An ASABE Meeting Presentation
Paper Number: 084045
UILU: 2008-7007
Combustion and Emissions Modeling of Biodiesel
Using GT-Power
Jonathon McCrady
Research Assistant, University of Illinois at Urbana Champaign 137 AESB, MC-644, 1304 W.
Pennsylvania, Urbana, IL 61801 USA, mccrady@uiuc.edu.
Alan Hansen
Associate Professor, University of Illinois at Urbana Champaign, 360S AESB, MC-644, 1304
W. Pennsylvania Avenue, Urbana, IL 61801, USA, achansen@uiuc.edu.
Chia-Fon Lee
Professor, University of Illinois at Urbana Champaign, 132 Mechanical Engineering Building,
1206 West Green Street, Urbana, IL 61801, USA, cflee@uiuc.edu.
Written for presentation at the
2008 ASABE Annual International Meeting
Sponsored by ASABE
Rhode Island Convention Center
Providence, Rhode Island
June 29 – July 2, 2008
Abstract. Previous investigations into the simulation of biodiesel combustion have shown some
significant differences between diesel fuel, soybean biodiesel, and rapeseed biodiesel. Upon review
of the methodology used, it was discovered that several improvements in the methodology needed to
be made.
Improvements were made in the injected fuel mass to better fit the operation of an unmodified diesel
engine and the estimation of the enthalpy of biodiesel liquid and vapor. These key improvements
The authors are solely responsible for the content of this technical presentation. The technical presentation does not necessarily reflect the
official position of the American Society of Agricultural and Biological Engineers (ASABE), and its printing and distribution does not
constitute an endorsement of views which may be expressed. Technical presentations are not subject to the formal peer review process by
ASABE editorial committees; therefore, they are not to be presented as refereed publications. Citation of this work should state that it is
from an ASABE meeting paper. EXAMPLE: Author's Last Name, Initials. 2008. Title of Presentation. ASABE Paper No. 08----. St. Joseph,
Mich.: ASABE. For information about securing permission to reprint or reproduce a technical presentation, please contact ASABE at
rutter@asabe.org or 269-429-0300 (2950 Niles Road, St. Joseph, MI 49085-9659 USA).
were then placed into the combustion model from the previous work to further study the fuel effects
on engine performance and emissions.
The load upon the engine had effects upon the performance of the engine. However, the emissions
of the soybean and rapeseed biodiesel combustion were higher than diesel regardless of load on the
engine.
Keywords. Biodiesel, combustion, physical properties, emissions
The authors are solely responsible for the content of this technical presentation. The technical presentation does not necessarily reflect the
official position of the American Society of Agricultural and Biological Engineers (ASABE), and its printing and distribution does not
constitute an endorsement of views which may be expressed. Technical presentations are not subject to the formal peer review process by
ASABE editorial committees; therefore, they are not to be presented as refereed publications. Citation of this work should state that it is
from an ASABE meeting paper. EXAMPLE: Author's Last Name, Initials. 2008. Title of Presentation. ASABE Paper No. 08----. St. Joseph,
Mich.: ASABE. For information about securing permission to reprint or reproduce a technical presentation, please contact ASABE at
rutter@asabe.org or 269-429-0300 (2950 Niles Road, St. Joseph, MI 49085-9659 USA).
Introduction
Previous research into the application of biodiesel fuel models using the GT-Power combustion
code (McCrady, et. al, 2007) has provided insight into the key processes that affect the fuels’
ability to atomize and combust. The study compared two biodiesel fuels, soybean and
rapeseed biodiesel, with the standard diesel fuel model provided with the GT-Power modeling
software. Both of these biodiesel fuels required a full fuel property definition to be constructed.
The results showed that both of the biodiesel fuels had a slightly shorter ignition delay, higher
peak cylinder pressure and peak heat release rate, and produced more NOx emissions than the
reference diesel fuel. The two biodiesel fuels were nearly identical in their results. This was
attributed to the two fuels being nearly identical in their fuel property definition.
This initial effort to use GT-Power to investigate biodiesel combustion proved to be useful in two
main areas. Knowledge was gained with regard to the construction of an engine simulation
model that could simulate various load schemes and fuels, and also in the construction of a fuel
library for varying sources of biodiesel fuels. However, further simulations and research
showed some limitations of the model utilized in the previous work. In the engine model, the
injected fuel mass required further investigation. The goal intended by McCrady, et al. (2007)
was to try to simulate an energy equivalent combustion comparison of both diesel and biodiesel
fuels. The only adjustment made to the model was to increase the mass of fuel injected. No
adjustments were made to any parameters of the fuel injection system. Some these parameters
include injection timing, injection duration, and injection pressure. The reason that this issue is
highlighted is that the simulated engine, a John Deere 4045HF, utilizes a common rail fuel
injection system. Studies completed by Postrioti et al. (2003) and Senatore et al. (2006) have
shown that significant changes in the injection strategies have to be performed to match an
engine’s performance when using biodiesel fuels. In order to obtain the same energy input
relative to diesel fuel, more biodiesel has to be injected into an engine due to the lower energy
content of the biodiesel fuel. To account for this increase in fuel mass, the start of injection and
the injection duration need to be altered to accommodate the total biodiesel fuel mass to be
injected.
Another limitation noted in the previous study was the lack of a model for the biodiesel enthalpy.
This included the enthalpy for both the liquid and vapor states. The previous work assumed the
enthalpy of the diesel fuel model to be adequate to represent biodiesel. In order to achieve a
more accurate simulation of an actual biodiesel fuel, an enthalpy model representing the methyl
esters of biodiesel needs to be used. It was not known if the enthalpy of biodiesel was similar in
any way to diesel fuel. It was also not known if large differences in this property (relative to
diesel fuel) would cause any major effects.
The objective of this paper was to improve upon some of the limitations of the previous work
and try to provide some additional insights into the combustion and emissions of soybean and
rapeseed biodiesel fuels. Details are be provided concerning the adjustments made to the fuel
injection characteristics and the methods used to estimate the enthalpy of the biodiesel liquid
and vapor. This is followed by details of the simulated engine and the engine operating
conditions. Lastly, the results of the combustion simulations are presented with details about
the combustion events and emissions of the biodiesel fuels and diesel fuel.
Modeling Approach
The modeling approach taken in this work is essentially the same as the previous work by
McCrady et al. (2007). As mentioned in the introduction, a couple of major changes were
2
provided to enhance the accuracy of the approach. These changes include adjustments in the
fuel injection characteristics and modeling the biodiesel enthalpy. The following sections
provide the details of the changes made in these two areas.
Adjustments in Fuel Injection Characteristics
As mentioned previously, the intention of the research is to model the combustion of biodiesel in
an energy equivalent case to diesel fuel. If biodiesel and diesel fuel had the same energy
content, the amount of fuel to inject into the cylinder would be the same. However, according to
Graboski and McCormick (1998), biodiesel typically contains 10-15% less energy than diesel
fuel. Due to this lower energy content, more biodiesel fuel than diesel fuel would need to be
injected into the cylinder. This would imply that the settings used in the fuel injection system of
the engine would not be adequate. In other words, the engine would not be able to inject
enough fuel into the engine to provide energy equivalency. Therefore, adjustments to the
parameters that control the fuel injection system in the engine need to be made. As detailed in
the works by Postrioti et al. (2003) and Senatore et al. (2006), these parameters include the
start of injection, injection pressure, and injection duration. However, after communications
about the engine model details with Chase (2007), the details in how to change these
parameters to gain energy equivalency could not be determined.
Without knowing how to adjust these parameters, it was decided to proceed with the research
assuming that the engine’s fuel injection system would be unmodified. This would assume that
the combustion study would no longer be energy equivalent. Furthermore, there would still
need to be further details into the amount of fuel injected into the cylinder. Since the engine
would be assumed unmodified, the conclusion was to try emulate what would happen in this
engine if biodiesel was introduced. The decision was made to simulate two different fuel mass
situations. The first being an equivalent fuel mass to the diesel fuel reference case. The
second case would include a small adjustment using the fuel’s density. This adjustment was
made based upon the assumption that the volume of fuel injected into the cylinder was
constant. Biodiesel and diesel fuel can have varying densities as shown by McCrady et al.
(2006). The difference in density of these two fuels can vary as much as 4-7%. To make the
adjustment in the injected fuel mass, the fuel mass of the diesel fuel case multiplied by the ratio
of the biodiesel density and the diesel fuel density. The densities of the soybean and rapeseed
(shown in Appendix 1) biodiesel fuels used in this study are both larger than the reference
diesel fuel leading to an increase in the injected fuel mass. The amount of fuel injected for each
loading condition is shown in Table 1.
Table 1. Adjusted fuel quantities used in simulations.
Diesel Fuel
Load
Low
High
mg
48.9
124.1
Soybean Biodiesel
Equal
Adj.
Mass
Mass
mg
mg
48.9
52.5
124.1
133.1
Rapeseed Biodiesel
Equal
Adj.
Mass
Mass
mg
mg
48.9
52.2
124.1
132.4
Biodiesel Enthalpy
In the previous study, the enthalpy of the reference diesel fuel was used for both biodiesel fuels.
The values for diesel fuel were used for both the liquid and vapor states for both biodiesel fuels.
In this work, more effort was applied to estimate the enthalpy of the liquid and vapor states of
both the soybean and rapeseed biodiesel fuels. The following two sections illustrate the
methods used to estimate the liquid and vapor biodiesel enthalpies.
3
Liquid Enthalpy
The enthalpy of the liquid fuel was estimated based upon the heat capacity of the fuel. Bondi’s
modification of Rowlinson’s method (Reid et al., 1987) was used to estimate the heat capacity of
the liquid. This is given in the following equation:
C pL − C po
R
[
= 1.45 + 0.45(1 − Tr ) + 0.25ω 17.11 + 25.2(1 − Tr ) Tr−1 + 1.742(1 − Tr )
−1
1/ 3
−1
]
where
CpL=specific heat of the liquid at constant pressure, J/mol-K
Cpo=specific heat at the reference state, J/mol-K
R=ideal gas constant, 8.3144 J/mol-K
Tr=reduced temperature
ω=Pitzer acentric factor given in Reid et al. (1987)
The temperature for the reference heat capacity was selected to be 0 K. To estimate the heat
capacity at this reference temperature, Reid et al. (1987) recommended the use of the Shaw
method. This method is a group contribution method based upon the structure of the molecule.
The group contributions are given in Table 2.
Table 2. Group contributions for Shaw’s method (Table 5-10 from Reid et al., 1987).
Group
-CH3
-CH2=CH-COO-
Value
36.8
30.4
21.3
60.7
The error of the Shaw method for the reference state heat capacity is less than 5% (Reid et al.
1987). For the modified Rowlinson equation, the error is also less 5% when tested across a
wide range of molecules (Reid et al. 1987).
The Pitzer acentric factor can be calculated from correlations provided by Reid et al. (1987).
This is given in the following equations:
ω=
α
β
α = − ln Pc − 5.97214 + 6.09648θ −1 + 1.28862 ln θ − 0.169347θ 6
β = 15.2518 − 15.6875θ −1 − 13.4721ln θ + 0.43577θ 6
θ=
Tb
Tc
where
Pc=critical pressure, bar
Tb=boiling temperature, K
Tc=critical temperature, K
4
The critical pressure and temperature were provided by Yuan (2005a). The boiling
temperatures for methyl esters were reported by Yuan (2005b). To find the heat capacity of the
biodiesel fuel, the heat capacity of each methyl ester was calculated individually and converted
into a mass basis. Then a mass weight averaging technique was used to compute the total
heat capacity using the mass composition of the biodiesel.
For the calculation of enthalpy, the heat capacity of the fuel can integrated with respect to
temperature. This method was illustrated by Heywood (1988) and is shown in the following
equation:
T
h − ho = ∫ C p dT
To
where
h=enthalpy of the fuel, J/kg
ho=enthalpy at the reference temperature To, J/kg
T=desired temperature, K
To=reference temperature, 0 K
The enthalpy values calculated from these methods needed to be fitted to the following equation
for use in the program.
h − href = a1 (T − Tref ) + a 2 (T − Tref
)
2
+ a 3 (T − Tref
)
3
where
h= enthalpy in J/kg
href= reference enthalpy (found from the reference enthalpy of the vapor minus the heat of
vaporization)
T=temperature in K
Tref=reference temperature, 298 K
a1, a2, and a3= fit constants.
The three constants, a1, a2, and a3 need to be determined for each fuel.
Vapor Enthalpy
Similar to the liquid enthalpy, the fuel vapor enthalpy was calculated by estimating the heat
capacity of each methyl ester in a vapor state. To estimate the heat capacity of the methyl
esters, Reid et al. (1987) recommended the use of the Lee-Kesler method shown in the
following equation:
C p − C po = (∆C p ) + ω (∆C p )
(0 )
(1)
where
Cp=heat capacity of vapor, J/mol-K
Cpo=reference state heat capacity, J/mol-K
5
(∆Cp)(0)=simple fluid contribution given in Table 5-8 of Reid et al. (1987)
(∆Cp)(1)=deviation function given in Table 5-9 of Reid et al. (1987)
ω=Pitzer acentric factor given in Section 3.3.6.1
The reference heat capacity is calculated using the Joback method reported by Reid et al.
(1987). This is a group contribution method dependent on the composition of the fuel. Joback’s
method is represented by the following equation:
⎛
⎞ ⎛
⎞
C po = ⎜⎜ ∑ n j ∆ a − 37.93 ⎟⎟ + ⎜⎜ ∑ n j ∆ b + 0.210 ⎟⎟T +
⎝ j
⎠ ⎝ j
⎠
⎛
⎞
⎛
⎞
⎜ ∑ n j ∆ c − 3.91e − 4 ⎟T 2 + ⎜ ∑ n j ∆ d + 2.06e − 7 ⎟T 3
⎜
⎟
⎜
⎟
⎝ j
⎠
⎝ j
⎠
where
nj=number of occurrences of reference groups of structures
∆a, ∆b, ∆c, ∆d=group contributions given in Table 6-3 of Reid et al. (1987)
The methodology for calculating the heat capacity of the fuel is the same as for the liquid
enthalpy. The heat capacity of the methyl esters are calculated individually, converted to a
mass basis, and then added to together using a mass weighted technique. To convert this into
an enthalpy, the heat capacity is integrated using the equation described in the previous
section.
To be used in GT-Power, the enthalpy values needed to be fitted to the following equation:
h − href = a1 (T − Tref ) + a 2 (T − Tref
)
2
+ a3 (T − Tref
)
3
+ a 4 (T − Tref
)
4
+ a5 (T − Tref
)
5
The variables in this equation are the same as for the liquid with the addition of the constants a4
and a5. The reference enthalpy is calculated from the reference temperature of 298 K. The
constants, a1 through a5 are calculated and then placed into GT-Power.
Combustion Model
The engine modeled for this study is a John Deere 4045HF475. Some of the details of the
engine are listed in Table 3. The dimensions of the actual engine and other operating
conditions were obtained from Chase (2007). These details of the engine were used in the GTPower software package. This software package is described in McCrady et al. (2007).
6
Table 3. Engine specifications.
Displacement
Stroke
Bore
Compression Ratio
Rated Speed
Rated Power
Peak Torque
Combustion System
Engine Type
Aspiration
Charge Air Cooling System
Fuel Injection System
4.5 L
127 mm
106 mm
17.2:1
2400 rpm
129 kW @ 2400 rpm
667 Nm @ 1400 rpm
Direct Injection
In-line, 4-stroke
Turbocharged
Air-to-Water
Common Rail
The details of the combustion model construction can be seen in McCrady et al. (2007). The
combustion model was only changed in the details of the mass of biodiesel fuel injected and the
enthalpy of the biodiesel liquid and vapor. The model was calibrated in the same manner as the
previous work.
Though this is a four cylinder engine, only a single cylinder was analyzed in order to save
computational time. The engine simulation was executed at two different loading conditions.
The first condition was at 257 N-m of torque at 2523 rpm, referred to as the low load case. The
second condition was at 667 N-m of torque at 1400 rpm, referred to as the high load case. A
complete summary of the test data is shown in Table 4.
Table 4. Test conditions for the engine and three test fuels.
Test Condition
Speed, rpm
Torque, N-m
Injection Timing, deg
Injection Duration, deg
Injection Pressure, bar
Fuel Quantity Injected, mg
Diesel Fuel
Low
High
Load
Load
2523
1400
257
667
-2.8
-1
15.4
28.4
1290
730
48.9
124.1
Soybean Biodiesel
Low
High
Load
Load
2523
1400
257
667
-2.8
-1
15.4
28.4
1290
730
Table 1
Table 1
Rapeseed Biodiesel
Low
High
Load
Load
2523
1400
257
667
-2.8
-1
15.4
28.4
1290
730
Table 1
Table 1
These simulations were completed to gain details into the combustion process and the resulting
emissions generated. Plots of the cylinder pressure, temperature and heat release were
studied along with the level of oxides of nitrogen (NOx) produced.
Results
Below are the results of the simulations of diesel fuel, soybean biodiesel and rapeseed
biodiesel. These results only contain data for the soybean biodiesel fuel. The rapeseed
biodiesel performed in a very similar manner. These results were omitted from discussion to
avoid any redundancy.
7
Low Load Analysis
Cylinder Pressure
In Figure 1, shown below, is the graph of cylinder pressure versus crank angle for the diesel fuel
and two soybean biodiesel combustion cases.
90
Cylinder Pressure, bar
80
70
Diesel Fuel
Mass Equivalent
Mass Adjusted
60
50
40
-20
-15
-10
-5
0
5
10
15
20
Crank Angle, deg
Figure 1. Cylinder pressure versus crank angle for the soybean biodiesel and diesel fuel low
load cases.
From Figure1, it can be seen that the initial pressure rise for all three cases is the same. This is
due to the injection starting just after head dead center. The initial pressure rise occurs from the
effects of the cylinder rising upwards, compressing the air entrained in the cylinder. The main
differences in the three cases come after head dead center. The pressure rise after head dead
center is not much larger than the motoring pressure in the engine, but there is another peak
present. Appendix 2 contains a table of the peak pressures obtained by all of the fuels. Though
difficult to see in the graph, the peak pressure of the mass equivalent and mass adjusted cases
of the soybean biodiesel perform nearly the same as the diesel fuel case. This is a very
interesting result due to the difference in energy content of both biodiesel fuels and diesel fuel.
Table 5, shown below, summarizes the energy content of low load cases.
Table 5. Energy differential of the diesel and biodiesel fuels in the high load condition.
Low Load
Mass
Equivalent
Mass Adjusted
Diesel Fuel
Fuel Mass,
Energy
mg
Difference
48.9
n/a
0%
n/a
Soybean
Fuel Mass,
Energy
mg
Difference
48.9
52.5
14.69
8.46
Rapeseed
Fuel Mass,
Energy
mg
Difference
48.9
52.2
14.6
8.88
Table 5 shows that all of the low load cases of biodiesel have a minimum of 8.5% less energy
than the diesel fuel. The difference in energy content would lead one to conclude that the peak
8
cylinder pressure for the biodiesel fuel should be less than that of the diesel fuel, but that is not
the case. Further investigations into the generation of the extra pressure needs to be
completed. It should also be noted that from Figure 1, the decay rate of the pressure was also
slower for the both biodiesel cases. Again, further exploration of this phenomena needs to be
completed.
Cylinder Temperature
Figure 2 shows the cylinder temperature versus crank angle profile of the soybean biodiesel
cases and for diesel fuel.
y
p
y
1300
Temperature, K
1200
1100
Diesel Fuel
Mass Equivalent
Mass Adjusted
1000
900
800
700
0
20
40
60
80
100
120
140
160
180
Crank Angle, deg
Figure 2. Cylinder temperature versus crank angle for the soybean biodiesel and diesel fuel low
load cases.
The soybean biodiesel combustion in the mass equivalent case caused the peak cylinder
temperature to be lower than the diesel and when the mass of the fuel was increased, the peak
temperature increased and became higher than diesel fuel case, even though there was less
energy injected into the cylinder. The peak cylinder temperatures are shown in Appendix 3.
The higher peak pressure of the mass adjusted cases could have been due to the combustion
starting sooner for the biodiesel fuel when the cylinder pressure was higher and cylinder volume
was smaller, causing a greater compression of gases in the cylinder at the start of combustion.
Heat Release
Figure 3 shows the heat release rate for the soybean biodiesel fuel cases and diesel fuel.
9
Heat Release (Normalized to 1.0)
0.035
0.03
0.025
0.02
Diesel Fuel
Mass Equivalent
Mass Adjusted
0.015
0.01
0.005
0
0
20
40
60
80
100
120
140
Crank Angle, deg
Figure 3. Heat release versus crank angle soybean biodiesel and diesel fuel low load cases.
The heat release rates in Figure 2 ndicate that the biodiesel fuel in both cases has a higher
initial peak rate of combustion than the diesel fuel. In addition, during the diffusion combustion
of the fuel, the biodiesel fuels again show a higher rate of combustion. This figure seems to
indicate that the biodiesel fuel has a tendency to want to burn at a faster rate than the diesel
fuel.
Nitric Oxide (NOx) Emissions
The NOx emissions generated by the combustion are shown in Figure 4 below.
10
60
50
NOx, ppm
40
Diesel Fuel
Mass Equivalent
Mass Adjusted
30
20
10
0
0
20
40
60
80
100
120
140
Crank Angle, deg
Figure 4. NOx concentration versus crank angle for soybean biodiesel and diesel fuel low load
cases.
The NOx formation rate of the biodiesel fuels in the mass equivalent condition show that overall,
less NOx is formed compared to the diesel fuel case due to the less energy being injected into
the cylinder. For the soybean biodiesel, the NOx concentration was 20% lower than the diesel
fuel case. The rapeseed biodiesel produced 22% less NOx in the mass equivalent case. When
the mass is adjusted, the NOx formation actually increases to be higher the diesel fuel case.
The soybean biodiesel produces 27% more NOx than diesel fuel. Additionally, the rapeseed
biodiesel produces 17% more NOx when the fuel mass is adjusted. This trend is in line with the
higher cylinder temperatures that are occurring within the cylinder. It again should be noted that
this is occurring with less energy being introduced into the cylinder. Further investigations into
the structure of the fuel itself needs to completed to look for any other unique properties of
biodiesel that could lead to answers as to why biodiesel burns at a faster rate, with higher
temperatures, and leads to higher NOx emissions.
High Load Analysis
Cylinder Pressure
Figure 5 shows the cylinder pressure versus crank profile of the mass equivalent and mass
adjusted soybean biodiesel cases and diesel fuel.
11
90
Cylinder Pressure, bar
80
70
Diesel Fuel
Mass Equivalent
Mass Adjusted
60
50
40
-20
-15
-10
-5
0
5
10
15
20
Crank Angle, deg
Figure 5. Cylinder pressure versus crank angle for the soybean biodiesel and diesel fuel high
load cases.
Figure 5 illustrates the pressure profile of the diesel fuel, the mass equivalent biodiesel, and the
adjusted fuel mass biodiesel combustion. The graph shows that the initial pressure rise of all
three load cases is exactly the same. This is due to the pressure rise from the motion of the
piston compressing the air. The peak pressure in all cases for all fuels in this load condition are
given in Appendix 2. The combustion of the fuel occurs after TDC. The pressure within the
cylinder decreases initially after TDC due to the late combustion of the fuel. The late
combustion negates the pressure rise due to the expanding volume.
As the combustion continues and the volume within the cylinder expands, the cylinder pressure
decreases. The rate of decrease is less dramatic for the biodiesel fuels than for the diesel fuel.
Cylinder Temperature
The cylinder temperature versus crank angle profiles for soybean biodiesel and diesel fuel are
given in Figure 6.
12
1900
1800
Temperature, K
1700
1600
Diesel Fuel
Mass Equivalent
Mass Adjusted
1500
1400
1300
1200
1100
1000
0
20
40
60
80
100
120
140
160
180
Crank Angle, deg
Figure 6. Cylinder temperature versus crank angle for the soybean biodiesel and diesel fuel high
load cases.
The peak cylinder temperatures for these cases are shown in Appendix 3. For both of the mass
cases, the peak cylinder temperatures do not exceed the peak cylinder temperature of the
diesel fuel case. This trend was to be expected due to two factors. First, the combustion event
occurred well into the expansion stroke. The cylinder volume is increasing during this stroke,
expanding the gases contained within it. This expansion will allow for decreases in both
temperature and pressure. Second, the energy content of the fuel is lower for the biodiesel
fuels. Table 6 shows the energy content of the simulations completed at this load condition.
Table 6. Energy differential of the diesel and biodiesel fuels in the high load condition.
High Load
Mass
Equivalent
Mass Adjusted
Diesel Fuel
Fuel Mass,
Energy
mg
Difference
Soybean
Fuel Mass,
Energy
mg
Difference
Rapeseed
Fuel Mass,
Energy
mg
Difference
124.1
0%
124.1
14.69
124.1
14.6
n/a
n/a
133.2
8.46
132.4
8.88
Table 6 shows that the diesel fuel has at least 8.5% more energy than any of biodiesel
simulations. This lower energy content would lead one to expect that the biodiesel fuel should
produce lower cylinder temperatures. Increasing the amount of energy injected into the cylinder
should increase the cylinder temperatures.
Heat Release
Figure 7 shows the heat release versus crank angle for the soybean biodiesel.
13
Heat Release (Normalized to 1.0)
2.50E-02
2.00E-02
1.50E-02
Diesel Fuel
Mass Equivalent
Mass Adjusted
1.00E-02
5.00E-03
0.00E+00
0
20
40
60
80
100
120
140
Crank Angle, deg
Figure 7. Heat release versus crank angle soybean biodiesel and diesel fuel high load cases.
Figure 7 shows that all three cases behaved in a very similar manner. The initial rise in heat
release is nearly identical. There is a small premixed peak that shows that the biodiesel cases
have a slightly higher premixed heat release peak. The biodiesel cases also have a higher heat
release peak during the diffusion phase of the combustion. The differences between the
biodiesel and diesel fuel are not very large, especially compared to low load condition.
Nitiric Oxide (NOx) Emissions
Figure 8 depicts the NOx concentration versus crank angle for the soybean biodiesel cases.
14
NOx of Diesel Fuel and Soybean Biodiesel Fuel
140
120
NOx, ppm
100
80
Diesel Fuel
Mass Equivalent
Mass Adjusted
60
40
20
0
0
20
40
60
80
100
120
140
Crank Angle, deg
Figure 8. NOx concentration versus crank angle for soybean biodiesel and diesel fuel high load
cases.
The NOx concentration of the soybean biodiesel in the mass equivalent case is 19% lower than
for the diesel fuel case. When the fuel mass is increased, the NOx emissions increased and
become 11% greater than the diesel fuel case. The rapeseed biodiesel behaved in a very
similar manner. In the mass equivalent case, the rapeseed biodiesel produced 21% less NOx
than the diesel fuel. In the mass adjusted case, the rapeseed biodiesel produced 6% more NOx
than the diesel fuel. The increase in NOx is more than likely due to the increase of energy
present in the cylinder coupled with higher cylinder temperatures. The overall fuel energy is still
smaller in this case and does not provide the full answer as to why this fuel produces greater
NOx emissions than diesel fuel.
Conclusions
The objective of this paper was to find and improve on some of the limitations found at the
conclusion of previous work on the simulation of biodiesel combustion. Though there were
several items to improve on, only two were selected for this work. Details were provided on how
the injected fuel mass was adjusted to provide more realistic results in an unmodified diesel
engine. Additionally, the methodology of estimating the enthalpy of the liquid and vapor
biodiesel was explained.
These two improvements in the modeling approach were then placed into the GT-Power
combustion code and used to simulate the combustion of soybean and rapeseed biodiesel and
diesel fuel. These simulations were completed at two different load conditions. One load was
to simulate a condition within the governor’s range and the other was used to simulate the peak
torque output of the engine.
15
The low load analysis showed that the biodiesel produced similar cylinder pressures, higher
peak cylinder temperatures and heat release rates, and greater concentrations of NOx. It was
quite intriguing that the biodiesel could produce higher temperatures, heat release rates and
NOx while have less energy being injected into the cylinder.
The high load analysis showed that the biodiesel also produces similar cylinder pressures.
However, the cylinder temperatures were lower and the heat release rates were very similar to
diesel fuel. The NOx concentration was also higher for the biodiesel fuels.
The only consistency between the two load conditions is that the biodiesel produces larger
concentrations of NOx. Further investigation into the chemical properties of the biodiesel
molecule may yield an explanantion that can find the reason why biodiesel produces higher
concentrations of NOx.
More research should also be applied into the estimation of the fuel properties of biodiesel in
order to enhance their accuracy. The methods used in this work were created to fit many
different types of molecules, not just esters. Being to adjust the methods used to estimate fuel
properties can only enhance the value of combustion modeling.
Acknowledgements
This work was supported in part by the Department of Energy Grant No DE-FC26-05NT42634
and by the Department of Energy GATE Centers of Excellence Grant No DE-FG26-05NT42622.
References
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16
Appendix or Nomenclature
Appendix 1: Fuel Densities
Fuel
Diesel Fuel
Soybean Biodiesel
Rapeseed Biodiesel
Density
kg/m3
830
890.7
885.7
Note: These densities are assumed at a temperature of 25oC.
Appendix 2: Peak Pressure During Combustion
Load
Low
High
Diesel
bar
86.09
89.84
Soybean
Equal
Mass
bar
85.76
89.85
Soybean
Adjusted
Mass
bar
86
89.83
Rapeseed
Equal
Mass
bar
85.72
89.85
Rapeseed
Adjusted
Mass
bar
85.88
89.83
Appendix 3: Peak Temperature During Combustion
Load
Low
High
Diesel
(K)
1244
1844
Soybean
Equal
Mass
(K)
1206
1742
Soybean
Adjusted
Mass
(K)
1271
1820
Rapeseed
Equal
Mass
(K)
1207
1744
Rapeseed
Adjusted
Mass
(K)
1266
1814
Appendix 4: Peak Nitric Oxide (NOx) Concentrations
Load
Low
High
Diesel
ppm
44.1
116.2
Soybean
Equal
Mass
ppm
35.3
94.4
Soybean
Adjusted
Mass
ppm
55.8
129.1
Rapeseed
Equal
Mass
ppm
34.4
92.2
Rapeseed
Adjusted
Mass
ppm
51.6
123
17