The results of this study allow us to draw the following conclusions:
1. The turbulent burning velocity coefficient of biogas-air mixture in the combustion chamber of a Honda Wave motorcycle engine fueled with compressed biogas containing 85% CH4 is approximately 1.3.
2. The indicated cycle work of a 110cc Honda Wave motorcycle engine reduces 28%
when switching from gasoline RON92 to compressed biogas containing 80% CH4.
3. For a given equivalence ratio, the indicated cycle work of the engine increases
almost linearly with CH4composition in the biogas fuel.
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Vietnam Journal of Mechanics, VAST, Vol. 37, No. 3 (2015), pp. 205 – 216
DOI:10.15625/0866-7136/37/3/5939
TURBULENT BURNING VELOCITY IN
COMBUSTION CHAMBER OF SI ENGINE FUELED
WITH COMPRESSED BIOGAS
Bui Van Ga, Nguyen Van Dong, Bui Van Hung
The University of Danang, Da Nang, Vietnam
E-mail: buivanga@ac.udn.vn
Received March 04, 2015
Abstract. Turbulent burning velocity is the most important parameter in analyzing pre-
mixed combustion simulation of spark ignition engines. It depends on the laminar burn-
ing velocity and turbulence intensity in the combustion chamber. The first term can be pre-
dicted if one knows fuel composition, physico chemical properties of the fluid. The second
term strongly depends on the geometry of the combustion chamber and fluid movement
during the combustion process. One cannot suggest a general expression for different
cases of engine. Thus, for accuracy modeling, one should determine turbulent burning
velocity in the combustion chamber of each case of engine individually. In this study, the
turbulent burning velocity is defined by a linear function of laminar burning velocity in
which the proportional constant is defined as the turbulent burning velocity coefficient.
This coefficient was obtained by analyzing the numerical simulation results and experi-
mental data and this is applied to a concrete case of a Honda Wave motorcycle engine
combustion chamber that fueled with compressed biogas. The results showed that the
turbulent burning velocity coefficient in this case is around 1.3 when the average engine
revolutions is in the range of 3000 rpm to 6000 rpm with biogas containing 80% Methane.
We can then predict the effects of different parameters on the performance of the engine
fueled with compressed biogas by simulation.
Keywords: Turbulent burning velocity, combustion simulation, biogas engine, spark igni-
tion engine, biogas fuel.
1. INTRODUCTION
Biogas is a renewable energy which is interchangeable with natural gas. However,
the problem which has existed until today is that all of the biogases yielded by different
biogas digestion tanks are of low pressure, low specific gravity and large specific volume.
The large quantity of CO2 present in biogas lowers its calorific value, flame velocity and
flammability range compared with natural gas. Therefore, the biogas needs to be purified
and compressed before it can be used in engines, especially in vehicle engines.
c© 2015 Vietnam Academy of Science and Technology
206 Bui Van Ga, Nguyen Van Dong, Bui Van Hung
Vietnam is tropical country and waste from agriculture production is abundant to
produce biogas. Besides, most of individual vehicles used in the country are motorcycles,
so the application of biogas on this kind of vehicle will be a good way for fossil fuel saving
and climate change mitigation. Research group GATEC of the University of Danang is
the pioneer in developing technology of gaseous fuel application on the motorcycle [1].
The research is initially carried out on LPG fuel [2] and now it is shifted to compressed
biogas [3, 4].
In order to convert a gasoline motorcycle engine to be a biogas engine, we should
carry out a theoretical study on engine performance before doing technical modification.
One of the most important unresolved problems of this study is the determination of
the turbulent burning velocity in the combustion chamber of the engine. There is no
consensus in literature whether the turbulent burning velocity is a characteristic quantity
that can be defined unambiguously for different geometries.
Turbulent premixed flame propagation was first investigated by Damko¨hler (1940).
He observed that the burning speed increases as Reynolds number increases and it was
affected by two different scales of turbulence: low intensity u′/SL, large scale turbulence
(weakly wrinkled flames); and high intensity u′/SL, small scale turbulence (strongly
wrinkled flames) [5, 6]. To develop mathematical models, Damko¨hler assumed that the
flamelet propagates with a constant velocity in a one-dimensional plane. Thus, the only
effect of turbulence is the wrinkling of the flame front which results in an increase in
flame surface area but internal structure and SL are unchanged. The turbulent burning
velocity was originally defined as follows [7, 8]
ST/SL = AT/AL, (1)
where SL is the laminar burning velocity, AT is the wrinkled flame surface area and AL
is the flow cross section area [5, 9, 10]. With such assumption, Damko¨hler developed the
first following model
ST
SL
= 1 +
u′
SL
, (2)
where u′ is the turbulence intensity.
The high pressure conditions have an effect on the thermo-physical properties of
the air-fuel mixture and the turbulent structures become finer as the pressure increases
[11], as well as a decrease of laminar burning velocity and the thickening of the laminar
flame front [12]. Kobayashi [11] has reported an increasing in turbulent burning velocity
as pressure increases and he suggested empirical correlation for methane/air flames
ST
SL
= 2.9
[(
P
Po
)(
u′
SL
)]0.38
, (3)
where P is the operating pressure and Po is the atmospheric pressure.
In these above expressions, laminar burning velocity SL depends on the physico
chemical characteristics of the mixture and it can be calculated if one knows the details
of rate of chemical reactions taking place in the combustion process. In general, this data
is established for mixture of pure fuel and air. M. Elia et al. [13] found a relationship
between laminar burning velocity and physical-chemical behavior of the mixture before
Turbulent burning velocity in combustion chamber of SI engine fueled with compressed biogas 207
combustion. R. Stone and A. Clarke [14] conducted experiments to determine the laminar
burning velocity CH4-air mixture diluted by CO2 atmosphere. Laminar burning velocity
of methane-air mixture can be established by empirical expressions as follows
SL = 0.366T1.42P−0.297 (m/s), (4)
where T = Tu/To is dimensionless temperature and P = Pu/Po is dimensionless pres-
sure.
According to Rallis and Garforth [15], laminar burning velocity of methane-air sto-
ichiometric mixture can be expressed by
SL = SL,oTαo , (5)
where αo is in interval of 1.37 and 2.33.
Biogas-air mixture could be considered as methane-air mixture diluted by CO2.
Laminar burning velocity in this case is presented in [16]. Thus, with a given cylin-
der pressure, mixture temperature, and composition of biogas, we can calculate laminar
burning velocity. Turbulent burning velocity depends not only on laminar burning veloc-
ity but also on fluid movement in the combustion chamber. In this work, we simplify the
relationship between turbulent burning velocity and laminar burning velocity by an as-
sumption of ST = f fSL. We try to determine the turbulent burning velocity coefficient f f
in the combustion chamber of a Honda Wave motorcycle engine fueled with compressed
biogas.
In view of the above, the specific aims of this study were defined as follows:
(1) To carry out experimental measurement of the indicated cylinder pressure
(2) To carry out simulation calculation with identical experimental condition
(3) To compare the simulation results with experimental data in order to determine
the turbulent burning velocity coefficient in the combustion chamber.
2. METHOD OF STUDY
2.1. Experimental setup
The experiment was conducted at the research laboratory of internal combustion
engines of Hanoi University of Science and Technology. The experimental facilities in-
clude a CD20” Chassis Dynamometer Test Bed AVL for motorcycles that is controlled by
Zoller software. Equipment Indiset 620 including a computer with the IndiWin620 soft-
ware for data acquisition from different sensors such as the pressure in the combustion
chamber, detonation, advance spark timing angle, TDC position, etc. The cylinder pres-
sure is obtained by means of a piezoelectric transducer. A synchronization optical crank
angle encoder 364C was used to acquire the cylinder pressure on the basic of crank angle
rather than the time. Operational parameters such as speed, acceleration and power of
the motorcycle are displayed on the control screen. Experimental data is treated by Con-
certo software. A schematic layout of the experimental facilities is shown in Fig. 1. Fig. 2
illustrates pictures of instruments for biogas motorcycle testing at the laboratory.
The following mass flow rates were determined: (1) air, by measurement of the
pressure drop across an orifice; (2) biogas, by means of rotameters; (3) petrol, by weighing
and timing. Hence, we can estimate the equivalence ratio of the mixture.
208 Bui Van Ga, Nguyen Van Dong, Bui Van Hung
5
parameters such as speed, acceleration and power of the motorcycle are displayed on the control
screen. Experimental data is treated
by Concerto software. A schematic
layout of the experimental facilities
is shown in Fig. 1. Fig. 2 illustrates
pictures of instruments for biogas
motorcycle testing at the laboratory.
The following mass flow
rates were determined: (1) air, by
measurement of the pressure drop
across an orifice; (2) biogas, by
means of rotameters; (3) petrol, by
weighing and timing. Hence, we
can estimate the equivalence ratio
of the mixture.
Fig. 2: Practical instruments for biogas motorcycle testing: Installing biogas motorcycle on
CD20" Chassis Dynamometer Test Bed AVL (a and b); Installing biogas cylinders (c); Pressure
regulators (d) and GATEC 25 biogas carburetor.
Compressed biogas from pressure cylinders is supplied to the engine with help of
conversion kit GATEC25 [18]. This is a special gas carburetor that ensures stable
equivalence ratio at any regime of the engine which was fitted to the engine's air intake
a. b.
c. d. e.
Fig. 1: Experimental facilities layout
1. AVL motorcycle test bed, 2. Force traducer; 3. Control
system, 4. Honda Wave motorcycle, 5. Cylinder pressure
sensor, 6. Optical decoder for engine speed, 7. Engine, 8.
GATEC 25 gas carburetor, 9. Air flowmeter, 10. Biogas
flowmeter, 11. Compressed biogas cylinder, 12. Ventilator
Fig. 1. Experimental facilities layout 1. AVL motorcycle test bed, 2. Force traducer; 3. Control
system, 4. Honda Wave motorcycle, 5. Cylinder pressure sensor, 6. Optical decoder for engine
speed, 7. Engine, 8. GATEC 25 gas carburetor, 9. Air flowmeter, 10. Biogas flowmeter, 11.
Compressed biogas cylinder, 12. Ventilator
a) b)
c) d) e)
Fig. 2. Practical instruments for biogas motorcycle testing: Installing biogas motorcycle on CD20”
Chassis Dynamometer Test Bed AVL (a and b); Installing biogas cylinders (c); Pressure regulators
(d) and GATEC 25 biogas carburetor
Turbulent burning velocity in combustion chamber of SI engine fueled with compressed biogas 209
Compressed biogas from pressure cylinders is supplied to the engine with help of
conversion kit GATEC25 [17]. This is a special gas carburetor that ensures stable equiv-
alence ratio at any regime of the engine which was fitted to the engine’s air intake up-
stream of the petrol carburetor. Biogas flow rate could be adjusted by a needle valve in
the gas inlet port. A pointer and scale arrangement indicating percentage of full throttle
opening was retrofitted to the butterfly valve. The fuel was inducted into the throat of
the venturi and the mixture flow rate was controlled by the throttle.
2.2. Numerical simulation
The combustion of biogas-air mixture in the combustion chamber of the engine is
simulated using Ansys FLUENT computational fluid dynamics software. The structural
and operational parameters of the engine are introduced into the program via the dy-
namic mesh option. Thermodynamic properties of the working fluid are set in PrePDF
table integrated into the FLUENT software. Turbulent combustion is simulated via the k-
ε turbulence model and the partially premixed combustion model with laminar burning
velocity SL which is determined empirically with fuel containing two major components,
which are CH4 and CO2. The turbulent burning velocity is determined via SL and a given
turbulent burning velocity coefficient by simplified relationship ST = f fSL.
6
upstrea of the pe rol carburetor. Biogas flow rate could be adjusted by a needle valve in
the gas inlet port. A pointer and scale arrangement indicating percentage of full throttle
opening was retrofitted to the butterfly valve. The fuel was inducted into the throat of the
venturi and the mixture flow rate was controlled by the throttle.
2.2 Numerical simulation
The combustion of biogas-air mixture in the
combustion chamber of the engine is simulated using
Ansys FLUENT computational fluid dynamics
software. The structural and operational parameters of
the engine are introduced into the program via the
dynamic mesh option. Thermodynamic properties of
the working fluid are set in PrePDF table integrated
into the FLUENT software. Turbulent combustion is
simulated via the k- turbulence model and the
partially premixed combustion model with laminar
burning velocity SL which is determined empirically
with fuel containing two major components, which are
CH4 and CO2. The turbulent burning velocity is
determined via SL and a given turbulent burning velocity coefficient by simplified relationship
ST=ff.SL.
Figure 3a illustrates the dimensions of the combustion chamber and cylinder of a typical
single-cylinder Honda Wave motorcycle engine. It is a four-stroke engine with bore, D = 50mm,
stroke, S = 49.50mm, and a rated power output of 5.1kW at 8000rpm. The compression ratio of
the engine is 9:1. The ignition system is powered by a CDI with an essentially fixed spark timing
Fig. 4: Variation of cylinder pressure,
fluid temperature and CH4
concentration during combustion
process of Honda Wave motorcycle
engine fueled with biogas containing
85% CH4
a. b.
Fig. 3: Dimension of cylinder and combustion chamber of
Honda Wave motorcycle engine (a) and meshing of
calculating space (b)
(a)
6
upstream of the petrol carburetor. Biogas flow rate could be adjusted by a needle valve in
the gas inlet port. A pointer and scale arrange ent indicating percentage of full throttle
opening was retrofitted to the butterfly valve. The fuel was inducted into the throat of the
venturi and the mixture flow rate was controlled by the throttle.
2.2 Numerical simulation
The combustion of biogas-air mixt re i the
combustion chamber of the engine is simulated using
Ansys FLUENT computational fluid dyna ics
software. The structural and operational para eters of
the engine are introduced into the program via the
dynamic mesh option. Thermodynamic properties of
the working fluid are set in PrePDF table integrated
into the FLUENT software. Turbulent combustion is
simulated via the k- turbulence model and the
partially premixed combustion model with laminar
burning velocity SL which is determined empirically
with fuel containing two major components, which are
CH4 and CO2. The turbulent burning velocity is
determined via SL and a given turbulent burning velocity coefficient by simplified relationship
ST=ff.SL.
Figure 3a illustrates the dimensions of the combustion chamber and cylinder of a typical
single-cylinder Honda Wave motorcycle engine. It is a four-stroke engine with bore, D = 50mm,
stroke, S = 49.50mm, and a rated power output of 5.1kW at 8000rpm. The compression ratio of
the engine is 9:1. The ignition system is powered by a CDI with an essentially fixed spark timing
Fig. 4: Variation of cylinder pressure,
fluid temperature and CH4
concentration during combustion
process of Honda Wave motorcycle
engine fueled with biogas containing
85% CH4
a. .
Fig. 3: Dimensi n of cylinder and combustion c r of
Honda Wave motorcycle engine (a) and meshin f
c lculating space (b)
(b)
Fig. 3. Dimension of cylinder and combustion chamber of Honda Wave
motorcycle engine (a) and meshing of calculating space (b)
Fig. 3a illustrates the dimensions of the combustion chamber and cylinder of a
typical single-cylinder Honda Wave motorcycle engine. It is a four-stroke engine with
bore, D = 50 mm, stroke, S = 49.50 mm, and a rated power output of 5.1 kW at 8000
rpm. The compression ratio of the engi e is 9:1. The ignition sys em is powered by a CDI
with an essentially fixed spark timing of 30◦ before TDC. The combustion chamber of the
engine is hemispherical in sh pe with a bowl-shaped cylinder he d and a flat piston top.
Fig. 3b shows the meshing of the computational space. To avoid the occurrence of
negative volume elements caused by element deformation during piston di placement,
the combustion chamber and the cylinder are meshed separately.
210 Bui Van Ga, Nguyen Van Dong, Bui Van Hung
3. RESULTS AND DISCUSSION
3.1. Simulation results
In the following section, the turbulent burning velocity coefficient is fixed at f f =
1.3 and then we predict the effects of the equivalence ratio and composition of biogas
to the performance of the Honda Wave motorcycle engine. The effect of advance spark
timing angle and engine speed has been published in previous works [18].
6
upstream of the petrol carburetor. Biogas flow rate could be adjusted by a needle valve i
the gas inlet port. A pointer and scale arrangement indicating percentage of full throttle
opening was retrofitted to the butterfly valve. The fuel was inducted into the throat of the
venturi and the mixture flow rate was controlled by the throttle.
2.2 Numerical simulation
The combustion of biogas-air mixture in the
combustion chamber of the engine is simulated using
Ansys FLUENT computational fluid dynamics
software. The structural and operational parameters of
the engine are introduced into the program via the
dynamic mesh option. Thermodynamic properties of
the working fluid are set in PrePDF table integrated
into the FLUENT software. Turbulent combustion is
simulated via the k- turbulence model and the
partially premixed combustion model with laminar
burning velocity SL which is determined empirically
with fuel containing two major components, which are
CH4 and CO2. The turbulent burning velocity is
determined via SL and a given turbulent burning velocity coefficient by simplified relationship
ST=ff.SL.
Figure 3a illustrates the dimensions of the combustion chamber and cylinder of a typical
single-cylinder Honda Wave motorcycle engine. It is a four-stroke engine with bore, D = 50mm,
stroke, S = 49.50mm, and a rated power output of 5.1kW at 8000rpm. The compression ratio of
the engine is 9:1. The ignition system is powered by a CDI with an essentially fixed spark timing
Fig. 4: Variation of cylinder pressure,
fluid temperature and CH4
concentration during combustion
process of Honda Wave motorcycle
engine fueled with biogas containing
85% CH4
a. b.
Fig. 3: Dimension of cylinder and combustion chamber of
Honda Wave motorcycle engine (a) and meshing of
calculating space (b)
Fig. 4. Variation of cylinder pressure, fluid
temperature and CH4 conce tration during
combustion process of Honda Wave motor-
cycle engine fuele with biogas containing
85% CH4
of 30° before TDC. The combustion chamber of the engine is hemispherical in shape with a
bowl-shaped cylinder head and a flat piston top.
Figure 3b shows the meshing of the computational space. To avoid the occurrence of
a) b)(a)
of 30° before TDC. The combustion chamber of the engine is hemispherical in shape with a
bowl-shaped cylinder head and a flat piston top.
Figure 3b shows the meshing of the computational space. To avoid the occurrence of
a) b)(b)
Fig. 5. Variation of CH4 mass fraction (a) and
O2 mass fraction (b) with crank angle during
combustion (n = 3000 rpm, ϕs = 30◦, biogas
containing 85% CH4, φ = 1)
Fig. 4 presents flame propagation at different crank angle and variation of temper-
ature and pressure as result in the combustion chamber of the engine fueled with biogas
Turbulent burning velocity in combustion chamber of SI engine fueled with compressed biogas 211
containing 85% CH4. We observe that flame front initially has a spherical shape and then
it is deformed during spreading out in space of the combustion chamber away from the
spark plug. A peak of pressure occurred at approximately 13 degrees after TDC and a
peak of combustion temperature occurred at about 5 degrees later.
Figs. 5a and Fig. 5b introduce the variation of concentrations of CH4 and O2 in the
combustion chamber with engine revolution speed of 3000 rpm, advance spark timing
angle 30◦ and equivalence ratio of 0.9, 1.0, and 1.5. The higher the slope of the curve,
the higher rate of fuel and oxidizer consumption. The results show that the highest fuel
consumption rate is achieved at an equivalence ratio of φ = 1. When φ = 0.9, in the early
phase of combustion, fuel consumption rate is not different with φ = 1 case but at the
end of combustion process, the difference is more evident on O2 consumption curve. The
slope of the curve for variation of CH4 and O2 with φ = 1.5 is significantly lower than
the previous two cases.
(a)
consumption rate is not different with = 1 case but at the end of combustion process, the
difference is more evident on O2 consumption curve. The slope of the curve for variation of CH4
and O2 with =1.5 is significantly lower than the previous two cases.
Figure 6a and 6b show the indicated cylinder pressure diagrams and the indicated cycle
(b)
Fig. 6. Effect of equivalence ratio to cylinder pressure diagrams (a) and cycle work diagrams
(b) (n = 3000 rpm, ϕs = 30◦, biogas containing 85% CH4)
Fig. 7. Variation of indicated cycle work with equivalence ratio
(n = 3000 rpm, ϕs = 30◦, biogas containing 85% CH4)
212 Bui Van Ga, Nguyen Van Dong, Bui Van Hung
Figs. 6a and 6b show the indicated cylinder pressure diagrams and the indicated
cycle work diagrams for various equivalence ratios with engine revolution speed of 3000
rpm, advance spark timing angle of 30◦ and biogas fuel containing 85% CH4. The indi-
cated cycle work of the engine is represented by the area bounded in the compression and
expansion curves. Variation of the indicated cycle work versus φ is presented in Fig. 7. We
observe that the indicated cycle work of the engine reaches its maximum value with φ in
range from 1 to 1.1 corresponding to the zone with the highest value of combustion rate.
(a) (b)
Fig. 8. Effect of CH4 composition in biogas to cylinder pressure diagram (a) and
to indicated cycle work diagrams (b) (n = 5000 rpm, ϕs = 35◦, φ = 1)
Fig. 9. Variation of indicated cycle work with CH4 composition in biogas
(n = 5000 rpm, ϕs = 35◦, φ = 1)
The following section presents the effects of biogas components on engine perfor-
mance. The calculations are carried out with advance spark timing angle 35◦, equivalence
ratio φ = 1 and engine speed n = 5000 rpm. Biogas fuel contains 60%, 70%, and 80% CH4.
Figs. 8a and 8b show the variation of indicated pressure and indicated cycle work versus
CH4 component in biogas as engine runs at 5000 rpm, advance spark timing angle of 35◦,
Turbulent burning velocity in combustion chamber of SI engine fueled with compressed biogas 213
equivalence ratio φ = 1. When the concentration of CH4 in biogas increases, the maxi-
mum pressure of the engine is increased leading to an increasing of the indicated cycle
work. Fig. 8a illustrates that at the fixed equivalence ratio, the combustion chamber peak
pressure decreases gradually with the introduction of carbon dioxide into the mixture,
due to the lower reactive charge inducted and the thermal release rate with the increase
of CO2 giving rise to the above observations. The result shows that the indicated cycle
work increases linearly with CH4 composition in biogas, as shown in Fig. 9.
3.2. Experimental measurements
The experiment was carried out firstly with gasoline RON92 and then mainly with
compressed biogas. The full load curves of the engine were established as the throttle
valve was fully opened. Figs. 10a and 10b present the comparison of the indicated cylin-
der pressure and the indicated work diagram of the Honda Wave motorcycle engine
fueled with gasoline RON92 and fueled with biogas containing 80% CH4 at the same op-
erating conditions: engine revolution speed of 3000 rpm, stoichiometric mixture, advance
spark timing angle 30◦. The results showed that when using biogas, the peak cylinder
pressure is 35bar which is lower than when gasoline was used (57bar). When switching
from gasoline to biogas, the maximum cylinder pressure drops and it can be explained
by two reasons: firstly, volume efficiency decreases because of gas fuel, and secondly, re-
duction in burning velocity and heat value of the mixture caused by the dilution of CH4
with CO2 in biogas. As a result, the indicated cycle work of the engine fueled with com-
pressed biogas containing 80% CH4 presents only 72% value of that fueled with gasoline
RON92. It confirms the observation of Jawurek et al. The authors observed that the en-
gine operates smoothly on gases containing up to 23% CO2, slightly noisily at 31% CO2
and harshly at 42% CO2. Maximum power output was 17% lower with CH4 than with
petrol. Increased CO2 content of the gas led to further losses, with a 45% loss (compared
with petrol) at 41% CO2 [19].
(a) (b)
Fig. 10. Comparison of cylinder pressure diagrams (a) and indicated cycle
work diagrams (b) of Honda Wave motorcyce engine fueled with
gasoline RON92 and with biogas containing 80% CH4
214 Bui Van Ga, Nguyen Van Dong, Bui Van Hung
The presence of carbon dioxide in the biogas reduces the burning velocity which
ultimately affects the performance of the engine. According to Bari [20], engine power
is lower when compared with that obtained in a diesel engine fumigated with natural
gas, while Neyloff found out that HC, NOx, and CO emissions from a CFR engine were
reduced [21]. Though the quantity of fuel admitted can be increased to ensure approx-
imately the same thermal loading [22], the indicated power output and cyclic variation
generally deteriorate with the increased proportion of carbon dioxide mixed with the
methane.
In the following section, we compare the indicated cylinder pressure given by the
simulation model and the experiment at different speed regimes in order to identify the
turbulent burning velocity coefficient. Compressed biogas contains 85% CH4, advance
spark timing angle of the engine is fixed at 27◦, equivalence ratio φ = 1 at full throttle
opening. Using each experimental result, we adjust the computational model’s pressure
diagram to match the experimental data.
(a) (b)
(c) (d)
Fig. 11. Comparison of cylinder pressures given by simulation and by experiment
(ϕs = 27◦, biogas containing 85% CH4, φ=1)
The comparison of indicated pressures given by experiment and by simulation
model at an engine revolution speed of 3000 rpm and 3620 rpm is shown in Figs. 11a
Turbulent burning velocity in combustion chamber of SI engine fueled with compressed biogas 215
and 11b with three turbulent burning velocity coefficient f f of 1.2, 1.3 and 1.5. The re-
sults showed that with the turbulent burning coefficient f f = 1.3, the indicated pressures
given by simulation are close to the experimental data. This coefficient is also consis-
tent with the case of n = 4070 rpm (Fig. 11c). As the engine revolution speed increases
to 5360 rpm, maximum indicated pressure decreases rapidly. If using the same coeffi-
cient f f = 1.3 as the above cases, the maximum indicated cylinder pressure given by the
computational model is higher than the experimental results by approximately 10%, as
shown in Fig. 11d.
4. CONCLUSIONS
The results of this study allow us to draw the following conclusions:
1. The turbulent burning velocity coefficient of biogas-air mixture in the combus-
tion chamber of a Honda Wave motorcycle engine fueled with compressed biogas con-
taining 85% CH4 is approximately 1.3.
2. The indicated cycle work of a 110cc Honda Wave motorcycle engine reduces 28%
when switching from gasoline RON92 to compressed biogas containing 80% CH4.
3. For a given equivalence ratio, the indicated cycle work of the engine increases
almost linearly with CH4composition in the biogas fuel.
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Các file đính kèm theo tài liệu này:
- turbulent_burning_velocity_in_combustion_chamber_of_si_engin.pdf