the air box was a long section of stamped
steel duct work in close proximity to the cylinder head. And
finally, a short section in front of and including the
throttle valve was heated by engine coolant. The coolant was
maintained at about 180°F on the engine dynamometer test
stand.
By applying these boundary conditions to the tubular heater
equation, the length of the intake tract necessary to recover
the original air temperature (after the temperature
depression due to evaporation) varies between a low estimate
of 85 inches (phi = 0.8) to a high value of 137 inches (phi =
1.0). The variation in these calculated estimates is due to
a range of assumed hydraulic diameters of the irregularly
shaped intake tract. The measured distance on the hardware
was approximately 80 inches.
The discrepancy between the measured tract length and the
calculated range can be attributed to several reasons.
First, mentioned before, is that more heat transfer could
have occurred in the air filter box than was accounted for in
the calculations. Second, and most likely, is the estimation
procedure itself. All of the temperatures were estimated and
not measured (note: if 100°F had been used for the final
wall temperature (To2) instead of 160°F, the calculation
would predict slightly less than a doubling in length).
Probably more important, however, is the procedure only
considered convective heat transfer to the gaseous
methanol-air mixture. Heat transfer with a change in phase
(fuel evaporation) can be very complex. The effects on the
system due to convective heat transfer on any liquid fuel
droplets that might have impinged on the wall could have had
a noticeable influence on the total heat transfer rates.
This effect, however, was not considered in the rough
approximation. Finally, there is the additional possibility
that some other unknown parameter influenced the volumetric
efficiency results.
Even though some people might consider the discrepancy
between the calculated and measured results to be minor
(i.e., substantially less than one magnitude of difference),
investigation into the results at Location 2 might support
one of the suggested reasons for the discrepancy — more
-------
-20-
specifically, the issue of neglecting the influence of heat
transfer to liquid fuel droplets. At Location 1, a small
amount of fuel was sprayed into the center of a large
diameter duct (4.5 inches) approximately 6 to 8 inches
upstream from the entrance to the normal inlet duct-work. At
Location 2, 65 percent more fuel was sprayed into an air
collector about 2 to 3 inches from the entrance of four 1.35
inch intake runners. Because of the physical set-up, more
liquid droplets would be expected to impinge on the duct
walls at Location 2 than at Location 1. If the heat transfer
effects from the droplets were significant, one would expect
the simple calculation procedure (i.e., neglecting droplet
effects) to show less agreement at Location 2 than at
Location 1.
The simple calculation predicts a path length for Location 2
of between 31.9 and 32.2 inches (phi = 0.8 and 1.0
respectively). The approximate measured distance is 14
inches. The over estimation by the simple approach is about
128% in this case (note: the over estimation value would
increase if lower wall temperatures were used). At Location
1, the over estimation for the smallest assumed duct diameter
is only about 6%, whereas the length of the largest assumed
diameter is over estimated by around 58%. From this
comparison, it is difficult to scientifcally say that the
simple approach is more accurate at Location 1 than at
Location 2. Hence, the hypothesis that liquid fuel droplets
impinging on the intake walls increased the heat transfer can
not be completely confirmed. However, since the convective
heat transfer coefficient (h) for boiling liquids is
approximately 100 times that for forced convection of gases,
and about 5 times that of gases for viscous fluids(4)(5),
there certainly is room for engineering judgement to expect
that any fuel droplets on the walls would substantially
enhance the heat transfer. Enhanced heat transfer would
shorten the effective length of the intake track necessary to
recover the original intake air temperature, thus bringing
the calculated lengths more in line with the actual hardware.
In conclusion, the effects of heat transfer appear to play an
important part in explaining the difference between the
expected VE results and the actual results. One may argue
that the estimated component temperatures are too high, or
that the study should have investigated the effects of
droplets on the walls more thoroughly, but in the final
analysis, the aggregate evidence says heat transfer plays a
role in the volumetric efficiency of the engine when methanol
is introduced into the inlet track. Furthermore, because of
-------
-21-
the difference in the heat of vaporization*, the heat
transfer role appears to be more dramatic for methanol than
gasoline, and therefore this role should be given more
consideration in the design of methanol fuel delivery systems.
*Obert(6) provides a procedure to calculate the dew point of
gasoline (a multi-component mixture). Between 10% and 90%
evaporated, the dew point temperature is very similar between
gasoline and methanol. However, due to differences in the
heat of vaporization and the specific heats, methanol will
have an adiabatic temperature drop eight times (on the
average) greater than for gasoline. Therefore, to maintain a
similar dew point mixture temperature in the intake manifold,
the heat transfer required for the methanol engine would be
substantially greater than for the gasoline engine.
-------
-22-
References
(1) Results of Methanol Fumigation Investigation, EPA memo,
April 22, 1983, B. Michael and W. Clemmens to C. Gray,
reprinted in Appendix VII.
(2) The Internal Combustion Engine in Theory and Practice,
Volume 1, second edition, C.F. Taylor, The MIT Press, MIT
Cambridge, Massachusetts, pages 183-185.
(3) Chemical Engineers Handbook,
Perry, 1973, p. 3-56.
5th edition, Robert H.
(4) Transport Phenomena, R. Byran Bird, Warrren E. Stewart,
Edwin N. Lightfoot, John Wiley and Sons, 1960, Library of
Congress (60-11717), Figure 13-2, p 400, and pp 398-407 which
is referenced to E.N. Sieder and G. E. Tate, Ind. Eng. Chem.
"28", 1429-1435 (1936).
(5) Principles of Heat Transfer, second addition, Frank
Kreith, International Textbook Company, Scranton, PA, 1967,
p. 15.
(6) Internal Combustion Engines and Air Pollution, Edward F.
Obert, Intext Educational Publishers, New York, 1973, pp.
254-263.
(7) Aeronautical Vest-Pocket Handbook, Pratt & Whitney
Aircraft, Eleventh Edition, Eighteenth Printing, May 1966.
-------
-23-
Appendix I
Partial Pressure Calculation
I. Basic Equation
Phi = 1
(AI-1) CHsOH + 1.5(02 + 3.73 N2 + .04 Ar) -* C02 + 2H20 + 1.5 (3.73N2 + .04Ar)
II. Molecular Weights
CH30H = 32.043 N2 = 28.014
02= 32 Ar = 39.948
III. Partial pressure (Pv) of Methanol
A. Equation for mole fraction with 100% vaporization
(AI-2) (mflOO) = (z)+(1.5)+(i.g)(1.73)+(1.5)(.04)
where z = % of the total fuel used at specific location
times 10~2
mf = mole fraction
or
(z)
(AI-3) (mflOO) =
(z) + (7.155/PHI)
where Phi = (f/a) actual/(f/a) stoichiometric = equivalence ratio
B. Equation for mole fraction wih x% vaporized
(AI-4) mfx = (Z)(x) + (7.155/PHI)
where x = % of z amount of fuel that is vaporized times 10~2
(e.g. 25, 50, 75, 100)
-------
-24-
C. Values from equation AI.4.
Configuration Z Phi mf 25 mf 50 mf 75 mf 100
Cl
C2
C3
*
. 1
1
4
4
. 2
2
4
4
. 4
4
injector*
injector*
injectors"1"
injectors*
injectors*
injectors*
injectors"1"
injectors"1"
injectors"*"
injectors"1"
(5
(5
(5
(5
(6
(6
(6
(6
(4
(4
total)
total)
total)
total)
total)
total)
total)
total)
total)
total)
.2
.2
.8
.8
.333
.333
.667
.667
1.0
1.0
1
.8
1
.8
1
.8
1
.8
1
.8
.007
.006
.027
.022
.012
.009
.023
.018
.034
.027
.014
.011
.053
.043
.023
.018
.045
.036
.065
.053
.021
.015
.077
.063
.034
.027
.065
.053
.095
.077
.027
.022
.101
.082
.044
.036
.085
.069
.123
.101
Fumigation Injector
Port Injector
D.
Partial
Pressure
(pv)
(AI-5) (pv) = (mfx)(local absolute pressure)
-------
-25-
Appendix II
Temperature Drop with Vaporization
I. Injector Location
A. Location 1
1. Upstream single injector with 4 port injectors.
2. Fuel flow split: 20% at Location 1, 80% at the
normal port injectors
3. PI = 29 in. HgA (736.6 mm HgA) , Tl = 80eF
(26.7°C)
4. Overall phi ratios examined = 1.0, 0.8
B. Location 2
1. Two air collector injectors with 4 port
injectors.
2. Fuel flow split: 33.3% at Location 2, 66.7% at
the normal port injectors
3. P2 = 15 in. HgA(381.0 mm HgA), T2 = 90°F
(32.2°C)
4. Overall phi ratios examined = 1.0, 0.8
II. Calculations
A. Temperature drop (AT) Equation for x% evaporated(2)
(AII-1) AT = C(x)(F)(HLG) + (Q)]/(1-F + xF)(Cp)
where
x = % evaporated = 0.25, 0.5, 0.75, 1.0
F = (f/a) = (flow split)(.155)(phi)
Cp = .240 for F less than 0.5; = .245 for F greater
than 0.5 (assumption for approximate calculations)
HLG = heat of vaporization = 474 BTU/lb.
AT = °F
Q = zero for the first stage of analysis
-------
-26-
1. Location 1
phi
1.0
1.0
1.0
1.0
.8
.8
.8
.8
2.
phi
1.0
1.0
1.0
1.0
1.0
.8
.8
.8
.8
X
.25
.5
.75
1.0
.25
.5
.75
1.0
Location 2
X
.25
.5
.75
.78
1.0
.25
.5
.75
1.0
AT1(°C)
8.7
17.3
25.7
34.0
6.9
13.8
20.5
27.2
AT2(°C)
14.7
29.1
43.0
44.7
56.6
11.7
23.1
34.3
45.3
f/a = (.2)(.155)(phi)
Tl-ATl(°C) Pv(mm) TDEW(°C)*
18.0
9.4
1.0
-7.3
19.8
12.9
6.2
- .5
f/a = ( .333)
T2-AT2(°C)
17.5
3.1
-10.8
-12.5
-24.4
20.5
9.1
-2.1
-13.1
5.2
10.3
15.5
19.9
4.4
8.1
11.8
16.2
(.155) (phi!
Pv(mm)
4.6
8.8
13.0
13.3
16.8
3.4
6.9
10.3
13.7
-25.6
-15.9
- 9.8
- 5.9
-27.8
-19.4
-13.9
- 9.1
TDEW(°C)*
-27.2
-18.2
-12.5
-12.1
- 8.6
-31.2
-21.7
-15.9
-11.7
* Calculated from: Log10 Pv = [-0.2185(A)/K]+B
A = 8978.8 Range: -44°C to 224°C
B = 8.639821
K = °K
Source: CRC Handbook of Chemistry and Physics, 53rd
edition, 1972-1973.
B. Specific Volume of Air
1. Location 1
v = RT/p
where
p = (29.0) in hg (70.73) = 2051.2 psf
T = (459.7 + 80) - ATI = (539.7° - ATl)°R
R = 53.345 ft-lb /°R
v = specific volume = cu ft/lb air
-------
-27-
phi x T (air, °R) v(cu ft/lb air)
^"-_ --
1.0
1.0
1.0
1.0
1.0
1.0
1.0
1.0
.8
.8
.8
.8
.8
.8
.8
.8
0
25
50
75
100
with heat transfer
100
100
100
0
25
50
75
100
with heat transfer
100
100
100
539.7
524.0
508.6
493.5
481.1
to 80°,100°,
539.7
559.7
579.7
539.7
527.3
514.9
502.9
490.8
to 80°, 100
539.7
559.7
579.7
14.04
13.63
13.23
12.83
12.51
and 120°F
14.04
14.56
15.08
14.04
13.71
13.39
13.08
12.76
°,and 120°F
14.04
14.56
15.08
2. Location 2
v = RT/p
where
p = 15 in HgA (70.73) = 1061.0 psf
T = (459.7 + 90°)-AT2 = (549.7 ° - AT2)°R
R = 53.345 ft-lb/°R
-------
-28-
phi x Tair
1.0
1.0
1.0
1.0
1.0
with heat
1.0
1.0
1.0
0.8
0.8
0.8
0.8
0.8
with heat
0.8
0.8
0.8
0
25
50
75
100
transfer
100
100
100
0
25
50
75
100
transfer
100
100
100
C. Specific Volume of
549.7
523.2
497.3
472.3
476.2
to 80°, 100
539.7
559.7
579.5
549.7
528.6
508.1
487.9
470.6
to 80°, 100°,
539.7
559.7
579.7
Fuel
27.6
26.3
25.0
23.7
23.9
°,and 120°F
27.14
28.14
29.15
27.6
26.6
25.5
24.5
23.7
and 120°F
27.14
28.14
29.15
1. Location 1
v = (RT/p) vapor
where
p = (29.0) in hg (70.73) = 2051.2 prf
T = (459.7 + 80) -ATI = (539.7° -ATl)°R
m = 32.043
mR = 1544 ft-lb/°R
v = specific volume = cu ft/lb of vapor
-------
-29-
1. Case 1
phi
1.0
1.0
1.0
1.0
1.0
1.0
1.0
1.0
0.8
0.8
0.8
0.8
0.8
0.8
0.8
0.8
Location
0
25
50
75
100
with heat
100
100
100
0
25
50
75
100
with heat
100
100
100
2
539.7
524.0
508.6
493.5
481.1
transfer
539.7
559.7
579.7
539.7
527.3
514.9
502.9
490.8
transfer
539.7
559.7
579.7
12.68
13.15
13.62
12.68
12.39
12.10
11.81
11.53
12.68
13.15
13.62
v = RT/p
where
p = 15 in HgA(70.73) = 1061 psf
T = (459.7 + 90)-AT2 = (549.7° - AT2)°R
m = 32.043
mR = 1544
-------
-30-
phi
1.0
1.0
1.0
1.0
1.0
1.0
1.0
1.0
0.8
0.8
0.8
0.8
0.8
0.8
0.8
0.8
Compute
0
25
50
75
100
with heat
100
100
100
0
25
50
75
100
with heat
100
100
100
mass Volume
549.7
523.2
497.3
472.3
476.2
transfer
539.7
559.7
579.7
549.7
528.6
508.1
487.9
470.6
transfer
539.7
539.7
579.7
25.0
23.8
22.6
21.4
21.6
24.51
25.42
26.33
25.0
24.0
23.1
22.2
21.4
24.51
25.42
26.33
D.
V = (mass air)(vl)+(mass fuel)(x)(v2)+(Mass fuel)(1-x)(v3)
where
mass air = 6.452 Ib (phi = 1)
mass fuel = Location 1: 20% (1 Ib) = .2 Ib for
phi of 1.0; =.16 Ib for
phi of 0.8
= Location 2: 33.3%(1 Ib) = .333 Ib
for phi of 1.0; = .266
Ib for phi of 0.8
= Carbureted : 1 Ib for phi = 1.0,
0.8 Ib for phi = 0.8
x = % of fuel evaporated = 0, .25, .50, .75, and 1.0
vl= Specific volume air (cu ft/lb)
V2= Specific volume evaporated fuel (cu ft/lb)
V3= Specific volume liquid fuel = .0201 cu ft/lb
V = Cubic feet of mixture per pound of fuel consumed
vc= Specific Volume of gaseous portion of mixture
VG= V/[(mass air) + (x) (mass fuel)]
-------
-31-
Table AII-D-1
Specific Volume of Partially
Fumigated Mix at Location 1
phi
1.0
1.0
1.0
1.0
1.0
1.0
1.0
1.0
0.8
0.8
0.8
0.8
0.8
0.8
0.8
0.8
X
0
.25
.50
.75
1.00
with
1.00
1.00
1.00
0
.25
.50
.75
1.00
with
1.00
1.00
1.00
Mvl
90.59
87.94
85.36
82.78
80.71
heat transfer
90.59
93.94
97.30
90.59
88.46
86.39
84.39
82.33
heat transfer
90.59
93.94
97.30
Mxv2
0
.62
1.20
1.74
2.26
80,
2.54
2.63
2.72
0
.50
.97
1.42
1.84
80,
2.03
2.10
2.18
M(x-l)v3
.004
.003
80.33
.001
0
100, 120°F
0
0
0
.0032
.0024
.0016
.0008
0
100, 120°F
0
0
0
V
90.59
88.56
96.56
84.52
82.97
93.13
96.57
100.02
90.59
88.95
87.36
85.81
84.17
92.61
96.04
99.48
Specific
Volume of
Mix (vc)
14.04
13.62
13.21
12.80
12.47
14.00
14.52
15.04
14.04
13.06
13.37
13.06
12.73
14.01
14.53
15.05
-------
-32-
Table AII-D-2
Specific Volume of Partially
Fumigated Mix at Location 2
phi
1.0
1.0
1.0
1.0
1.0
1.0
1.0
1.0
0.8
0.8
0.8
0.8
0.8
0.8
0.8
0.8
X
0
.25
.50
.75
1.00
with
1.00
1.00
1.00
0
.25
.50
.75
1.00
with
1.00
1.00
1.00
Mvl
178.08
169.69
161.30
152.91
154.20
Mxv2
0
1.98
3.76
5.34
7.19
heat transfer to
175.08
181.56
188.05
178.08
171.62
164.53
158.07
152.91
8.16
8.46
8.77
0
1.60
3.08
4.44
5.69
heat transfer to
175.08
181.56
188.05
6.52
6.76
7.00
M(x-l)v3
.0067
.0050
.0033
.0017
0
80°, 100°,
0
0
0
.0054
.0040
.0027
.0013
0
80°, 100°,
0
0
0
V
178.08
171.67
165.06
158.25
161.40
and 120°F
183.24
190.03
196.82
178.08
173.22
167.61
162.51
158.60
and 120°F
181.60
188.32
195.05
Specific
Volume of
Mix (vc)
27.60
26.27
24.94
23.61
23.79
27.01
28.01
29.01
27.60
26.57
24.45
24.43
23.61
27.03
28.03
29.03
-------
-33-
Table AII-D-3
Specific Volume of Carbureted Mix
with Complete Fraction of Fuel and
a Charge-Air Temperature of 100°F
phi v Mvl Mxv2 M(x-l)v3 V Vc VB
Location 1
1.0
0.8
1.00
1.00
93.92
93.92
13.15
10.52
0
0
107.06
104.43
14.37
14.40
14.56
ii
Location 2
1.0
0.8
1.00
1.00
181.56
181.56
25.42
20.33
0
0
206.98
201.90
27.78
27.84
27.60
«
-------
-34-
Table AII-D-4
Specific Volume of Partially Fumigated
Mix (all locations) vs. Fraction Evaporated
Phi
1.0
11
II
II
ii
.8
M
ll
II
tl
with
1.0/80°F
1.0/100°F
1.0/120°F
0.8/80°F
0.8/100eF
0.8/120°F
X
0
.25
.50
.75
1.00
0
.25
.50
.75
1.00
Location
VB
14.04
M
11
11
II
14.04
II
II
M
II
heat transfer (assumed
100
II
ti
100
II
II
14.04
14.56
15.08
14.04
14.56
15.08
1
Vc
14.04
13.62
13.21
12.80
12.47
14.04
13.70
13.37
13.06
12.73
charge
14.00
14.52
15.04
14.01
14.53
15.05
Location
VB
27.60
M
II
II
II
27.60
ii
M
ii
M
temperature)
27.60
28.14
29.15
27.60
28.14
29.15
2
Vc
27.60
26.27
24.94
23.61
23.79
27.60
26.57
25.45
24.43
23.61
27.01
28.01
29.01
27.03
28.03
29.03
-------
-35-
Appendix III
Volumetric Efficiency Change Due to
_ Adiabatic Temperature Drop
In order to compare the effects of the adiabatically cooled
charge on volumetric efficiency, it would be useful to
perform this comparison at a constant manifold pressure
(which is also consistent with the analysis in Reference 1).
If we consider the Bernoulli equation of the form;
1) 0.5d1(V1)2 + P-L = 0.5d2(V2)2 + P2*
we can rearrange it to,
(AIII-2) PI - ?2 = 0.5d2(V2)2 - 0.5d1(V1)2
If we consider condition (1) to be at ambient conditions,
then the term (P^ - P2 ) would be our constant manifold
vacuum (MV) , and the V^ term would be zero. Therefore
Equation AIII-2 would become
(AIII-3) MV = 0.5d2(V2)2
Now we can evaluate the effects of charge-air cooling on the
velocity (V) term by forming a ratio of the baseline case (B)
to the cooled case (C).
I-4) MVC/MVB = [0.5 w(V2)2]c/[0.5 w(V2)2]B
Since MVB = MVc we can change (AIII-4) to,
(AIII-5) C(V2)2]C/C(V2)2]B = W
Also recognizing that the weight density (w) can be expressed
as the inverse of the specific volume, we have Equation
AIII-6 which says the velocity change in the system is
inversely proportional to the square root of the ratio of the
specific volumes.
(AIII-6) [V2]c = CV2]B (vc/vB)0-5
The mass flow change due to the charge cooling can now be
calculated from the information on the velocity change and
the specific volume change. This mass change can then be
related to a change in volume flow at the inlet to the
induction system which is at constant temperature and
pressure. First, the mass rate (M) can be described as,
*d = mass density = weight density/gravitational constant
= w/g.
v = velocity.
-------
-36-
(AIII-7) M = (d)(V)(A)
Next, we can look at the change in mass from the baseline
due to charge cooling (MC) by forming a ratio,
8) MC/MB = CdVA]c/[dVA]B
Cancelling the areas (A) in AIII-8, and substituting for YC
from AIII-6, we have
9) MC/MB = (dcVB)(vc/vB)0.5/(dBvB)
Substituting for d (d = w/g) and cancelling VB» we have
10) MC/MB = wc (vc/vB)°'5/wB
Substituting the specific volume (v) for the density (w) (w =
1/v),
(AIII-11) MC/MB = (vB/vc)(vc/vB)0-5
Equation AIII-11 describes the change in mass flow at
constant manifold vacuum due to charge cooling in terms that
we have calculated in Section D of Appendix II (specific
volume). Note that the specific volume of the cooled charge
(vc) is the specific volume for a mix of vaporized fuel
plus air while the specific volume of the base case (VB) is
for air alone. Therefore, the mass flow change includes the
change in the mass of air, and the change in the mass of
fuel.* In order to identify the change in volumetric
efficiency, we must separate out the change in the mass of
the intake air from the total mass change. This can be
determined essentially on a unit ratio basis where the
original air mass is unity, and the portion of the change in
air mass (M^c) is t^16 ratio of the mass of air (a) to the
sum of the mass of air (a) plus the mass of fuel that has
evaporated (xf).
(AIII-12) MAC = (Mc/MB) (a/a + xf)
or,
(AIII-13) MAC = (Mc/MB)[l/(l + xf/a)]
Since we considered the old air mass equal to unity, we can
now compute the change in air mass (CAM) between the two
conditions .
*The change in mass of fuel refers to the gaseous portion and
hence is essentially governed by the fraction evaporated.
Liquid fuel is neglected.
-------
-37-
(AIII-14) CAM = MAC - 1
or,
(AIII-15) CAM = (vB/vc)(vc/vB)0.5[a/(a + xf)] - 1
We can use this change in mass flow to compute the change in
volume flow at the inlet to the system (i.e., at the air flow
meter). Recognizing that the change in mass flow is uniform
across the system, the change at the inlet can be written as
a ratio of the difference between the change in air flow due
to cooling (MAc), and the baseline case, or
(Ain-16) CAM = (MAC - MB)/MB
Using the continuity Equation (AIII-7) and inlet conditions,
we can rewrite AIII-16 as,
(AIII-17) CAM = [(dVA)ACi - (dVA)Bi]/[dVA]Bi
At the inlet to the system, the temperature and pressure for
all practical purposes is the same between the charge cooled
case, and the baseline case because all of the potential
charge cooling would occur farther downstream. Hence the
inlet density for the charged cooled case is essentially the
same, and cancels in AIII-17 leaving only the velocity term
(V) and the area term (A). The product of the velocity (V)
and area (A) happens to be the volume flow rate (Q). The
result of substituting this identity into AIII-17 provide us
with a change in volume flow rate
(AIII-18) CAM = (QACi - QBi)/Qfii = CQi
If we define volumetric efficiency (VE) as inlet volume flow
rate (Qi) divided by engine displacement per unit time, the
change in volumetric efficiency (CVE) would be equivalent to
the change in volume flow rate (CQi).
(AIII-19) VE = Qi/(Displ./Time)
(AIII-20) CVE = CQi
(AIII-21) CVE = CAM
Substituting the conditions resulting from any charge-air
cooling for CAM (AIII-15) into equation AIII - 21, we now
have a function for the change in volumetric efficiency (CVE)
as a function of the specific volume resulting from the
temperature change due to fuel evaporation, and the amount of
fuel that has evaporated.
(AIII-22) CVE = (vB/vc)(vc/vB)°-5[a/(a + xf)] - 1
-------
-38-
Appendix IV
Observed Volumetric Efficiency Loss
The volumetric efficiency (VE) change is based on a constant
pressure drop across the throttle (i.e., constant manifold
vacuum, MV). Analysis of the test results from Figure 3 in
Appendix VII indicate the following relationships.
Baseline: VE = -2.43 MV -I- 70.50 (RSQ = .96)
Location 1: VE = -2.57 MV + 71.17 (RSQ = 1.0)
Location 2: VE = -1.99 MV + 63.44 (RSQ = .77)
Substituting a range of manifold vacuums from 10 to 15 inches
Hg, we arrived at the following tables:
Table AIV-1
Volumetric Efficiency vs. Manifold Vacuum
MV
10
12.5
15
Baseline
34.05
40.13
46.2
Location 1
32.62
39.05
45.47
Location 2
33.15
38.13
43.10
Table AIV-2
Volumetric Efficiency Changes (from Base)
MV CVE (%)-Location 1 CVE (%)-Location 2
10 -1.6 -6.7
12.5 -2.7 -5.0
15 -4.2 -2.6
Average Loss 2.8 4.1
-------
-39-
Appendix V
Heat Transfer Calculation
I . Equation for correlation coefficient (C) by Sieder and
Tate (4)"
(AV-1) (Tb2 - Tbl)(D)(Pr)*667(u)~°-14/4(L)[(To - Tb)ln] = C
where
[(To - Tb) In] = (A - B)/[ln(A/B)]
A = Tol - Tbl
B = To2 - Tb2
Pr = (Cpu/k) @ Tb
u = (u @ Tb/u @ To)
II . Boundary Conditions
A. Boundary Conditions for All Locations
k)
1 @
60°F (ref. 2)
1. Air flow = 19.6 CFM = 1176.4 CFH
2. k = .014 (approx. = air) (conduction
3. Cp = .240 for air; = .245 for phi of
a. Cp = (R/J)[k/k-l]
b. R = mR/m = 1544/m
c. m = (% fuel) (phi) (32.043) + [!-(% fuel) (phi )] (28 .85 )
d. k = 1.4 for air (ratio of specific heats)
e. k = 1.38 for menthanol/air at phi = 1 (ref. 2)
4. Geometry similar to tubular heater
5. Re, Pr , Nu numbers computed at bulk temperature
(Tb) = (.5) (Tbl + Tb2)
B. Boundary Conditions for Location 1
1. For phi = 1
a)
b)
c)
d)
e)
f)
g)
h)
Tbl
Tb
ub
uo
Cp
Tol
To2
Tb2
= 19°F
= 50°F
= 3.7 (E-7)
= 3.9 (E-7)
= .240
= 80°F
= 160°F
= 80°F
(32.17) (3600) Ibm/
(32.17) (3600) Ibm/
ft.hr
ft.hr
@ 50eF
@ 80°F
i) Re neglects fuel mass flow.
For phi = .8
a)
b)
c)
Tbl = 31
Tb =56
ub = 3.7
(E-7) (32.17) (3600) Ibm/ft.hr @ 50°F
-------
-40-
d) uo = 3.9 (E-7) (32.17) (3600) Ibm/ft.hr @ 80°F
e) Cp = .240
f) Tol = 808F
g) To2 = 160°F
h) Tb2 = 80°F
i) Re neglects fuel mass flow
C. Boundary Conditions for Location 2
1. For phi = 1
a) Tbl = 7.5°F
b) Tb2 = 908F
c) Tb = 48.7°F approx = 50°F
d) ub = 3.7 (E-7) (32.17) (3600) Ibm/ft.hr @ 50°F
e) uo = 3.95 (E-7) (32.17) (3600) Ibm/ft.hr @ 90°F
f) Tol = 180°F
g) To2 = 240°F
h) Re neglects fuel mass flow
i) Cp = .241 for k est. @ 1.395
2. For phi = .8
a) Tbl = 8.4°F
b) Tb2 = 90°F
c) Tb = 49.2°F approx = 50°F
d) ub = 3.7 (E-7) (32.17) (3600) Ibm/ft.hr @ 50°F
e) uo = 3.95 (E-7) (32.17) (3600) Ibm/ft.hr @ 90°F
f) Tol = 180°F
g) To2 = 240°F
h) Re neglects fuel mass flow
i) Cp = .241 for k est. @ 1.395
-------
-41-
III. Calculations
A. Location 1, phi = 1
0 L/D = [(Tb2 - Tb)(Pr)-667(u)-0.14]/4(c)[(To - Tb)ln]
_(Pr).667 = (Cp u/fc.).667 = C( .240) (3.7)(E-7) (32.17) (3600)/(0.
- (u)-°-14 = (ub/uo)-0.14 = (3.7/3.9)-0.14 = 1.0074
0 [(To - Tb)ln)] = (A - B)/[ln(A/B)] = 70.1
- A = Tol - Tbl = 80-19 = 61
- B = To2 - Tb2 = 160-80 = 80
8 Re = Dvd/u
Where
d = slugs/cu. ft. = 2.3(E-3) slug/cu. ft. @ 59°F
v = ft./sec.
u = Ib.-sec./sq. ft.
D = duct diameter.
0 For various estimates of intake duct hydraulic diameter
D (in.) D ft.) A (sq.ft.) v (ft.sec.) Re
2
2.5
3.0
.167
.208
.25
.022
.034
.049
14.989
9.586
6.662
1.56(E+4)
1.28(E+4)
1.04(E+4)
.00375
.00380
.00395
L = (D/C)(K)(Tb2 - Tbl)/4[(To - Tb)ln]
- K = .8021
For various estimates of intake duct hydraulic diameter
Estimate of L L (hardware
D (in.) Required to Achieve Tb2 measurement)
2.0" 95.2" 80"
2.5" 117.4" 80"
3.0" 137.2 80"
-------
-42-
B. Location 1, phi .8
Pr, Re, C, and u are the same as in Case 1, phi = 1.0
[(To - Tb)ln] = (A - B)/[ln(A/B)] = 63.2
- A = Tol - Tbl = 80-31 = 49
- B = To2 - Tb2 = 160-80 = 80
For various estimates of intake duct hydraulic diameter
Estimate of L L (hardware
D (in.) Required to Achieve Tb2 measurement)
2.0" 84.8" 80"
2.5" 104.6" 80"
3.0" 122.3 80"
C. Location 2, phi = 1
D of intake track runner = 1.35 in. = .1125 ft.
Q per runner = Q/4 = (.327 ft.3/sec.)/4 = .0818 ft.3/sec.
v = 8.22 ft./sec.
Re = dvD/u = 5.748 E+3
C = .004
pr.667 = .7376
(U)-.14 = 1.009
K = (Pr)-667 (U)-.14 = .7442
[(To - Tb)ln] = (A - B)/[ln(A/B)] = 161
- A = Tol - Tbl = 172.5
- B = T02 - Tb2 = 150
L = (D/C)(K)(Tb2 - Tbl)/4[To - Tb)ln]
L (calculations) = 32.2 in.
L (hardware measurement) = 14 in. (approx.)
-------
-43-
D. Location 2, phi .8
Pr, Re, C, D, and u are the same as in Case 2, phi .8
[(To - Tb)ln] = (A - B)/[ln(A/B)] = 160.5
- A = Tol - Tbl = 171.6
- B = To2 - Tb2 = 150
L (calculated) = 31.9 in.
L (hardware measurement) = 14 in. (approx.)
-------
-44-
Appendix VI
Changes in Pumping Work Due to Temperature Changes
It is generally accepted that if the air-charge of an engine
is cooled, the volumetric efficiency is increased. The
question is then what is the general magnitude of the
differences between the work to pump cold air versus hot air.
As indicated in other appendices, if fuel is used to cool the
charge, the volumetric efficiency may or may not improve.
For the purposes of this appendix, we will assume that there
is no fuel in the air charge, and that the air charge is
cooled by some means other than fuel evaporation.
The simple model we will use is a long tube of constant cross
section connected to a pump which exits to the atmosphere.
In order to consider only the work necessary to pump a given
quantity of fluid, and not the work to speed up the air, we
will assume the pump exit area to be the same as the long
inlet tube. The following conditions will be assumed:
Position 1 (inlet)
P! = Pa
Vi = 0
Position 2 (inside tube)
V2 = 9.586 ft/sec (from Section III, Appendix V)
D2 = 2.5 in = .208 ft (from Section III, Appendix V)
Position 3 (Exit)
P3 = Pa = PI
V3 = V2
D3 = D2
Equation AVI-1 (pgs. 213, 216, Reference 4) is a restatement
of the Bernoulli equation, and in effect says that the work
needed to pump a given amount of fluid across our simple
system from position 1 to position 3 is equal to the work
needed to change the velocity, plus the work needed to change
the hydraulic head, plus the work needed to change the
pressure, plus the work needed to overcome friction.
(AVI-1) W = (.5)V2 - g h-(RT/M)ln(P3/P1)-(.5V2Lf)/R
The change in velocity is simply the outlet velocity minus
the inlet velocity.
-------
-45-
(.5)V2 = .5(V3)2 - .
Vx = 0
V3 = V2
(AVI-2) (.5)V2 = .5(V2)2
The change in hydraulic head is zero.
(AVI-3) h = 0
The change in pressure is zero
Pi = P2
(AVI-4) In (Pi/P2) = In 1 = 0
If we look at the change in work per unit length at position
2, the unit work due to friction is
= (.5)(V2)2(D(f)/(.5)(D2)
(AVI-5) = (V2)2(f)/D2
Substituting AVI-2 to 5 into AVI-1, provides for the work per
unit length to pump fluid through our simple system,
(AVI-6) W = -(.5)(V2)2 - (V2)2(f)/D2
The friction factor (f) is related to the Reynold's No.
Re = DVd/u
d = P/RT*
(AVI-7) u = (.3170)(E-10)(T)l-5[734.7/(T+216)]**
From Appendix II, we can assume a dry air temperature of 80°F
at condition 2. The lowest dew point of the fuel air mixture
was around -12°C (10.4°F). Therefore, we will assume that we
can cool dry air to that temperature for this comparison.
Using these assumptions results in the following:
Temperature u d Re
80°F 3.86(E-7) 2.214(E-3) 1.42(E+3)
10.4°F 3.46(E-7) 2.542(E-3) 1.82(E+3)
* d = mass density = w/g.
**T = °R, from Reference (7).
-------
-46-
Since the Reynold's No.s calculated are in the laminar range,
the friction factor (f) becomes, .
(AVI-8) f = 16/Re
Therefore, the work Equation (AVI-6) becomes,
(AVI-9) W = -(.5)(V2)2 - (V2)2(16)/(D)(Re)
To determine the magnitude of the difference in work per unit
length to pump hot air (WH) versus cold air (We), a
simple ratio can be formed with Equation AVI-9, and the
appropriate values for position 2 and the Reynold's No. can
be substituted.
(AVI-10) WC/WH = Z
or,
(AVI-11) Wc = .9785 WH
Because the velocity (V2) and area (A2) remained the same
between the comparison of the required pumping work with hot
air to the required work with cold air, the results of
Equation (AVI-11) are applicable to constant volume
flow-rate. In this case, artifically reducing the
temperature of dry air by approximately 70°F, reduced the
estimated work per unit length by slightly more than 2
percent.
However, power is a function of mass of air flow, not
volume. If we want to hold mass flow (M) constant for this
exercise, then Equation AVI-13 tells us that mass density (d)
and velocity must vary with temperature.
(AVI-13) M = dVA
Since we have previously calculated the change in mass
density (d) with temperature, we can determine the subsequent
change in velocity by forming a ratio of the cold mass flow
rate to the hot mass flow rate. Cancelling terms, we have
(AVI-14) VC/VH = dH/dc = .87097
Substituting AVI-14 into AVI-9 and 10 results in the ratio of
work per unit length to pump the same mass flow at two
different temperatures.
(AVI-15) WC/WH = .87097 (Z)
-------
-47-
or,
(AVI-16) We = .8522 WH
The results of AVI-16 indicate that the work to pump the same
mass of cold fluid is approximately 15 percent less than that
required for a hot fluid. But remember that this is for air
that was cooled by some means other than fuel evaporation.
Also this was a very simple example that did not consider the
intricacies in the geometry of a real engine. The potential
effects of the practical engine cycle (intake, compression,
power, exhaust, and valve overlap) were similarly not
considered. And finally the myriad of effects involved in
pulsating flow were ignored. Nonetheless, the general trend
is that if you can cool the air by some means other than
adding fuel, it will require less power during the intake
stroke.
The amount of work reduction is, however, obviously not only
dependent on the percentage reduction, but also on the base
level of intake work. Furthermore, the relation of the
intake work to the other losses in the engine cycle must be
of sufficient magnitude, in order for a small percentage
improvement in intake work due to a cooler charge to be
measureable in the overall engine performance.*
*Note:enginepower increases due to a cooler intake charge
are essentially due to an increase in charge mass density
which increases the mass throughput of the engine.
-------
-48-
Appendix VII
EPA MEMO: Results of Methanol
Fumigation Investigations
-------
-49-
I
UNITED STATES ENVIRONMENTAL PROTECTION AGENCY
ANN ARBOR. MICHIGAN 48105
DATE: ADD « « iqnq
nrn C C. 1JO\I
SUBJECT: Results of Methanol Fumigation Investigations
OFFICE OF
A(R NOISE AND RADIATION
PROM
TO
THRU
R. Bruce Michael
Technical Support Staff
William Clemmens/ Project Manager
Technical Support Staff
Charles L. Gray, Jr., Director
Emission Control Technology Division
Phil Lorang, Chief
Technical Support Staf
<
The hypothesis of this study was: "Can methanol injected into
the upstream air passages (additionally to the normal port
injection) cool the inlet air sufficiently through vaporiza-
tion to increase the volumetric efficiency, and will the
expected increase in volumetric efficiency translate into
improved thermal efficiency?" This hypothesis is derived
from the mathematical equation of volumetric efficiency in
which the efficiency is proportional to the mass of air per
unit of time flowing through the engine relative to the swept
volume of the engine during that time. Vaporization (through
fumigation) in the inlet air should cause a temperature drop
of the air, which would increase density and mass flow. To
test this hypothesis, two methods of fumigation were tested,
both occurring along with the normal port injection of the
Nissan 2.0 litre NAPS-Z engine. Neither method improved
volumetric or thermal efficiency. In fact, the normal port
injection without fumigation gave slightly better results.
Because none of the fumigation systems tested demonstrated
any increase in efficiency at the lower power levels tested,
we recommend that we abandon any plans for fumigation testing
on this engine in the future.
Test Configuration
Fumigation of methanol was obtained by adding EFI style fuel
injectors upstream of the standard port injectors. The
fumigation injectors were controlled by the EFI computer in
-------
-50-
the same manner and simultaneously with the standard port
injectors. The standard port injectors are located approxi-
mately 3 to 5 inches from the intake valve seat.
Three locations for the fumigation injectors were tested.
The first location was in the intake manifold collector. The
collector is a log style manifold downstream of the single
throttle valve. Four intake runners enter the bottom of the
collector. They are ranged in pairs - cylinder 1 and 2, and
cylinders 3 and 4. The intake runners are on the order of 12
inches from the intake valve seat to the collector. The
single fumigation injector was mounted on the top side of the
collector, and roughly equally spaced between the intake
runner pairs of cylinders 1 and 2 and cylinders 3 and 4
(i.e., between cylinders 2 and 3).
The second location was also in the collector, but two
injectors were used instead of a single injector. The two
injectors were located over the intake runner pairs - one
injector over the pair of runners for cylinders 1 and 2, and
the other injector over the pair of runners for cylinders 3
and 4.
The third location utilized a single injector placed upstream
of the entrance to the EFI vehicle system. This injector was
centrally located in a short section of 4 inch O.D. tubing
connected to the downstream end of the intake air flow
meter. This configuration is identified as "central
injection" or "central fumigation".
Engine Operation
i
Engine operation with fumigation eminating from the first
location was not satisfactory. Rough engine running, extreme
sensitivity to ignition timing and ever present detonation
were characteristics of this location. We speculate that the
central location of the injector created a situation in which
cylinders 2 and 3 robbed fuel from cylinders 1 and 4 creating
a very lean condition in cylinders 1 and 4. Because of these
problems, all fumigation testing at this location was
terminated.
The second location for the fumigation injectors seemed (to a
degree) to solve the assumed distribution problems, but on a
qualitative basis the engine still did not operate quite as
well as in the standard configuration (i.e., no fumigation).
For instance, some of the tests at 90 ft.lb. and 2000 RPM
were scrubbed from the test program because of an instability
in the MBT point for ignition timing, which led to creeping
detonation. Another handicap with the two injectors was that
even with the adjustable air/fuel control box for the EFI,
the sum of the fuel flow from the two fumigation injectors
-------
-51-
plus the four standard injectors made it difficult to achieve
equivalence ratios leaner than 0.8. Operation at equivalence
ratios richer than 0.8 would normally depress the brake
thermal efficiency somewhat, but Table 1 and Figure 1 suggest
that peak efficiency with the two injector set-up is slightly
richer than 0.8. Directionally the shift to a richer point
from the baseline (0.7 to 0.8) for peak efficiency is in the
wrong direction.
Only the central injector seemed to operate as well as the
standard configurations. No difference in driveability was
observed. The only handicap encountered was that the central
fumigation system was potentially capable of running leaner
than our adjustable EFI control box would allow.
Results and Conclusions
The most interesting result of this testing was that upstream
injection (fumigation) of methanol did not improve volumetric
efficiency. In fact, as shown in Table 1 (and Figures 1 and
2), fumigation of methanol on this port injected engine
actually decreased volumetric efficiency by 4 to 10 percent.
In three out of four cases, the decrease in volumetric
efficiency resulted in a noticeable loss in thermal
efficiency (5-9%). The only case that showed an increase in
thermal efficiency (0.5 percentage point), with a decrease in
volumetric efficiency was the test point at high power using
the central upstream injector (Inlet to EFI, Table 1). The
trend of this increase in efficiency as shown in Figure 2
(central fumigation) is plateauing around an equivalence
ratio of 0.8, the limit of our EFI control with the central
injector. Potentially, improvement could be obtained at
leaner equivalence ratios, but returning to Figure 1, we see
that we were able to sufficiently enlean the system to get a
peak in the efficiency trend with the central injector. At
the lower power in Figure 1, the peak efficiency for the
central injector was below the peak efficiency for the
baseline.
It is not known exactly why fumigation did not increase
volumetric efficiency, but information from the recent Nissan
work may shed some light. Nissan was surprised to find that
efficiency went down with methanol, as compared to gasoline,
and theorized that fuel vaporization was caused mainly by
absorbing heat from the cylinder walls rather than the air.
This would, of course, result in little or no gain in
charging efficiency since the incoming air would not change
density. The lack of a density change results in a negative
effect on volumetric efficiency for the following reason.
When the fuel vaporizes, it increases the gas volume in the
air stream and has the effect of restricting or slowing down
the air flow. This lowers charging efficiency, more than
-------
-52-
offsetting any gain. While this is not exactly analogous to
our fumigation work, it may be that the restriction of air
flow (due to fumigation) was the dominant effect, which
lowered the volumetric efficiency.
-------
-53-
Table 1
Methanol Fumigation
Test Results
Injector Position
Intake Intake Inlet to
Collector Collector EPI
Baseline* 1 Injector 2 Injectors 1 Injector
90 ft-lbs @ 2000 RPM
Best Average Efficiency 38.9
Equivalence Ratio .79
Volumetric Efficiency 69.3
Number of Tests 9
29.5 ft-lbs @ 1500 RPM
Best Average Efficiency 26.4
Equivalence Ratio .70
Volumetric Efficiency 38.0
_**
35.5
.93
65.5
3
24.4
.86
33.8
39.4
.83
65.1
3
25.2
.78
35.5
Number of Tests
12
*Based on replicate tests before fumigation testing (V87,V90)
and replicates after fumigation (ZM2-VO). The best average
(thermal) efficiencies were calculated as follows. First,
the data were separated into different equivalence ratio
ranges for each version. Second, the average thermal
efficiencies of the tests in each range for each version was
calculated. Third, for the equivalence ratio range yielding
the overall best efficiencies, the average efficiencies for
the three versions were averaged. For the higher torque
level, the equivalence ratio range which yielded the highest
average efficiency was .77-.80. For the lower torque level,
the range was .67-.73.
**A11 testing was aborted in this configuration due to severe
detonation problems (no data taken) .
-------
-54-
1500BPM
35
30
20
'
6*i*«.;w*
•'-A^-J^M* i _/"
-* ** + ' r-4-K_
i i i 1 — 1
X BASELINE
N-15
» TWO OUTSIDE
INJECTORS
N-3
* CEMTRfiL INJECTOR
N-6
0 DISASSOCIRTION
BASELINE
N-9
.65 .70 .75 .80
PHI
.85 .90
.35
Methanol Fumigation
Figure 1
-------
-55-
2000RPM
z
UJ
40
35
30
25
20
.E
c*«7(\A<-
Far*i(.AriOtJ
")4ti|T)hr->^ '
G^Sii'-'vtf A.
* C.ENTRRL INJECTOR
N»S
0 -DISRSSOCIRTION
BftSELINE
N-9
5 .70 .75 .30 .85 .90 .95
PHI
Methanol Fumigation
Figure 2
-------
-56-
Purther analysis of the data was performed to compare
volumetric efficiency to manifold air pressure (manifold
vacuum). Theoretically, if the upstream fumigation of
methanol cooled the inlet air, the volumetric efficiency
should increase for the same manifold air pressure.
Figure 3 shows scatter plots and regression lines of
volumetric efficiency versus manifold pressure (MAP) for the
normal port injection (baseline) and the two fumigation
versions. The regression lines, which are all nearly
parallel, again show no advantage to fumigation.*
Figure 4 shows thermal efficiency (BTE) versus manifold
pressure for the same tests. In all cases, for a given MAP,
efficiency was higher in the baseline configuration, thus
suggesting that methanol fumigation does not improve
volumetric efficiency.
*We should mention that the MAP measurements were assumed to
be "dry" measurements (no partial pressure of the fuel in the
manifold), which is not really correct. The fumigation would
really result in a somewhat wet condition in the manifold,
making the measured MAP a wet measurement which is by nature
greater than a dry measurement. Therefore, the fumigation
points and regression lines really should be slightly to the
left of those shown in Figure 3 which translates into worse
results. No calculations were made to account for this,
because the effect is slight.
-------
o
z
UJ
*—«
CJ
52
50
48
46
44
42
UO
38
36
34
32
30
VOL. EFF. VS MflP
6
_L
10 11 12 13 14
MANIFOLD VACUUM
15 16
6 BflSELINE. TQ=29.5
N=33 RSQ=.96
60=70.50 Bl°-2.43
+ TWO INJECTORS
N=ll RSQ=.77
80=63.44 B1=-1.99
•» 1 CENTRflL
INJECTOR
N=7 RSQ=1.0
80 = 71 .17 Bl=-2.57
Figure 3
-------
28
27
u 26
25
cr
a:
cc
cc
22
8
BTE VS MflP
» +
+
10 11 12 13
MANIFOLD VACUUM
14
15
• BflSELINE, TQ=29.5
N = 33
+ TWO INJECTORS
N=l 1
« CENTRflL INJECTOR
Ul
oo
Figure 4
-------
-59-
Prom another aspect, graphs of BTE vs. MAP can also be
valuable. The air-fuel ratio was varied from an equivalence
ratio (PHI) of about .6 to .8 for many test points. For the
power output and rpm to remain constant, the throttle opening
has to be varied somewhat inversely to the different PHI
levels - a lower PHI requires a larger throttle opening, for
example, which in turn lowers the throttling (pumping) loss
because of less restriction of the air flow. Therefore, it
can be determined from the results if throttling loss changes
have an appreciable effect on thermal efficiency when leaner
mixtures (leaner than .8) are used.
Results from baseline tests show that leaner mixtures (with
lower throttling losses) did not improve thermal efficiency;
in fact the reverse was true. Figures 5-8 show baseline (no
fumigation) test data for four categories of PHI. Figure 5
includes test points at the lowest levels of PHI and MAP,
which produced the lowest levels of BTE. Figure 6, which
includes intermediate low values of PHI and MAP, shows mixed
results ranging from nearly the lowest to the highest levels
of BTE. PHI values in this category appear to be the
transition from low to high BTE. The highest BTE values were
from PHI at .66-.67 and the higher MAP values, while the
lowest BTE values were from PHI at .63 and the lowest MAP
values. Figures 7 and 8 show consistently high levels of BTE
for PHI greater than .67 and MAP greater than 13 inches of
mercury. Thus it appears that throttling losses are not
dominant characteristics on this engine under these loads.
In summary, fumigation did not improve either volumetric or
thermal efficiency and we therefore recommend not pursuing it
further on this engine.
-------
BTE VS MflP BflSELINE .58
-------
BTE VS MflP BflSELINE .63
-------
BTE VS MflP BflSELINE .68
-------
-63-
Appendix VIII
Test Data
(Summary Sheets)
-------
TEST NO.: HO-81O809
ENGINE: 38O NAPS-ZM
87
TEST 0/T: B-31-82 O:, O
REPORT D/T: O9-27-B3 13:57
PO625B1
SUMMARY REPORT
TORO. MAX THR IGN. INJ. CORR.
RJ>»L FT-LB TO. POSN TIMG. T1MG. BHP
34.91
35. 12
34 .74
11.41
8 67
8.67
200O
2OOO
2OOO
15OO
15OO
1500
90
91
90
39.
30
30.
.5
.O
.O
5
0
O
100
1O1
99
10O
76
76
2 1O
110
oao
1270
13<4O
1O70
22
23
25
32
33
37
.OB
.SB
.SB
.SB
.SB
.OB
O.
O.
0.
O.
O.
O.
MEAS.
A>
7
8
8
a
8
9
If
.75
.21
.42
O3
.87
.80
PHI
EOUIV
RATIJJ
O.83
0 79
0.77
OBI
0 73
0.66
BSFC
GAS
BSFC EOUIV
FUEL LB/ LB/
LB/HR BHPHR BHPHR
ENERGY
EFFICIENCY
1000*
MBTU/
% BHPHR
206 7 26.66 O.764 0.359 38.6
217.9 26.54 O.756 0.355 39.O
22O.2 26.14 0.753 O.354 39.1
9O.8 11.31 O.991 0.466 29.7
87 6 9.88 1.14O O.536 25.8
6.60
6.53
6.51
8.57
9.85
BRAKE SPECIFIC EMISSIONS (G/BHPHR)
BSHC BSCO BSC02 BSNOX BSALDYH BSPART
1 .62O
1.878
1 .895
2.707
6.33O
1O3 3 1O.55 1.216 0.571 24.2 1O.52 14.888
1.388 408.16 9.555
2.537 398.25 7.SOS
2.S6O 398.O1 6.318
3.427 56O.25 11.878
5.272 597.08 4.804
6.761 473.22 1.673
O.O
O.O
O.O
0.0
O.O
O.O
-------
TEST NO.: HO-81O81O
ENGINE: 38O NAPS-ZM
87
TEST D/T: 9- 1-82 8:3O
REPORT D/T: O9-27-83 13:56
PO625B1
SUMMARY REPORT
TORO. MAX THR IGN.
RPM FT-LB TQ. POSN TIMG.
2OOO
2OOO
2OOO
150O
15OO
1500
91
90
89
29
29
28
.O
.0
.7
.5
5
0
IOO
99
99
IOO
IOO
95
47O
1 IO
080
146O
1310
1060
22
25
25
33
31
34
OB
. 58
OB
. 8B
OB
2B
O.
O.
0.
0.
O.
O.
PHI
INJ. CORR. MFAS. EOUIV
TIMG. BHP A/F RAMO
BSFC
GAS
BSFC EOUIV
FUEL LB/ LB/
/HR BHPHR BHPHR
ENERGV
EFFICIENCY
1OOO*
MBTU/
% BHPHR
BRAKE SPECIFIC EMISSIONS (G/BHPHR)
BSHC BSCO BSC02 BSNOX BSALOYH BSPART
35 43 7.71 O 84 2O5 8 26.7O O.754 0.354 39.1 6.52 1.961 2.677 441.49 9.687
26.12 O.745 0.35O 39.6
26.36 O.754 0.354 39.0
9.55 I.1O9 O.521 26.6
9.83 1.142 0.537 25.8
10.32 1 .267 O.595 23.2
O.
0.
0.
O.
O.
35
34
a
a
8
07
94
.GO
.60
. 15
8
8
8
8
10
26
.23
.38
92
O1
0
O
O
O
0
78
79
77
73
G5
215
217
80
87
IO3
. 7
.O
O
6
3
6
6
9
9.
IO.
44
,52
59
87
95
2
2
8
a
14
.384
.599
.597
.588
.545
2
2
4
5
7
.510
.515
.495
.394
. 6O1
433
436
618
614
623.
.92
.46
.02
77
48
8
7
8
3.
1 .
.678
.837
. 3O5
911
312
0.0
O.O
O.O
O.O
O.O
O.O
U1
I
-------
TEST NO.: HO-8IOS58
ENGINE: 380 NAPS-ZM
90
TEST 0/T: 9-21-82 9:45
REPORT O/T: O9-27-83 13:56
P062581
SUMMARY REPORT
BSFC ENERGV
GAS EFFICIENCY
% PHI BSFC EOUIV 1OOO*
TORQ. MAX THR IGN. INJ. CORR. MEAS. EQUIV AIR FUEL LB/ LB/ MBTU/
RPM FT-LB TQ. POSN JJ.MG_._ T IMG . _BHP_ A/F RAT_m _LB/HR LB/HR BHPHR BHPHR % BHPHR
BRAKE SPECIFIC EMISSIONS (G/BHPHR)
BSHC BSCO BSC02 BSNOX BSALOVH BSPART
20OO
15OO
9O.5
29.5
1OO
1OO
110 24.OB
I3OO 38.OB
0.
O.
35. 10
8 54
7 .88
9. 16
O.82
0.7 I
208 9
89 9
26.51
9.81
755 0.355 39.O
148 0.539 25.7
6.53
9.93
1 .608
5.231
2. 118
4 . 3O4
439.1O
605.85
8. 058
3.511
O.O
O.O
I
a\
-------
TEST NO.: HD-8IO859
ENGINE: 3BO NAPS-ZM
90
TEST D/T: 9-21-82 «O: O
REPORT D/T: O9-27-83 13:55
PO62581
SUMMARY REPORT
y. PHI
TORQ. MAX TIIR IGN 1NJ. CORR. MEAS. EOUIV AIR
RPM FT-LB TQ. POSN T1MG. TIMG. JHP_ A/F RAM0 LB/HR
2OOO 90.O IOO I4O 24.OB
2OOO 90.5 1O1
2OOO 89.5
150O 29.5
15OO 29.5
150O 29.5
99
IOO
too
IOO
9O 24 OB
80 27.OB
145O 36.OB
1'JO 4 2 . OB
91O 48.OB
O. 34 65 7 82 O 83 2O5 8
O. 34.83 7 96 O 81 2103
O.
O.
O.
O.
34 .45
851
8.51
851
8. 17
8.56
9.42
10. 3O
O.79 2112
O.76 80 4
O.69 89 9
O.63 1O9 6
BSFC ENERGY
GAS EFFICIENCY
BSFC EOUIV 1OOO*
FUEL LB/ LB/ MBTU/
LB/HR BHPHR BHPHR % BHPHR
BRAKE SPECIFIC EMISSIONS (G/BHPHR)
BSHC BSCO BSC02 BSNOX BSALDYH BSPART
6
6
6.
9
9
.57
56
49
54
.69
1.
1
1
4
5
.346
. 8O6
9O4
1-19
439
2
2
2
3
4
. 149
.262
.247
259
.057
335
437
433
6OO.
596
.40
. 19
69
13
. 13
a
6
6
4 .
1
.523
.777
.662
.909
.498
O
O
O.
O
O.
O
.0
.0
0
O
26.33 O.76O O.357 38.8
26.42 O.759 O.356 38.8
25.85 O.75O O.353 39.2
9.39 1.1O4 O.519 26.7
9.54 1.121 O.527 26.3
1O.64 1.250 O.587 23.6 1O.BO 17.268 8.327 594.76 O.91O
O.O
I
-------
TEST NO.: HD-8IO86O
ENGINE: 3BO NAPS-2M
90
TEST D/T: 9-22-82 8:45
REPORT D/T: O9-27-83 13:55
P06258t
SUMMARY REPORT
RPM
2OOO
2OOO
2OOO
15OO
15OO
1500
TORO.
FT-LB
90. 0
90.0
89.3
29.6
29.5
30.8
%
MAX
10^
1OO
100
99
1OO
1OO
1O4
THR
POSN
12
O9
O8
142
125
95
IGN.
T1MG.
22. OB
22.08
24 .OB
44. OB
47. OB
47. OB
1NJ.
TIMG.
O.
O.
0.
O.
O.
O.
CORR.
BHP
34 .88
34. 9O
34 5O
8 56
8.54
8 92
MEAS.
A/F
7 .79
7 .80
7 96
8.51
9. 34
1O. 19
PHI
EOUIV
RA1 10
O.83
0 83
O 81
0 76
O.69
O.64
AIR
LB/
208
210
211
83
92
IO9
HR
.O
.3
.2
1
.6
2
FUEL
LB/HR
26.70
26.95
26.53
9.77
9.91
1O. 72
BSFC
LB/
BHPHR
O.765
O.772
O.769
1 . 142
1 . 161
1 .202
BSFC
GAS
EOUIV
LB/
BHPHR
0.36O
O.363
0.361
O.536
O.545
0.565
ENERGY
EFFICIENCY
%
38.5
38. 1
38.3
25.8
25.4
24.5
1OOO«
MBTU/
BHPHR
6.62
6.68
6.65
9.87
1O.O3
1O.39
BSHC
1.351
1.411
1.542
16.7O4
7.O44
17.257
BRAKE SPECIFIC EMISSIONS
BSCO
1 .823
1 .932
1 .987
3. 195
4.57O
6.97O
BSC02
451 . 18
449.84
446.63
621 .66
613.92
578.51
BSNOX
9. 1O7
7.654
8.247
9.857
4.772
1 .592
(G/BHPHR)
BSALDYH BSPART
O.O
0.0
0.0
O.O
0.0
O.O
CO
I
-------
TEST
NO. : HO-81O861
SUMMARY REPORT
'/,
TORQ. MAX
RPM
2OOO
2OOO
2OOO
2OOO
15OO
I5OO
1501
15OO
FT-LB
9O.O
89.5
89.5
89.5
29.5
30.5
29.5
29.5
TO.
too
99
99
99
1OO
103
1OO
100
THR
POSN
I2O
too
O9O
O8O
143O
13-1O
1 170
9-40
ENGINE : 38O NAPS-ZM
IGN. 1NJ.
TIMG. T1MG.
21
23
22
22
38
42
55
56
.OB
OB
.88
BE
OB
.58
5B
.58
O.
O.
O.
O.
O.
O.
0.
O.
CORR.
BMP
34
34
34
34
8
a
8
8
.at
62
61
61
.56
84
56
56
MEAS.
A/F
7.61
771
7.73
7 57
8.O7
8.42
9.26
0.0
PHI
ECU IV
RATIO
O
o
o
o
0
0
o.
o
85
84
84
85
80
77
7O
O
9O
AIR
TEST
0/T:
BSFC
FUEL LB/
LB/HR
204
2OG
206
204
78
83
93
O
4
. 7
7
4
6
1
.5
O
LB/
26
26
26
26
9
9
IO
IO
'HR BHPHR
87 O.
. 80 O.
.75 O.
99 O.
74 1 .
.87 1 .
09 1 .
.64 1 .
772
774
773
7 BO
139
1 16
178
243
: 9-22-82 13:45
BSFC
GAS
EOUIV
LB/
ENERGV
EFFICIENCY
1OOO*
MBTU/
BHPHR %
0.
0.
0.
O.
O.
0.
O.
O.
363
364
363
366
535
524
553
584
38
38
38
37
25
26
25
23
.2
.O
. 1
.8
.9
.4
.O
.7
BHPHR
6
6
6
6
9
9
1O
IO
67
.69
.68
74
.84
65
. 18
.74
REPORT
BSHC
1
1
1
1
2
3
a
19
.462
.583
.623
.637
.882
. 58O
. 55O
.079
D/T:
O9-27-83
13:54
PO625B1
JAKE SPECIFIC EMISSIONS (G/BHPHR)
BSCO
1
2
2
2
3
3
5
7
.912
.014
. 1O2
. 1 14
. O91
.433
.355
.4 IO
BSC02
455
45O
451
455
624
604
609.
593.
.68
.69
.36
.56
.25
.45
22
28
BSNOX BSALDVH BSPART
9
8
8
8
9
7
4 .
2.
.348
.974
. 194
200
.696
.237
131
O66
O.O
O.O
O.O
0.0
O.O
0.0
00
O.O
-------
TEST NO.: HO-B1OB62
ENGINE: 38O NAPS-ZM
91
TEST D/T: 9-22-82 16:30
REPORT D/T: O9-27-83 13:48
PO62581
SUMMARY REPORT
X PHI
TORQ. MAX THR IGN. INJ. COPR. MEAS. EOUIV
RPM FT-LB TO. POSN TIMG. T1MG. BMP A/F RATIO
2OOO
2OOO
2OOO
15OO
150O
9O.O 1OO
91.O 1O1
89.5 99
1 1
O8
OS
29.O 1OO 145
29.O 1OO 134
17. 5B
16.OB
17 .SB
39. 5B
4O.8B
15OO 3O.O 1O3 1180 42.2B
BSFC
GAS
BSFC EOUIV
FUEL LB/ LB/
LB/HR BHPHR BHPHR
ENERGY
EFFICIENCY
1OOO*
MBTU/
% BHPHR
BRAKE SPECIFIC EMISSIONS (G/BHPHR) --•
BSHC BSCO BSC02 BSNOX BSALDYH BSPART
8 69 7.89 O 82
202 2 32.02 O.918 0.431 32.1 7.94 5.2O7
202 2 31.95 O.908 0.426 32.5 7.85 4.1O3
201 3 31.O4 0.896 0.421 32.9 7.74 4.796
78.6 1O.98 1.3O4 O.613 22.6 11.27 15.957
81 8 11.23 1.335 0.627 22.1 11.54 15.030
92.1 11.67 1.342 O.631 21.9 11.6O 8.423
37.4O1
32. 175
25.769
55.446
44.581
471 .76
474.61
475.41
625.82
632.67
4.9O4
5.049
6.O97
9.828
9.25O
5.276 6OO.16 4.O7O
0.0
O.O
O.O
O.O
O.O
O.O
I
-J
O
I
-------
TEST NO.: HO-S10863
ENGINE: 38O NAPS-ZM
91
TEST D/T: 9-23-82 9: O
REPORT D/T: O9-27-83 13:49
PO62581
SUMMARY REPORT
TORQ. MAX THR 1GN. INJ.
RPM FT-LB TO. POSN TIMG. 1IMG.
O.
O.
O.
O.
O.
O.
2OOO
2000
2000
1500
15OO
15OO
89
90
90
29
3O
29
.5
0
.2
.5
.2
.5
1OO
1O1
1O1
1OO
1O2
1OO
35O
210
14O
95O
146O
139O
2O
17
18
32
4O
4 1
.OB
OB
SB
.OB
58
8B
PHI
CORR. MEAS. EQUIV
BHP A/F RAI1O
34. 79
34 .83
34 .83
8 55
8.74
8 55
AIR
LB^HR
6.58
7 . 16
7.56
2O
56
8.23
O.98
0.9O
O 86
O.9O
O 8G
O 79
BSFC ENERGY
GAS EFFICIENCY
BSFC EOUIV 1OOO»
FUEL LB/ LB/ MBTU/
LB/HR BHPHR BHPHR % BHPHR
BRAKE SPECIFIC EMISSIONS (G/BHPHR)
BSHC BSCO BSC02 BSNOX BSALOYH BSPART
179 7 27.3O O.785 O.369 37.5 6.78 1.872
195 5 27.31 O.784 O.368 37.6 6.78 1.609
2O4 4 27.03 0.776 O.365 38.O 6.71 1.654
71.0 9.86 1.154 O.542 25.5 9.97 5.228
76.4 1O.11 1.156 O.543 25.5 9.99 4.242
81 8 9.93 1.162 O.546 25.4 1O.O4 4.O99
4.982 46O.89 1O.ISO
1.485 459.36 8.972
1.599 457.28 9.109
3.944 628.36 10.944
2.912 635.69 14.O91
3.458 636.92 14.477
O.O
O.O
O.O
O.O
0.0
O.O
-------
TEST NO. : HD-81O864
SUMMARY REPORT
TORO. MAX TMR
RPM FT-LB
2OOO 9 1 . O
2OOO 87. O
20OO 88. O
150O 29. O
15OO 29.5
1500 3O.O
TO. POSN
1OO 35O
96 27O
97 1OO
1OO 152O
1O2 138O
1O3 1 1OO
ENGINE
IGN.
T1MG.
2O. OB
21 5B
24 OB
42 .OB
42 48
46. OB
1NJ.
T1MG.
0.
O.
O.
O.
0.
O.
: 38O NAPS-ZM
CORR.
BMP
35 19
33.64
34 .Ol
841
8 56
8.71
Ml AS
A/F
6 57
7 45
7 . 76
7 .47
8.39
9.41
PHI
EOUIV
RATIO
0 98
O 87
O.83
O 87
0.77
O.69
92
AIR
LB/HR
182 0
19 1 O
202 2
74 1
83 1
99 3
TEST 0/T
FUEL
LB/HR
27.69
25.63
26.07
9.92
9.91
IO.56
BSFC
LB/
BHPHR
O.787
O.762
O.767
1 . 18O
1 . 158
1.212
: 9-23-82
BSFC
GAS
EOUIV
LB/
BHPHR
0.370
0.358
0.36O
0.554
0.544
0.570
13: O
ENERGV
EFFICIENCY
1OOO*
MBTU/
•/,
37 .4
38.7
38.4
25. O
25.4
24 .3
BHPHR
6.8O
6.59
6.63
1O.2O
1O.01
tO. 48
REPORT 0/T:
BSHC
O.692
O.995
1 .540
2.841
3.354
8.411
O9-27-83 13:5O PO6258I
RAKE SPECIFIC EMISSIONS (G/BHPHR)
BSCO
1 .474
1 .798
2.O69
2.971
3.453
5.268
BSC02
436.91
4 12 .89
413.55
579.60
565.09
599.31
BSNOX
12.993
7.277
4.735
12.O3O
4 . 162
4 . 064
BSALDYH BSPART
O.O
0.0
O.O
0.0
00
0.0
IVJ
I
-------
TEST NO.: HD-8IOB65
ENGINE: 38O NAPS-ZM
92
TEST D/T: 9-23-82 14: O
REPORT D/T: O9-27-83 13:5O
PO62S81
SUMMARY REPORT
y. PHI
TORQ. MAX TIIR IGN. INJ. CORR. MEAS COUIV AIR
RPM FT-LB TO. POSN TIMG. TIMG. _BHP_ A/F RA11J) J.B/HR
2OOO 90.5 1OO 39O 18.OB
2OOO BB.5 98 ?9O 21.OB
2OOO 90.O
150O 29.5
I5OO 28.8
15OO 31.O
99
1OO
98
105
12O 25.5B
149O 40 58
138O 41.58
1200 45.BB
O. 35.O3 6.74 O.96 179.7
O. 34 25 7.34 O.8B 189 6
O.
O.
O.
O.
34 84
8 56
8 36
9 OO
7 97
7 .44
8.37
9.3O
O.81 206.7
O 87 76 4
077 83 I
O.7O 94 4
FUEL
LB/HR
26.66
25.83
25.95
IO.26
9.93
IO. 15
BSFC
LB/
BHPHR
O.761
0.754
O.745
1 . 198
1 . 188
1 . 128
BSFC
GAS
EQUIV
LB/
BHPHR
O.358
O.354
O.35O
O.563
O.558
O.530
ENERGY
EFFICIENCY
1OOO*
MBTU/
_%_
38
39
39
24
24
26
. 7
. 1
.6
.6
.8
. 1
BHPHR
6.58
6.52
6.44
1O.36
1O.27
9.75
BSHC
O.8O9
0.977
1 .283
3.456
3.829
5. 100
RAKE SPECIFIC EMISSIONS (G/BHPHH)
BSCO
1
1
1
2
3
3
. 29O
.512
.764
.888
.483
965
BSC02
428.27
417.86
4O9.08
582.72
57O.88
502.53
BSNOX
1 1 .570
9.264
7.549
14 .407
4.O36
O.798
BSALOVH BSPART
O.O
0.0
O.O
0.0
0.0
O.O
I
-J
-------
TEST NO.: HO-81O866
ENGINE: 38O NAPS-ZM
92
TEST D/T: 9-23-82 15: O
REPORT D/T: O9-27-83 13:51
PO6258 1
SUMMARY REPORT
TORO. MAX THR IGN. INJ. CORR. MEAS.
RPM FT-LB TO. POSN TIMG. TIMG. BHP A/F
PHI
EOUIV AIR
RATIO LB/HR
BSFC
GAS
BSFC EOUIV
FUEL LB/ LB/
LB/HR BHPHR BHPHR
ENERGY
EFFICIENCY
IOOO*
MBTU/
% BHPHR
2OOO 89.O 98 3OO 21.58
20OO 9O.2 99 130 28.OB
1SOO 29.2 1OO 1490 41.OB
15OO 30.0 1O3 1410 49.8B
15OO 29.5 1O1 125O 49 2B
O.
O.
O.
0.
O.
34
34
8
8
a
41
.87
46
7O
56
7
8
7
a
9
.32
.08
52
. 1 1
.05
O
0
0
O.
O
88
BO
86
BO
7 1
188
2O7
76
81
9O.
7
6
4
a
a
BRAKE SPECIFIC EMISSIONS (G/BHPHR)
BSHC BSCO BSC02 BSNOX BSALDYH BSPART
2OOO 91.0 10O 42O 18.OB O. 35.19 6.66 O.97 178 8 26.86 O.763 0.359 38.6 6.6O O.851
25.77 0.749 0.352 39.3 6.47 O.9O1
25.70 0.737 0.346 4O.O 6.37 1.115
1O.16 1.2O1 0.564 24.5 1O.38 3.248
1O.O3 1.153 O.542 25.5 9.97 3.019
10.03 1.1720.551 25.1 1O.13 4.05O
1.253 43O.36 12.169
1.415 414.51 1O.553
1.678 406.91 9.544
2.817 568.25 12. 14O
2.78O 553.85 7.2O4
3.896 512.O6 1.124
0.0
O.O
O.O
O.O
O.O
O.O
-------
TEST NO.: HO-B1OBG7
ENGINE: 38O NAPS-ZM
91
TEST D/T: 9-24-82 16: O
REPORT D/T: O9-27-83 13:49
PO62SS1
SUMMARY REPORT
TORO. MAX THR IGN. 1NJ.
RPM FT-LB TQ. POSN I IMG. TIMG .
15OO 3O.O 1OO 150O 4O.OB O.
150O 29.5
15OO 29.5
15OO 29.O
15OO 29.5
15OO 29.8
98 I45O 44 .SB O.
98 1350 24 .58 O.
97 154O 21.58 O.
98 145O 23 OB O.
99 136O 53 58 O.
PHI
CORR. MEAS. EOUIV AIR
BMP A/F RAMO LB/HR
BSFC ENERGV
GAS EFFICIENCY
BSFC EQUIV 1OOO*
FUEL LB/ LB/ MBTU/
LB/HR BHPHR BHPHR % BHPMR
BRAKE SPECIFIC EMISSIONS (G/BHPHR)
BSHC BSCO BSC02 BSNOX BSALDYH BSPART
8. 74
8 6O
8 59
8. 44
8 59
8 .68
7. O2
7.68
8 6O
7.63
8. 23
8.57
O.92
O 84
0. 75
0.85
0 79
0.76
78 6
80 9
89 0
76 4
82 7
87 6
1 I .20
10.53
1O.34
1O.O1
10. 04
10.23
1 .281 O.6O2
1 .225 O.576
1 . 2O4 O.566
1 . 185 O.557
1 . 169 O.549
1 . 178 O.553
23.0 1 1 .07
24.0 1O.59
24 .5 1O. 4 1
24.9 1O. 25
25.2 1O. 1O
25. O 1O. 18
3.2O3
3. 1 14
4.228
2.269
2.320
2.966
13.291
5.575
4 . 066
1 1 .843
4.813
3.3O3
64O. 18
633. 10
633.52
615.45
61O.47
636.52
14 .966
12.415
6.504
4 .221
6.O49
1 1 .OO2
O.O
O.O
O.O
0.0
O.O
0.0
Ul
I
-------
TEST NO.: HD-81O87t
ENGINE: 3BO NAPS-ZM
9O
TEST D/T: 9-27-82 9: O
REPORT D/T: O9-27-83 13:53
PO62581
SUMMARY REPORT
RPM
2OOO
20OO
2OOO
15OO
15OO
15OO
2OOO
2017
2OOO
15OO
150O
1516
TORO.
FT-LB
89. 0
89.5
90. 0
29.5
3O.O
3O.O
89.5
B9.5
90.2
29. 0
3O.2
29. 0
y.
MAX
TO.
IOO
1O1
1O1
IOO
102
1O2
IOO
IOO
IOO
IOO
1O4
IOO
THR
POSN
12O
oao
oao
136O
1 13O
aao
120
O8O
oao
144O
129O
B7O
IGN.
INJ.
TIMG. TIMG
2O
22
22
44
50
46
22
21 .
21
28
44 .
52
5B
OB
5B
5B
OB
5B
OB
5B
OB
OB
SB
OB
0.
0.
O.
O.
O.
O.
O.
O.
O.
O.
O.
0.
CORR.
. BMP
34
34
35
8
8
a
34
35
35
8
8
8
80
87
09
62
77
77
89
17
14
47
83
56
MEAS.
A;
7
7
7
9
1O
10
7
a
7
a
9
10
if
.96
97
92
1O
O6
60
98
05
87
82
70
87
PHI
EOUIV
RATIO
O
0
0
0
O
O
O
0
o
o
o
o
.8 1
81
.82
.7 1
61
61
81
BO
82
73
67
6O
AIR
BSFC
FUEL LB/
LB/HR LBj
211
2 12
211
87
101
113
211.
215
211
81
9O.
1 16 .
2
1
2
6
t
2
2
7
2
a
a
a
26
26
26
9
1O
1O
26
26
26
9
9
1O.
fHR BHPHR
.51 0
.62 0
. 65 O
.63 1
,O5 1
69 1
48 0
81 O
, 84 O
27 1 .
36 1 .
75 1
.762
764
760
. 1 17
. 146
218
759
762
764
O94
O61
255
BSFC
GAS
EOUIV
LB/
ENERGY
EFFICIENCY
1OOO*
MBTU/
BHPHR %
O
O
O.
0
O
O.
0.
0.
O.
O.
0.
O.
358
359
357
525
538
572
357
358
359
514
498
59O
38.7
38 6
38.8
26.4
25.7
24 . 2
38.8
38.6
38.6
26.9
2/8
23.5
BHPHR
6
6
6
9.
9
1O.
6.
6.
6.
9.
9.
1O.
59
6O
57
65
91
53
56
59
60
45
17
85
Bi
1
1
1
O
1 t
19
2
t
2
2
2
2
„ p(
5HC
.241
.541
.559
.O
.336
.589
108
983
O63
O63
OG3
.063
JAKE SPECIFIC EMISSIONS (G/BHPHR)
BSCO
2
2
2
O.
5
8.
2.
2.
2.
2.
2.
2 .
O5B
. 148
192
.O
768
538
161
24O
232
232
232
232
BSC02
441
397
437
O
566
559
445.
442.
444.
444.
444.
444.
.85
.81
.53
. 15
.38
37
25
91
74
74
74
74
BSNOX BSALOYH BSPART
7.O61
6.836
7. 167
o.ota
3.759
1 .332
7 .759
6.4O4
6.4O8
6.408
6.4O8
6.408
0.0
O.O
0.0
O.O
O.O
O.O
O.O
0.0
0.0
O.O
0.0
O.O
-------
TEST
NO. : HD-8
SUMMARY REPORT
%
TORO. MAX
RPH
2OOO
2OOO
2OOO
15OO
15OO
15OO
2OOO
2OOO
2OOO
150O
15OO
15OO
FT-LB
89.5
9O.S
89. b
29. O
30. O
3O.O
90.0
90. 0
90.5
29.5
29. O
29.8
TO.
1OO
1O1
100
IOO
103
1O3
IOO
too
101
too
98
1O1
1O872
THR
POSN
MO
OHO
O80
137O
1 16O
87O
MO
O8O
O80
14 2O
125O
9-IO
ENGINE
IGN.
TIMG.
22 .58
22. 8B
23. 2B
44 .OB
44 OB
42 .28
22.58
22. 58
22.08
39. OB
45. OB
4O.5B
INJ.
TIMG.
O.
O.
0.
O.
O
O.
0.
O.
0.
O.
O.
O.
: 380 NAPS-ZM
CORR.
BMP
34
34
34
8
a
8
34
34
34
8
8
8
47
85
44
37
66
66
68
66
82
52
37
61
MEAS.
A/F
8.OO
8 12
817
9.O6
9.79
10.38
8 O3
8. OB
7.95
9. 37
10. 3O
10.97
PHI
EOUIV
RATIO
081
0 80
O 79
071
O G6
O 62
O 81
O.BO
O 81
0.69
O 63
0.59
9O
AIR
LB/HR
208 9
213-1
2125
87 6
98 9
114.6
2112
2134
2112
89 9
99 8
1 16 8
TEST
FUEL
LB/HR
26. 13
26.30
26. 02
9.67
10. 1O
1 1 .04
26.31
26.43
26.55
9.6O
9.69
1O. 65
D/T
BSFC
LB/
BHPHR
O
0.
O
1 .
1 .
1
0.
O.
O
1
1
1 .
758
755
756
156
166
275
759
763
763
127
157
237
: 9-28-82 14: O
BSFC
GAS
EOUIV
LB/
BHPHR
O. 356
O.355
O.355
O.543
O.548
O.599
O.356
O.358
0.358
O.529
O.544
O.581
ENERGY
EFFICIENCY
1OOO*
MBTU/
%
38.9
39.0
39.0
25.5
25. 3
23. 1
38.8
38.6
38.6
26. 1
25.4
23.8
BHPHR
6.55
6.52
6.53
9.99
1O. 08
1 1 .02
6.56
6.59
6.59
9.74
1O.OO
10.69
REPORT
BSHC
1
)
1
4
1O
22
1 .
1 .
1
5.
6
12.
776
621
679
, 4O5
816
8O8
796
769
.799
.229
.744
. 148
D/T:
O9-27-83 13:53 PO6258 1
RAKE SPECIFIC EMISSIONS (G/BHPHR)
BSCO
1
2
2
4
5
9
2
2
2
3
5
8
.976
.096
. 154
. 14 1
.993
.047
.212
. 138
. IOO
.358
656
.664
BSC02
453
446
445
616
572
591
448
443
444
626
6O7.
693.
.53
.93
.23
.63
.88
.60
.89
88
.55
62
3O
74
BSNOX
7.787
6.372
6. 136
12.112
4.412
2.O82
7.788
6.324
6.219
4 .457
1 .833
2. 181
BSALDVH BSPART
O.O
0.0
0.0
O.O
O.O
O.O
O.O
O.O
00
O.O
O.O
O.O
-------
TEST NO.: HO-81O873
ENGINE: 380 NAPS-ZM2
TEST D/T: 9-29-82 1O: O
REPORT D/T: O9-27-83 13:57
PO6258 1
SUMMARY REPORT
% PHI - BSFC
TORO. MAX THR IGN. INJ. CORR. MEAS. EOU1V AIR FUEL LB/
RPM FT-LB TQ. POSN TIMG. TIMG. _BHP_ A/F RATIO LB/HR LB/HR BHPHR BHPHR % BHPHR
BSFC ENERGY
GAS EFFICIENCY
EOUIV 1OOO*
LB/ MBTU/
2OOO 9I.O 10O 11O 2O.OB O. 35.O9 7 86 O 82
2OOO 89.8 99 O8O 21.5B 0. 34 63 8.06 0.80
2OOO 9O.2 99 07O 21 .8B O. 34.SO 7.95 OBI
15OO 29.O 1OO 144O 3O.OB 0. 8.31 8.66 O 75
15OO 28.7 99 12OO 48 OB 0. 8.23 9.97 O 65
15OO 29.5 1O2 930 51.OB O. 8 46 1O.25 O.63
211 2 26.87 0.766 O.36O 38.5 6.62
213 4 26.47 O.764 O.359 38.5 6.61
212 1 26.69 0.774 0.363 38.1 6.69
81 8 9.44 1 136 0.534 25.9 9.82
98 O 9.83 1.194 0.561 24.7 10.32
BRAKE SPECIFIC EMISSIONS (G/BHPHR)
BSCO
BSC02 BSNOX BSALDYH BSPART
2.O94
1 .841
1 .837
4.496
5.956
1 .883 • 449.19 7.O38
2.O42 443.94 6.324
2.O66 450.21 6.387
3.665 616.35 6.1OO
6.145 59O.2O 5.151
108.7 10.61 1.254 0.589 23.5 tO.B4 11.2OO 8.684 582.28 2.521
O.O
0.0
O.O
O.O
O.O
0.0
CD
I
-------
TEST
NO. : HD-8
SUMMARY REPORT
%
TORQ. MAX
RPM
2000
2OOO
2OOO
15OO
1500
1SOO
2OOO
2OOO
2000
15OO
15OO
15OO
FT-LB
89.5
89. O
90. 0
29.8
29.5
29.4
9O.O
9O.O
89.8
30.5
30.0
28.0
10^
too
99
1O1
10O
99
99
100
too
1OO
too
98
92
1O874
THR
POSN
120
oao
070
145O
134O
990
O90
OBO
oao
1430
137O
1O60
ENGINE
IGN.
TIMG.
21 .58
23. OB
22. OB
44. SB
54 .OB
55. OB
17 .OB
20.58
21 .88
52. OB
47 OB
50. OB
INJ.
TIMG.
O.
O.
0.
O.
O.
O.
O.
O.
O.
O.
O.
O.
: 38O NAPS-ZM2
CORR.
BMP
34
34
34
8
a
8
34
34
34
8
8
8
.39
18
.57
. 58
'IS
.45
.56
.57
49
.78
63
.05
MEAS .
A/F
7.89
8.12
7.93
8.65
9 G5
10.53
7.62
7.83
7.95
8.51
9 47
10. 46
PHI
EOUIV
RATIO
0 82
0 SO
0.82
O. 75
O.67
O 61
O.B5
O.83
O 81
O.76
0 68
O 62
AIR
IB/
2O8
213
21 1
81
89
107
21 1
21 1
213
82
87
103
0
TEST
D/T:
9-29-82 14:
BSFC
GAS
BSFC EOUIV
FUEL LB/ LB/
MR
9
4
.2
a
.9
.a
.2
2
.0
7
.6
3
IB;
26
26
26
9
9
IO
27.
26.
26.
9.
9
9.
'HR BHPHR BHPHR
.49 O
30 O
.62 O
46 1
31 1
24 1
70 O
96 O
79 O
71 1
25 1
88 1 .
.770 O.
.769 O.
.770 O
. 1O3 O
. O98 O
.212 O
.802 O
.780 O.
.777 O
107 0
.072 O.
227 O.
362
361
362
518
516
569
377
366
365
52O
, 5O4
576
0
ENERGV
EFFICIENCY
10OO*
MBTU/
y.
38.2
38.3
38.2
26.7
26.8
24.3
36.7
37.8
37.9
26.6
27.5
24. 0
BHPHR
6
6.
6.
9.
9
10
6
6
g
9
9
10
66
65
66
53
49
48
93
74
71
57
27
61
REPORT
BSHC
2
2
2
8
6
9
2
2
2
8
6
7.
.753
.321
.254
.931
.532
. 161
.827
. 30O
. 30O
.461
.071
955
D/T:
O9-27-83 13:57 PO6258
RAKE SPECIFIC EMISSIONS (G/BHPHR)
BSCO
2
2
2
3
3
a
2
2
2
2
3
7.
. 35O
.367
.351
.042
. 8O8
. 186
.328
.269
. 3O5
.835
234
941
BSC02
447
448
448
6O2
578
577
437
425
423.
609.
581 .
6O4.
.81
.51
.39
.59
.29
.20
35
07
25
72
48
71
BSNOX
6.O22
5.87O
5.5OO
1O.454
2.S8O
1 . 1O9
4.783
5.204
5.671
13.727
2.879
O.946
BSALOVH BSPART
0.0
O.O
O.O
O.O
O.O
O.O
0.0
0.0
O.O
O.O
O.O
0.0
to
-------
TEST NO.: HD-8IO877
ENGINE: 38O NAPS-ZM2
TEST D/T: 10-19-82 13: O
REPORT D/T: O9-27-B3 13:59
PO62S81
SUMMARY REPORT
% PHI
TORO. MAX THR IGN. INJ. CORP. MEAS. EOUIV AIR
RPM FT-LB TO. POSNI T1MG. TIMG. _BHP_ A/F RAJJ.O LB/HR
150O 29.5 tOO 142O 44.OB O. 8.51 8.69 O 74 82.7
150O 29.1 99 1280 48.5B O. 8.38 9.74 O66 92 1
15OO 3O.O 102 9OO 45.5B O. 8.64 10.42 O.62 111.4
FUEL
LB/HR
BSFC
LB/
BHPHR
BSFC
GAS
EOUIV
LB/
BHPHR
ENERGV
EFFICIENCY
10OO»
UQ Til/
MB I U/
% BHPHR
BSHC
BRAKE SPECIFIC EMISSIONS (G/BHPHR)
BSCO BSCO2 BSNOX BSALDYH BSPART
9.52 1.119 0.525 26.3 9.67 3.299
9.46 1.128 0.53O 26.1 9.75 5.934
10.70 1.237 0.581 23.8 1O.69 19.985
3.123 599.4O 11.074
4.TOO 586.13 3.888
8.4O5 564.77 1.58O
O.O
O.O
O.O
CD
O
I
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