Simulation of High Efficiency Heavy Duty SI Engines Using
Direct Injection of Alcohol for Knock Avoidance
Paul N. Blumberg*, Leslie Bromberg*, Hyungsuk Kang**, Chun Tai**
*Ethanol Boosting Systems LLC, Cambridge, MA
**Volvo Powertrain North America, Mack Truck, Inc, Hagerstown, MD
March 10, 2008
Submitted for publication to SAE Powertrains, Fuel and Lubricants meeting, 2008
ABSTRACT
The use of direct injection (DI) of a second fuel, ethanol or methanol (or their
concentrated blends), is explored, via simulation, as a means of avoiding knock in
turbocharged, high compression ratio spark-ignited engines that could replace diesels in
certain vocational applications. The Ethanol Turbo Boost TM concept uses the second fuel
only under conditions of high torque to avoid knock, while using only conventional
gasoline throughout the rest of the engine operating range. This approach is an attractive
alternative for heavy duty vehicles that operate intermittently at high torque and within a
confined locale, reducing the logistical issues of supplying the knock-suppressing fuel.
The combination of GT-Power for engine calculations and a sophisticated chemical
kinetics code for predicting knock were used in the study. After benchmarking the
engine model against experimental data from an 11-liter heavy duty diesel engine, the 13mode emissions speed/load points were investigated operating in ethanol boosted SI
mode at a compression ratio of 14:1. For the baseline 11-liter case, two additional points
were added: one with a higher, knock-free torque at the B-rpm (limited to 190 bar
cylinder pressure); and one at 2100 rpm to achieve higher horsepower.
The ethanol boosting technology produces very high specific output, as the engine
operates stoichiometrically (with three way catalyst for emission control) and can also
operate at higher maximum engine speed. For a given engine size, it was determined that
ethanol-boosted SI engine could operate with about twice the torque and at higher
horsepower than the baseline diesel engine. Consequently, a scaled, downsized 7-liter
engine was simulated, which exhibited higher predicted efficiency than the baseline 11liter diesel engine over most of the B-rpm torque range of the 13-mode test.
The paper evaluates E85 and methanol as the knock-suppressing fuels. Relative to E85,
the use of methanol reduces the storage requirements of the second fuel by about a factor
of two. Further, the SI engine can operate with premium gasoline at the B50 point
without any knock-suppressing fuel (e.g., if it is exhausted) with the use of DI gasoline
and aggressive spark retard.
INTRODUCTION
The demand for high fuel efficiency is increasing due to environmental concerns, the
desire to reduce petroleum use and increasing fuel costs. While diesel engines have
traditionally been shown to be more efficient than conventional spark ignition engines,
they must overcome substantial hurdles in order to comply with emission regulations.
On the other hand, emissions from gasoline engines have been reduced to very low levels
through the use of the highly developed and effective three-way catalyst. The
phenomenon of knock in spark ignition engines is a major factor that prevents it from
reaching increased efficiencies. In general, the additional efficiencies that could be
achieved if knocking phenomena could be prevented derive from increased compression
ratio and engine downsizing with increased boost. Compared to heavy duty diesel
engines, in particular, operation at stoichiometry at constant torque/power requires less
turbocharging and overall engine mass flow. Furthermore, in achieving higher
efficiency, it would be advantageous to use renewable fuels [Haupt, Haput1, Nord,
Volvo, Turner, George].
A novel approach to avoid the knock in spark ignition engines has been developed that
exploits the use of direct injection of an alcohol in an otherwise conventionally fueled
gasoline engine employing port fuel injection. In this paper, the concept is described and
compared with a conventional heavy duty diesel engine as a baseline. The concept of
direct alcohol injection for knock control is first presented. The computational model is
then described. The results of simulation comparisons between a conventional diesel
engine and three alcohol-enabled, high compression, turbocharged, spark ignited engines
are described: the first engine uses the same engine size as the diesel baseline with
ethanol in the form of E85 as the direct injected fuel; the second, a downsized more
highly turbocharged engine with E85 as the DI knock suppressing fuel; and the third,
similar to the second but with methanol as DI fuel.
ALCOHOL BOOSTED CONCEPT
The ethanol boost concept was originally developed at the Massachusetts Institute of
Technology. The ethanol boost technology [Cohn1, Bromberg1, Bromberg2, Cohn2]
uses direct injection (DI) of a knock-suppressing second fuel when the engine is prone to
knock (usually at conditions of high torque). The cooling, which derives from the high
latent heat of vaporization of the knock-suppressing fuel, typically a lower alcohol such
as ethanol or methanol, is primarily responsible for this fact. Due to the charge cooling
from the DI process, the effective octane rating [Bromberg3] greatly exceeds the
chemical octane rating that these fuels would exhibit using conventional induction
methods such as port fuel injection (PFI). In the ethanol-boosted concept, DI of the
knock-suppressing fuel is used only in the amount required to prevent knock and gasoline
is supplied to the cylinder by conventional PFI. Since the engine operates at
stoichiometry (using a typical O2 feedback sensing system), a very high specific torque
output can be produced while emissions can be maintained at low levels through the wellproven and relatively simple three-way catalyst system without the use of EGR as a
major diluent. The technology opens the possibility of a spark-ignited gasoline engine
operating at high compression ratio (12 – 14) and high boost ratios of 2.0 – 2.5 times
ambient pressure, which is sufficient to produce a torque output equivalent to or greater
than more highly turbocharged heavy duty diesel engines operating lean with significant
EGR.
The ethanol-boosted technology offers an attractive alternative to diesel engines in light
duty applications and in certain heavy duty vocational uses involving central refueling
where the need for a second fuel would not pose a significant problem. It avoids the cost
of a complex emission control system and the very high pressure fuel injection system in
favor of a gasoline type DI injection system (~130 – 150 bar) and three-way catalysts.
Further, due to the high specific output at stoichiometry (compared to lean and EGRdiluted diesel operation at high torque needed to control NOx emissions), an engine of
smaller displacement can produce the required torque and power and operate at a higher
top speed. Under these conditions, the ethanol-boosted engine can be as efficient (as
measured by BTE, brake thermal efficiency) as the replaced diesel and have lower
specific CO2 emissions as well. However, due to the lower energy density of gasoline
and the second fuel on a volumetric basis, the volumetric fuel consumption will be
higher, as measured in liters (or gallons) per hour.
The concept has been demonstrated and proven in systematic dynamometer tests at Ford,
in collaboration with EBS [Stein]. In addition, Honda has independently investigated the
concept [Honda].
The ethanol-boosted concept is being developed by Ethanol Boosting Systems LLC, a
small spin-off company from MIT. The Volvo Powertrain (VPT) Division North
America (VPT) of Mack Truck, Inc. and EBS undertook a short analytical study to
examine the applicability of ethanol-boosted technology to the VPT MD11 diesel engine.
COMPUTATIONAL MODEL
A combined model to investigate the direct injection of alcohol in a spark ignition engine
was constructed. The model consists of two modules. The first module models the
engine flows, pressure and temperature evolution (including the effect of alcohol
evaporation and heat losses), residuals, self-consistent turbocharger, etc. The second
module performs the chemical kinetic evaluations using the pressure evolution and
composition of the cylinder charge determined from the first model.
ENGINE MODEL
The engine model is built upon the GT-Power platform supplied by Gamma
Technologies Inc. A single cylinder GT-Power, intercooled, turbocharged model,
equipped with DI and PFI dual fueling capability was used for the calculations. It
employs a two-zone combustion model so that temperatures in the unburned gas can be
tracked separately. The model also has a built-in finite element wall temperature solver
so that the effects of speed and load on heat transfer are included the gas temperature
calculation.
Conventional valves and timing of the baseline MD11 were used in the calculations (i.e.,
there was no use of VVT or VCT at lighter loads). No EGR has been used in the
calculations even at the light load points to reduce throttling losses. The compression
ratio of the engine is assumed to be CR = 14.
The dynamics of alcohol vaporization have substantial impact on the temperature of the
unburned “end gas” in which knock occurs. It is assumed that the alcohol is completely
vaporized by IVC. The effect of in-cylinder ethanol vaporization on unburned mixture
temperature was calculated from thermodynamic properties of the alcohol, with imposed
evaporation rates. The GT-Power model was modified to include a fuel with E85
properties, representing a concentrated form of ethanol, and contains its own methanol
fuel capability. The E85 was composed of 80% ethanol and 20% gasoline, which is in the
typical range for an 85% denatured ethanol and gasoline mixture.
Ethanol has a high latent heat of vaporization, about three times that of gasoline, as
shown in Table 1. The latent heat of vaporization is even larger for methanol. The ratio
of heat of vaporization to the heat of combustion for either alcohol to that of gasoline is
even higher, due to the lower heat of combustion of the alcohol relative to gasoline The
DI vaporization model allows specification of the half-life of the injected fuel in the
liquid state prior to vaporization. The half-life was chosen to insure complete
vaporization by IVC. In the present calculations, the port-fuel injected gasoline is 50%
vaporized at the time of injection. Since the heat of vaporization of gasoline is low, this
does not have a significant impact on the results.
Combustion timing and duration have been investigated. It is assumed that the 10%-to90% combustion duration (CA1090) of all the SI points were 25 CA degrees in duration.
This is achievable through the use of the right type of turbulence (i.e., tumble) and
possibly aided by dual ignition. As for combustion timing, MBT is assumed at all points
that do not require alcohol injection. The peak BTE at MBT was determined, via
analytical sweeps of the 50% burned point (CA50), to be 7.5 CA degrees ATDC. When
alcohol is required (because of knock), a 7.0 CA degree combustion retard was applied to
CA50 in order to minimize the alcohol requirement. A small amount of combustion
retard has a very pronounced effect on peak temperatures and pressures and a much
smaller effect on thermal efficiency.
For the turbomachinery, a compressor efficiency of 80% has been assumed, with a
turbine efficiency of 72%. Downstream from the compressor, an intercooler with an
efficiency of 95% has been assumed.
Premium gasoline (98 RON) has been used as the primary fuel, introduced through PFI.
Two DI knock-suppressing fuels have been considered: E85 (an available concentrated
form of ethanol), and M100 for methanol.
CHEMICAL KINETIC MODEL FOR KNOCKING
The model developed by Bromberg [Bromberg1] has been modified in order to determine
the knocking tendency of an engine. In order to provide useful predictions of knock
suppression throughout the engine operating regime, the end-gas conditions were
modeled using a volumetric compression of the unburned end-gas mixture. The method
used in reference [Bromberg] to determine the pressure/temperature evolution of the end
gas has been replaced with results from the engine model described in the previous
section.
The chemical kinetics code CHEMKIN 4.1 [Kee] was used for the chemical calculations.
The CHEMKIN code is a software tool for solving complex chemical kinetics problems.
This model uses the chemical model and reaction rate information based upon the
Primary Reference gasoline Fuel (PRF) mechanism from Curran et al. [Curran] to
represent the auto-ignition of the fuel. The mechanism also includes ethanol and
methanol chemical models [Marinov].
Table 1. Physical properties of gasoline, ethanol and methanol [SAEJ1297]
Ethanol
E100
Methanol
M100
RON
129
133
MON
102
105
(R+M)/2
115
119
Fuel type
Gasoline
Specific gravity
kg/l
0.75
0.794
0.796
Net heat of Combustion (LHV)
MJ/l
32
21
16
Net heat of Combustion (LHV)
MJ/kg
43
27
20
Latent heat of vaporization
BTU/gal
800
2600
3300
Latent heat of vaporization
MJ/kg
Vaporization energy/heat of combustion
Stoic air/fuel ratio
Equiv. Latent heat of vaporization
MJ/kg air
0.30
0.91
1.16
0.007
0.034
0.058
14.6
9
6.4
0.02
0.10
0.18
The gasoline is modeled as a mixture of primary reference fuels. RON 98 octane
(premium gasoline) was used (2% n-heptane, balance isooctane, by volume). E85
properties were constructed by using 80% ethanol and 20% gasoline by volume,
respectively.
The amount of alcohol (and the resulting initial temperature) is varied in the engine
module and the chemical kinetics module is then used to determine whether the specified
conditions result in knocking. The knock chemistry is stiff; small perturbations in the
initial temperature can result in large changes in behavior. It has been determined that
small initial temperature changes result in substantial changes in the timing of the autoignition. Auto-ignition onset varies by 5 crank angle degrees for one degree of change of
the initial temperature (corresponding to 1% change in ethanol energy fraction). Thus the
alcohol required to avoid knock is not very sensitive to the timing of auto-ignition.
Figure 1 shows typical results of the model, for knocking and non-knocking conditions.
The pressure curves, obtained from GT-Power are shown, as well as the cases with two
E85 fractions. The pressure curves are only slightly different for each case and they
overlap. The same is the case for the earlier times in the temperature traces. In the case
of the temperature knock curve, close to peak pressure, the temperature trace shows a
very fast increase, with very large derivatives. By comparison, a case with slightly more
E85 (“temperature-no-knock” case), there is still a substantial heat release, but the
derivatives are much smaller. The borderline knock case occurs when the temperature of
the unburned gas reaches 1100 K at the time of peak pressure.
Figure 1. Temperature and pressure traces for knocking and non-knocking conditions.
To determine the sensitivity of knock to the timing of the knock and the unburned fuel
fraction, a parametric study of the ethanol fraction required for different knock timing
assumptions was undertaken. The results shown in Figure 2, for intermediate speed and
highest torque, indicate that the required amount of DI ethanol is not very sensitive to the
assumed timing of knock, that also results in substantially varying amount of unburned
fuel fractions at time of knock. As the time to knock increases (at later crank angles) the
amount of ethanol required to avoid knock increases, while the mass of unburned fuel
decreases. In this paper, borderline knock conditions are assumed to occur at the time of
peak pressure.
The chemical kinetics module starts at a time where all the alcohol has evaporated, but
before any chemistry has had time to occur and continues through the time of peak
pressure. It was determined that very little chemistry (pre-combustion heat release)
occurs for temperatures lower than 650 K, by which time all alcohol has evaporated.
Figure 2. Ethanol fraction at borderline knock as a function of the assumed knock
timing.
ENGINE MAP
The European Stationary Cycle (ESC) (also known as OICA/ACEA cycle) has been used
to estimate the engine performance. The ESC is a 13 mode, steady-state procedure that
tests an engine at 4 different relative torques (25%, 50%, 75% and 100% of maximum
torque) at 3 speeds (A-, B-, and C- rpm), plus idle. Since knock is not expected at idle,
this point has not been included in the calculations. The ESC is a procedure used for
emission certification of heavy-duty diesel engines in Europe starting in the year 2000. It
is also used in the U.S. for development and in the form of the Supplemental Emission
Test to ensure compliance with “not to exceed” limits.
To investigate peak torque and horsepower capability, two additional points were
selected for analysis in the first case studied, i.e., the full size 11-liter engine. The first
one is maximum torque (limited by peak cylinder pressure) at B-rpm, and a second one at
a higher engine speed at slightly higher torque than B100.
ENGINE COMPUTATIONS RESULTS
The baseline engine chosen for the exercise is the VPT MD11, an 11.0-liter diesel engine
used in heavy duty vocational applications. Three cases were examined:
a) An EBS version of the MD11 engine “as is” (i.e., no modifications to the
valvetrain) at the same displacement, using premium grade gasoline (RON 98) as
the primary fuel and E85 as the second directly injected fuel. This will be
referred to as the EBS-11L engine.
b) A scaled, downsized version of the MD11 of 7.0-liter displacement referred to as
the EBS-7L engine with fuels as in case a).
c) A scaled downsized version of the MD11 with methanol (M100) employed as the
DI second fuel. Methanol was chosen due to Volvo’s potential interest in this fuel
and also because it is more effective in suppressing knock, requiring about 55%65% of the amount of E85, the overall lower volumetric fuel flowrate required
and the reduction of on-board storage capacity required.
d) Operational implications of exhausting the alcohol knock-suppressing fuel.
MD11 ENGINE “AS IS”
Prior to carrying out any calculations employing the EBS concept, the engine model was
used to compare the GT-Power prediction for the B100 diesel point using diesel fuel and
the experimental Rate of Heat Release (ROHR) data. The predicted efficiency was
approximately 3.5% less than the experimental data for otherwise equivalent conditions.
It was considered accurate enough to proceed with the assessment of the EBS engines in
comparison to the MD11 diesel.
The computational model has been used to investigate the performance of the EBS-11L
engine over the ESC points. Figure 3 shows the E85 fraction (by mass) required at
borderline knock. The BMEP, as well as all other results to follow, have been normalized
to the value of the predicted B100 point for confidentiality reasons. As expected, the
ethanol fraction slightly decreases with rpm and increases substantially with engine
torque.
Substantial amounts of ethanol are required even for 25% torque at the low engine
speeds. Thus for an engine operating for long-distance hauling, large amounts of the DI
E85 fuel would be required and issues with replenishing the E85 might be difficult
because of lack of fuel distribution infrastructure. However, for vocational applications
where the vehicles return to a central station at the end of the day or at several times
during the day, a refueling depot of the second fuel can easily be envisioned.
It can be seen in Figure 3 that at peak torque, especially for intermediate (B-rpm of the
ESC test matrix) and high speeds (C-rpm) substantially less than 100% E85 is required.
The question arises as to whether it would be possible to increase the maximum torque
expected from the engine if higher levels of E85 were to be used, while keeping the peak
cylinder pressure in the cylinder under 200 bar. The results are shown in Figure 4, for the
B-rpm of the EBS-11L engine. The high torque point is referred to as “BMAX”.
Figure 3. Required E85 fraction (by mass) at borderline knock for the EBS-11L engine
over the 13-mode operating range excluding idle.
Figure 4. Normalized BTE and E85 fraction (by mass) as a function of percent full load
at B-rpm for the conventional MD11 and the EBS-11L engines.
The green dot in Figure 4 is the predicted efficiency of the diesel engine using the same
model as for the EBS-11L engine, as discussed previously. The efficiency of the EBS11L engine is lower than that of the simulated engine operating as a diesel by about 1.5%.
Within the accuracy of the baseline these may be said to be virtually equivalent.
However, the engine can produce about twice as much torque as the MD11 engine, in
addition to producing more power because of higher allowable engine speeds. It should
be pointed out that in achieving the very high torque, no additional combustion retard
above 7.0 degrees was required to remain under the 200 bar peak cylinder pressure limit.
The implications of the EBS-11L conversion engine on the turbomachinery are discussed
next. The mass flow rates for the B-points in the ESC matrix for the conventional and
EBS-11L engine are shown in Figure 5 while the turbocharger compressor outlet
pressures are shown in Figure 6.
Figure 5. Normalized mass flow rate as a function of percent of full load for B-rpm of
the MD11 diesel and EBS-11L engines.
Figure 5 shows that there is substantial mass flow through the turbocharger in the MD11
diesel engine at B100 load, due to the fact that the engine must operate with an air-fuel
ratio lean of stoichiometric and also with heavy EGR. By comparison, the EBS-11L
engine operates stoichiometrically and without EGR. Thus, even at the BMAX point
(about twice the torque), the mass flow rates in the EBS conversion engine are
comparable to those in the base MD11 diesel engine at its maximum B-torque.
Figure 6 shows the manifold pressures for the same conditions as in Figures 4 and 5.
Because of the lack of dilution gases, the pressure of the case with almost double the
torque is similar to that of the baseline diesel engine at the diesel engine maximum
torque. The difference in compressor outlet pressure is about a factor of 2.5 for the
conditions of maximum torque of the baseline engine.
Figure 6. Normalized manifold pressure downstream from turbocharger compressor as a
percent of full load at B-rpm for baseline MD11 diesel engine and EBS-11L engines.
Due to rates of combustion, diesel engines have engine speed limitations that spark
ignited engines do not. Thus higher power can be achieved in EBS engines from both
increased torque and engine speed. The engine speed was raised from about 1830 rpm
(C-rpm in the ESC matrix) to 2100 rpm, while the peak torque was raised to 23 bar
BMEP, resulting in a power of increase of approximately 46% relative to the baseline
MD11. The ethanol fraction at this high power point was still a moderate 56% by mass.
Operation without much dilution does result in increased inlet turbine temperature. The
gas temperatures at the turbine are 930, 950 and 970 K for the A100, B100 and C100
points. The turbine temperature at the B-point with twice the torque is about 1000 K, and
it is about 1030K for the peak power point at the higher engine speeds.
In summary, the following points can be made from the comparison between the baseline
MD11 and the EBS-11L conversion:
•
•
•
•
Efficiency: the efficiency of the EBS-11L engine is competitive to the baseline
diesel at 75% and 100% load points, but falls short at light load conditions,
especially as no attempt to reduce pumping losses was investigated.
Peak Torque at B-rpm of EBS conversion engine is about twice that of the B100
point in the conventional diesel engine, strongly suggesting that 11 liters is too
large for the specific output required to match that of the baseline MD11 and the
possibility of significant engine downsizing.
Power at 2100 RPM is 46% greater than the baseline with a 10% larger torque
than B100.
Peak cylinder pressures are manageable and below ~120 bar except for the
BMAX point at ~200 bar.
•
•
Peak cylinder burned gas temperatures are less than 2565 deg K, including
BMAX.
E85 requirement (excluding the high power point) varies between zero and 64%
mass fraction at A100.
EBS-7L ENGINE
To investigate the consequences of aggressive engine downsizing, the MD11 has been
scaled down to 7.0 liters, keeping constant the compression ratio and bore-to-stroke ratio.
The valve diameters, lifts, port dimensions, inlet and exhaust dimensions have been
scaled with bore.
The combustion duration has been held constant (25 CA degrees for CA1090). However,
MBT has been redetermined at CA50 burn time at 9.5 deg ATDC. When E85 is used,
timing is retarded (CA50 = 16 deg ATDC) resulting in a ~ 1% loss in absolute efficiency,
but significantly reducing the peak pressures and the E85 requirements for reasons stated
earlier.
Turbocharger and intercooler efficiencies have been decreased slightly relative to the
EBS-11L engine due to higher boost ratios. Compressor, turbine and intercooler
efficiencies of 77.5%, 70% and 92.5% were employed. As in the previous case, neither
VVT, VCT nor EGR have been employed at the 25% load point (which is throttled).
Only the B-rpm points have been investigated for this case.
Figure 7. Normalized BTE and E85 fraction (by mass) as a function of percent of full
load for B-rpm for the baseline MD11 engine and the EBS-7L engines.
Figure 7 shows normalized efficiency for the baseline MD11 engine and the EBS-7L
engine using E85. The downsizing not only makes the efficiency of the gasoline engine
slightly higher than the baseline at the higher torque, but also makes it comparable to the
diesel baseline at the lower torque, which in this case is somewhat throttled but does not
employ VVT, VCT or EGR (which would further increase the efficiency at the light
loads). The E85 fraction is also shown, indicating substantial requirements at the high
loads (about 70% E85 by mass at the B100 point).
Although not shown in the figure, higher power operation has also been investigated. In
this case, the BMEP was 33 bar with an engine speed of 2100 rpm, producing a 35%
increase relative to the MD11 baseline. The ethanol fraction (by mass) at this point was
55%, so increased torque (at the expense of increased E85 fraction) could be possible at
this point.
The peak pressures in the cylinder of the EBS-7L engine at the B100 point is under 160
bar. The peak cylinder pressure of the baseline MD11 diesel engine is substantially
higher than this value, while the BMEP of the diesel engine is about 2/3 that of the EBS7L engine.
The gas temperature upstream from the turbine is about 970 K for the B100 case, and
about 1030 K for the high power case.
The flow rates through the B-speed points of the EBS-7L engine are about 30-40% lower
than those of the baseline MD11 diesel engine. Similarly, the manifold pressures are
about 30-40% lower.
In summary, the downsized EBS-7L engine competes very favorably with the MD11
diesel in efficiency, and it could be redesigned for even increased efficiency by taking
advantage of the reduced peak cylinder pressures and consequent reduction of ring and
bearing friction. However, the component temperatures need to be evaluated and may
require increased cooling and/or higher temperature materials for specific components, in
particular the turbocharger.
METHANOL USE IN AN EBS-7L ENGINE
In this section the EBS-7L engine operating with methanol as the antiknock DI fuel is
compared with both the baseline MD11 diesel engine and the E85 EBS-7L engine with
respect to BTE, fuel requirements, fuel flows and CO2 emissions. Methanol has two
advantages: the antiknock effect is higher (due to both higher intrinsic octane and higher
heat of vaporization), allowing of increased torque (if allowed by peak cylinder pressure),
and lower consumption for borderline knock than E85, decreasing the fueling
requirements.
Figure 8 shows the results for the B-rpm points of borderline knock with M100. The
efficiency of the EBS-7L engine operating with M100 is very similar to that of the same
engine operating with E85 as the antiknock fuel, as a comparison between Figures 7 and
8 confirms. However, the methanol requirements are substantially decreased so that a
smaller fraction of the gasoline (by volume) is needed to provide the same antiknock
capability.
Figure 8. Normalized BTE and methanol fuel fraction (by mass) as a function of percent
of full load for B-rpm for the EBS-7L engine.
It should be noted that although there is less consumption of the antiknock fuel this does
not mean that there is less volumetric consumption of total fuel. Figure 9 shows the
normalized fuel consumption of the MD11 diesel engine, compared with the total fuel
consumption of gasoline/M100, and gasoline/E85. Note that the total fuel consumption
(by volume) for the EBS-7L engine is very similar whether E85 or M100 is being used.
The reason is that although M100 is more effective at knock suppression, it also has
larger volumetric flow rates for a given energy content. Thus the total volumetric flow
rate is nearly independent on the nature of the alcohol.
Figure 9. Normalized volumetric fuel consumption as a percent of full load for the EBS7L engine, for E85 and M100 as the knock-suppressing DI fuels.
Also shown in Figure 9 are the gasoline fuel consumption, the E85 fuel and the M100
fuel consumption, separately for the cases of gasoline/E85, and gasoline/M100. Two
gasoline curves are shown corresponding for the case of E85 as the DI fuel and one for
the case of M100 as the DI fuel. It should be noted that the volumetric flow rate is
decreased in the case of gasoline/M100, with favorable consequences to the necessary
size of the secondary tank that contains the alcohol.
Table 2 shows the normalized brake specific CO2 emissions as a function of load for the
3 cases investigated. B100 for the MD11 baseline has been used as the normalizing
point. The use of alcohol fuels is, in general, favorable for reducing CO2 emissions.
However, for the EBS-11L engine, the inefficiency compared to the MD11 baseline
overrides this benefit. For the EBS-7L cases, there are small reductions in CO2
emissions, as expected, at the higher loads.
It should be mentioned that a significant part of the CO2 emissions in Table 2 for the EBS
engines, particularly at the higher loads where most of the fuel is consumed, derive from
renewable alcohol fuels. Although there is much debate about the net CO2 benefit from
use of fuels from various feed stocks and/or biomass, if the net benefit on CO2 from
production of these fuels is shown to be positive, the EBS engines would show a much
more favorable CO2 comparison on a “well-to-wheels” basis.
Table 2. Normalized CO2 emissions for the MD11 baseline and the three simulated EBS
engines.
MD11
B25
B50
B75
B100
1.15
1.04
1.01
1.0
EBS 11L
E85
1.29
1.10
1.02
0.99
EBS 7L
E85
1.15
1.03
0.97
0.94
EBS 7L
M100
1.15
1.03
0.98
0.95
The following points summarize the conclusions of this part of the study:
•
•
•
•
•
M100 reduces the storage requirement of the 2nd fuel by almost a factor of 2
relative to E85
Brake thermal efficiency of DI M100 in the EBS-7L engine is virtually equivalent
to that of DI E85 in the same engine.
Brake thermal efficiencies of 7-liter EBS engines with either DI E85 or DI M100
compare favorably with baseline MD11 diesel.
Brake specific CO2 is equal to or lower at all operating points with EBSconversion 7-liter engine compared to MD11 diesel, using either DI E85 or M100.
For equivalent mileage, overall fuel flow and storage is 25% and 33% greater for
the EBS-7L engine relative to the MD11 diesel engine at B75 and B100,
respectively.
ENGINE OPERATION IN THE ABSENCE OF 2ND FUEL
Without a DI knock-suppressing fuel, an EBS engine must be derated to avoid potential
damage from knocking. In this section, the operational issues when the second fuel is
exhausted are discussed.
Although a relatively small effect compared to alcohol fuels (e.g., ethanol, E85 or
methanol), DI of gasoline helps improve the knock resistance. It has been assumed that
when the alcohol fuel has been exhausted and there is available refueling, the second tank
is filled with premium, high octane gasoline or that high octane premium from the main
tank is used as the DI fuel and that all the required fuel is introduced via DI. In addition,
an increase of 10 degrees of combustion retard has been assumed. The model indicates,
under these conditions, that the maximum torque that the engine can produce without
knock is 50% of rated B-torque could be achieved without knock at the B-rpm (i.e., B50).
Use of VVT or VCT on the inlet valves to delay inlet closing and reducing effective CR
would increase knock tolerance further. Although the effective displacement of the
engine would be reduced, the combustion retard could also be reduced and the boost
increased. This option was not investigated quantitatively.
DISCUSSION
The previous sections described a preliminary comparison between a diesel engine and SI
engines that use direct injection of an alcohol for “on-demand octane” TM increase for
knock avoidance. The numerical calculations suggest that the SI engine may be an
alternative to a diesel powertrain for heavy duty applications, especially in circumstances
where the refueling infrastructure can be addressed. In this section, potential means of
further improve the concept will be described.
Because the end gas region is localized, it may be possible to decrease the amount of
alcohol required through stratification of the alcohol injection, predominantly injecting
the alcohol in those regions that will encompass the end gas. Substantial decreases in the
alcohol requirement could be achieved in this manner [Bromberg1, Bromberg3].
An additional attractive concept would be to operate in the EBS mode under conditions
of high torque, while operating in a more efficient mode at low torque. Relatively simple
solutions, such as operation with EGR can be used, but have not been quantified. In
addition, VVT or VCT can be used for minimization of the throttling losses.
CONCLUSIONS
The major conclusions of the study are as follows:
1. Due to its high specific output, an EBS conversion engine can be significantly
downsized relative to the baseline MD11 diesel; an engine in the range of 7 liters
was chosen given the output capability of the 11-liter baseline engine.
2. The brake thermal efficiency of an EBS-7L engine with premium gasoline as the
primary fuel and E85 as the secondary fuel is comparable to the baseline MD11.
Use of methanol as the DI fuel does not materially affect the BTE comparison.
3. Volumetric fuel flows are higher (~30%) with the EBS-7L engine compared to
the diesel MD11 due to the lower density of gasoline and lower energy densities
of the alcohol fuels relative to No. 2 diesel fuel.
4. Use of methanol rather than E85 allows approximately a factor of 2 reduction in
the amount of the second (DI) fuel required.
5. CO2 emissions on a specific basis are less at all points with the EBS-conversion 7liter engine as compared to the MD11 diesel.
6. Emissions treatment (current and future) in EBS-conversion engines can be
accomplished using the highly efficient and proven three-way catalyst system that
is enabled by operation at stoichiometric air-fuel ratios.
7. In the absence of a knock-suppressing second fuel, premium gasoline can be used
as the second fuel, allowing 50% of rated B-torque.
8. As a result of the deletion of the high pressure fuel injection equipment, the use of
a much simpler exhaust aftertreatment system and the reduction in the size of the
engine, a significant reduction in up-front cost of the engine should be realizable.
ACKNOWLEDGEMENTS
This work was partly sponsored by Volvo Powertrain North America Division of Mack
Truck, Inc.
Special thanks are given to Jan Wiman for providing and verifying data for the MD11
engine at the 13 ESC points.
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