1. Simulation Models for Spark Ignition
Engine: A Comparative Performance Study
Ashish Chaudhari, Upendra Garg, Vinayak Kulkarni, Niranjan Sahoo
Indian Institute of Technology Guwahati
Guwahati, Assam, INDIA-781039
IV th International Conference on Advances in Energy Research, IIT BOMBAY 2013

2. PRESENTATION OUTLINE
 Introduction
 Background
 Objectives
 Model Development
 Results and Discussion
 Validation of Simulink Model
 Validation of CFD Model
 Conclusions
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3. INTRODUCTION
INTRODUCTION
- The Internal combustion engines have been present since a century and their
performance, fuel economy has been greatly improved from the past and will
continue to increase.
- There has been a lot of research going on worldwide experimentally to bring new fuel
injection and combustion technologies for gasoline and diesel engines such as HCCI
(Homogeneous Charge Compression Ignition), GDI (Gasoline Direct injection) etc.
which are more efficient and produce lesser emissions [1].
- All the technologies introduced, cannot be tested experimentally as it is costly affair
and time consuming.
- The mathematical model based simulation using different environments like
Simulink, CFD etc makes the analysis easy, less time consuming and economical, for
the different technologies.
- The results of these simulations can be implemented for the new technology in the
engine.
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4. BACKGROUND
BACKGROUND
Sr No
Name of author
Work reported
Fuel used
Outcome
01
Ismail and
Mehta(2011)
Estimation of irreversibility
using quasi –three dimensional
combustion model
Hydrogen fuel More than 90% fuel
undergoes combustion and
20-30% availability found
destroyed.
02
Rakopoulos et
al.(2005)
Computationally development of
combustion model for observing
the effect of irreversibility
Hydrogen and
hydrocarbon
fuel
Availability destroyed
found to be14.99 to
15.02%..
03
Hongqing and
Huijie (2010)
Second law analysis with two
zone combustion model
Surrogate
fuels
Found availability
destroyed around 18.9%.
04
Al-Baghdadi
(2006)
the effect of different fuels alone
or their mixtures on the
performance parameters and
toxic emissions using a 2-zone
simulation model.
Gasoline,
ethanol,
hydrogen
Ethanol add with 30%
gasoline improves output
power, reduces NOx. it
increases heat release rate
and also sfc.
05
Wu et al.,
(2009)
developed the modified heat
transfer model for predicting the
heat transfer rate value
Gasoline
accurate and improved
over the other models
developed by researchers.
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5. OBJECTIVES
OBJECTIVES
- To develop a computer model for simulating a 4-stroke SI engine running on
any hydrocarbon fuel.
- Model should take into account realistic heat release during combustion and
heat loss from walls.
- Model should predict the cylinder pressure and temperature for one
thermodynamic cycle.
- It should predict performance related variables such as Indicated W.D. ,
IMEP, efficiency, power and torque for a given engine at any operating
condition.
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6. MODEL DEVELOPMENT
MODEL DEVELOPMENT
- A Single-zone, zero dimensional thermodynamic model is chosen for computer
simulation for a four stroke SI engine.
- The model is divided into parts (subsystems) with different processes
(compression, combustion, expansion, exhaust and intake) during 720º crank
angle (CA) of one-thermodynamic cycle in the following sequence:
- intake valve closing (IVC),
- spark timing,
- end of combustion (EOC),
- exhaust valve opening (EVO),
- exhaust process (from EVO to EVC/IVO) and
- intake process (from IVO to IVC)
- During the combustion process, the mixture can be assumed to be
unburned and in expansion process (end of combustion to EVO), the
cylinder consists of burned gases.
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7. MODEL DEVELOPMENT
MODEL DEVELOPMENT
- Spatial homogeneity of pressure for the whole cylinder, and spatial homogeneity of
temperature for each zone are considered.
- From IVC to EOC, the specific heats are calculated considering only fresh gases to
be present in the mixture and from EOC to EVC the properties are calculated
considering only burned gases.
- The gaseous mixture follows ideal gas law.
- It is assumed that exhaust valve closes at the instant the intake valve opens.
- Heat transfer is considered to happen only during the closed cycle operation i.e.
IVC to EVO.
- The Simulink model has following computation methodologies for determination
of various parameters;
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8. MODEL DEVELOPMENT
MODEL DEVELOPMENT
Volume:
- Using the geometry parameters of the engines, the cylinder volume at any crank
angle by following relation;
 π B2  
2
V ( θ ) = Vc + 
1 + a ( 1 − cos θ ) − l 2 − ( a sin θ ) 
÷
(1)


 4 
Heat release during combustion:
- The combustion investigation in engines has been carried out by analyzing cylinder
pressure data in single-zone thermodynamic model using Weibe function for
calculating heat release during combustion [6].
- A functional form often used to represent the mass fraction burned versus crank angle
curve [4].
  θ − θ 0  m +1 
(2)
X b = 1 − exp  −a 
÷ 
  ∆θ  
- If δ Q is the heat released from combustion, it is computed from the following
relation;
0, θ ivc < θ ≤ θ 0 and θ > θ eoc
dQch 
=
(3)
dx
dθ
Qmax b , for θ 0 < θ ≤ θ eoc


dθ
ch
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9. MODEL DEVELOPMENT
MODEL DEVELOPMENT
- Here,
dxb
dθ
is the fuel mass burn rate which can be calculated from Weibe function
(Eq. 2). So the equation of mass burn fraction takes the following form;
m
  θ − θ 0 m +1 
dxb a ( m + 1)  θ − θ 0 
a
dθ
=
∆θ

÷ exp  −a 
÷ 
 ∆θ 
 ∆θ  

(4)
-Where,
is combustion completeness factor and tells about the efficiency of combustion;
-It depends upon the intensity of charge motion and engine design and falls in the range of
0.35 ≤ a ≤ 0.8
- ∆θ is the combustion duration
- m is the form factor that affects the shape of mass burned profile.
- These parameters determine the shape of Weibe function and hence its accuracy to
predict
the actual heat release during combustion.
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10. MODEL DEVELOPMENT
MODEL DEVELOPMENT
Combustion duration:
- Total mass burned and the heat release are independent of the combustion duration (∆θ ).
- It is also very important factor because if same heat is released in small combustion
duration, the rate of heat release will be more and pressure peak would be higher as
compared to the large combustion duration. [7].
∆θ ( r , N , ϕ , θ ) = f ( r ) f ( N ) f ( ϕ ) f ( θ ) ∆θ
0
1
2
3
4
0
1
2
2
r
r
N
N
f1 ( r ) = 3.2989 − 3.3612  ÷+ 1.0800  ÷ ; f 2 ( N ) = 0.1222 + 0.9717  ÷+ 5.0510 × 10 −2  ÷
 r1 
 r1 
 N1 
 N1 
2
2
ϕ 
ϕ 
θ 
θ 
f3 ( ϕ ) = 4.3111 − 5.6393  ÷+ 2.3040  ÷ ; f 4 ( θ 0 ) = 1.0685 − 0.2902  0 ÷+ 0.2545  0 ÷
 ϕ1 
 ϕ1 
 θ 01 
 θ 01 
(5)
Cylinder temperature:
- The first law of thermodynamics applied to a closed volume of combustion chamber in
the following manner.
δ Qch = dU s + δ Qht + δW
dT
dθ
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=
⇒
δ Qch
dT dQht
dV
= mCv
+
+P
dθ
dθ
dθ
dθ
dV 
 dQch dQht
−
−P
mCv ( T )  dθ
dθ
dθ 


1
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11. MODEL DEVELOPMENT
MODEL DEVELOPMENT
Convective heat transfer rate:
- Using the Annand‘s correlation [2, 6] for convective heat transfer coefficient is obtained.
b
_
ρS B
_
hc B
p

÷ ; b = 0.7 and S p = 2aN ( mean piston speed )
=a
 µ ÷
k
60


µ air
3.3 × 10 −7 × T 0.7
9γ − 5
µ=
=
; k=
µ Cv
Here
1 + 0.027ϕ
1 + 0.027ϕ
4
(8)
(9)
Pressure rise:
- Ideal gas law for closed cycle operation is used to obtain the pressure rise in the
combustion chamber. The ideal gas equation can be differentiated with respect to crank
angle.
P=
m0 RT
V
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⇒
dP
dθ
=
1
dT
dV 
m0 R
−P
V 
dθ
dθ 


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12. MODEL DEVELOPMENT
MODEL DEVELOPMENT
Gas exchange process:
- In this process, the system is considered to be open system, from which gas can leave or
enter the system.. It is correlated with the valve opening area and pressure variation with
crank angle [8].
Aex = Aexo
dp
dθ
  180 ( θ − θ evo )
sin 
  θ evc − θ evo

÷

1
3
  180 ( θ − θ ivo )  
; Ain = Aino sin 
÷
  720 + θ ivc − θ ivo  
1
3
(11)
dp
 RT dm p dV 
 1 dm 1 dV 
−
;
= γ2  m
−
÷
÷
 m dθ V dθ exhaust dθ
 V dθ V dθ int ake
= γ1 p 
Here, Aino and Aexo are maximum opening in intake and exhaust valves. Both the equations
require mass flow rate which are to be found from the equations of fluid mechanics.
Indicated torque and mean effective pressure:
- Indicated torque T ( θ ) as a function of crank angle [9] is obtained from the piston ( FP )
force by neglecting inertia of various components and friction.
Piston Force
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FP = ( p − patm )

 n= l
sin 2θ
 B2 
Ap ; T ( θ ) = Fp a  sin θ +
A =π 
÷ ;
÷ (12)
2
2
a ; p
 4 
2 n − sin θ 

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13. MODEL DEVELOPMENT
MODEL DEVELOPMENT
Indicated work done:
w.d . ( J ) =
ivc + 720
∫
ivc
Indicated Power:
P ( watt ) =
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dV 

p(θ )
d

÷θ
dθ 

w.d .
120
×N
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14. PROCEDURE OF SIMULATION
PROCEDURE OF SIMULATION
-Based on the above formulation for the zero-dimensional and single zone thermodynamic model
simulation, the Simulink model is being introduced for a four-stroke SI engine.
-This model has been built in solvers for ordinary differential equation, when the model is formulated
in appropriate logical manner. The flow chart of Simulink model is shown in Fig. 1.
- First, this Simulink model is validated with the experimental conditions for the engine as given in the
reference [8].
-Side by side, a CFD model has been developed for the combustion chamber of a test engine to predict
its performance (Fig. 2-a).
-It is a four-strokes, air cooled, 1.8HP (HONDA make) engine with rated speed of 3600 rpm (Table 1).
-In this model, one should be able to perform transient simulation with dynamic/moving mesh(Fig.2-b).
-For a fixed compression ratio, it uses a module “in-cylinder” using the commercial package Fluent 6.3.
-The premixed combustion model is used for simulating engine combustion for different speeds. The
performance results obtained from both the computational models are then compared.
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15. Simulink
Simulink
Model
Model
Fig.1 Flowchart for Simulink modeling
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16. CFD MODEL
CFD MODEL
-
A 2-dimensional CFD model with
Axisymmetric swirl was made for the
new SI engine installed in Workshop.
-
The aim was to carry out the unsteady
simulation of the engine for closed cycle
period (IVC to EVO) and to obtain
pressure, temperature, and mass burn
fraction curve with crank angle.
-
The CFD model should be more
accurate because its solves conservation
equations for mass, momentum and
energy and also predicts combustion.
Fig.2 (a) 2-D geometry of
half combustion chamber
at clearance volume
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Fig.2 (b) Structured mesh of
cylinder at TDC (360 degree CA)
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17. CFD MODEL
CFD MODEL
•
Mesh Motion
Mesh is moved to crank angle corresponding to IVC i.e. 30 deg aBDC (570
deg.) as simulation would be setup from that crank angle.
Fig. 7 Clearance volume geometry
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18. CFD MODEL
CFD MODEL
Setting up the physics
•
Three user defined functions (UDFs) are used in this problem
1. Initialize.c : To initialize the flow field.
2. Laminar flame velocity.c: Function to calculate laminar flame speed at
different times.
3. work.c : to compute averaged values of pressure, temperature and mass
burn fraction.
•
Energy equation and turbulence are set on.
•
Premixed combustion model is used. Spark location and its timing is
specified.
•
Setting Boundary conditions and finally selecting solvers.
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19. ENGINE SPECIFICATIONS
ENGINE SPECIFICATIONS
Table 1 Specifications of engine
Sl. no
Specification
Simulink validation with
experimental results [8]
Simulink with
CFD validation
1
79.4
52
2
Bore ( B ) , mm
Stroke ( l ) , mm
111.2
46
3
Connecting rod (2a ), mm
233.4
81
4
Compression ratio (
7.4
4.8
5
Speed ( N ), rpm
4000 at 5 HP
3600 at 1.8HP
6
Spark timing ( º)
-25
-20
7
Intake manifold pressure ( Po ), atm
1
1
8
Relative Fuel-air ratio ( ϕ )
1
1
9
Fuel
C8H18 (Gasoline)
C8H18 (Gasoline)
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20. RESULTS AND DISCUSSION
RESULTS AND DISCUSSION
Validation of Simulink model
- Using the heat release rate obtained from experimental data [8], the
combustion duration for those operating conditions of experimental engine,
the reference values were found after tuning combustion model.
- At crank angles 25o before top dead centre (bTDC), the heat release rate
becomes zero and can be approximately taken as end of combustion so that the
combustion duration remains as 45o. The following reference parameters
obtained.
r1 = 9.2,
N1 = 4000 rpm,
ϕ1 = 1.1,
0
0
θ 01 = −20 and ∆θ1 = 45 .
- With this as input to the main Simulink model for simulating the cycle for any
operating condition.
- The parametric studies were done and results were compared with
experimental data
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21. RESULTS AND DISCUSSION
RESULTS AND DISCUSSION
(a)
(c )
Po=0.91 bar
(b)
Po=0.71 bar
(d)
Po=0.81 bar
Po=0.61 bar
Fig. 3 (a),(b),(c ) and (d) Variation of peak pressure with crank angle at different intake pressure
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22. RESULTS AND DISCUSSION
RESULTS AND DISCUSSION
Effect of intake pressure on performance with change in speed
Fig. 4 Thermal Efficiency and Power variation with speed
Fig. 5 Combustion heat release rate and pressure variation with speed
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23. RESULTS AND DISCUSSION
RESULTS AND DISCUSSION
Validation of CFD model
- HRR Analysis using Gasoline Experimental Data
Fig. 6 (a) and (b) pressure variation with crank angle using SIMULINK
Reference values for tuning combustion model are
r1 = 4.8,
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N 1 = 3600 rpm,
ϕ1 = 1,
0
0
θ01 = −20 and ∆θ1 = 140 .
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24. RESULTS AND DISCUSSION
RESULTS AND DISCUSSION
• Progress Variable (its value is 0 in unburned regions and 1 in burnt
regions).
Fig. 7 Clearance volume at end of suction stroke
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25. RESULTS AND DISCUSSION
RESULTS AND DISCUSSION
•
Progress variable at end of combustion for different rpm. We can find
combustion duration using contour plots as well.
Fig. 8 Combustion process at N=1000,2000.3000 rpm.
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26. RESULTS AND DISCUSSION
RESULTS AND DISCUSSION
-The CFD model is prepared using FLUENT was verified for its results at the condition
of rated power 1.8 HP at 3600 rpm.
- Simulink model which emulates the combustion could then be compared with results of
CFD model.
(b)
(a)
Fig. 9 (a) and (b) pressure and mass burn fraction variation with crank angle at N=1000 rpm.
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27. RESULTS AND DISCUSSION
RESULTS AND DISCUSSION
Fig. 10 (a) and (b) pressure and mass burn fraction variation with crank angle at N=2000 rpm.
Fig. 11 (a) and (b) pressure and mass burn fraction variation with crank angle at N=3000 rpm.
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28. RESULTS AND DISCUSSION
RESULTS AND DISCUSSION
- Brake power
Fig .12 Brake Power variation with speed
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29. CONCLUSIONS
CONCLUSIONS
- In this paper, an attempt has been made to address all the thermodynamic processes
occurring in engine through appropriate and suitable relations.
- A single-zone zero-dimensional Simulink model presents good interface to user.
- Side by side, the CFD analysis is also performed using commercial package (Fluent).
- In the present study, the Simulink model is validated from the experimental data and
found to be well agreement with the reported results.
- The Simulink analysis is performed to predict the maximum pressure after combustion.
- The deviation of mean pressure is about 6% higher as compared to experiment.
However, it reduces to 1% with reduction in intake charge pressure from 0.91 bar to
0.61 bar.
- During experiment, when the speed increases from 1000 to 4000 rpm, the heat release
rate rises from 3800 kJ to 4800 kJ, where as simulation shows constant heat release rate
of 2000 kJ for all speeds.
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30. CONCLUSIONS
CONCLUSIONS
- After successful validation of Simulink model, the numerical analysis was
carried out at rated rpm of engine (1.8 HP @ 3600 rpm) using CFD and
subsequently both the results are compared .
- Here, the pressure curve shows slight variation. At low speed 1000 rpm, the
maximum pressure recorded by CFD model showed relatively large deviation
(≥ 8%). But with increase in speed (at 3000 rpm), the variations reduces and
deviation remains below 3%.
- Mass burn fraction shows almost same trend of combustion duration. The
experimental brake power shows close match with Simulink and CFD results.
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31. REFERNCES
REFERNCES
[1] Ganesan, V. (2007) Internal Combustion Engines. Tata McGraw Hill, Third Edition.
[2] Wu, Y.Y., Chen B.C. and Hsieh, F.C. (2006) ‘Heat transfer model for small-scale air-cooled
spark-ignition four-stroke engines’. International Journal of Heat and Mass Transfer, Vol. 49,
pp.3895–3905
[3] Rakopoulos C.D. and Kyritsis D.C. (2005) ‘Hydrogen enrichment effects on the second law
analysis of natural and landfill gas combustion in engine cylinders’. International Journal of
Hydrogen Energy, Vol. 31, pp. 1384 – 1393
[4] Rakopoulos C.D., Michos C.N. and Giakoumis E.G. (2008) ‘Availability analysis of a syngas
fueled spark ignition engine using a multi-zone combustion model’. Energy, Vol. 33, pp.1378–
1398
[5] Ismail, S. and Mehta, P.S. (2011) ‘Second Law Analysis of hydrogen-air combustion in a spark
ignition engine’ International Journal of Hydrogen Energy, Vol. 36, pp.931 – 946
[6] Heywood, J.B. (1988) Internal Combustion Engine fundamentals. Tata McGraw Hill,
[7] Bayraktar, H. and Durgun, O. (2004) ‘Development of an empirical correlation for combustion
durations in spark ignition engines’. Energy Conversion and Management, Vol. 45, pp. 1419–
1431.
[8] Ganesan, V. (1996) Computer Simulation of Spark-Ignited Engine Process. Universities press
(India) Limited.
[9] Rattan, S. S. (2009) Theory of Machines. Third Edition, Tata McGraw Hill
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