Analytical study to minimize the engine exhaust emissions and safe knock limit of CNG powered four-stroke SI engine

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International Journal of Energy and Environment (IJEE), Volume 1, Issue 1, 2010, pp.31-52
ISSN 2076-2895 (Print), ISSN 2076-2909 (Online) ©2010 International Energy & Environment Foundation. All rights reserved.

35
ranges.
Knock integral method
is then proposed by Livengood and Wu [21]

by which an ignition delay
correlation could be used to determine knock occurrence crank angle (KOCA). The knock integral has
the following form;
∫
=
=
i
t
t
dt
0
1
τ
(9)

In terms of Crank angle
∫
=×
i
o
d
N
θ
θ
θ
τ
1
006.0
11
(10)

where
τ
is the induction time at the instantaneous temperature and pressure for the mixture, t (or
θ
crank angle) is the elapsed time from the start of the end-gas compression process (t=0 or at
0
θ
crank
angle: after ignition lag), t
i
(or
i
θ
crank angle) is the time of autoignition, and N is the revolutions per
minute. Using this concept, the most extensive tested correlation is produced by Douaud and Eyzat [22],
which is widely accepted by Heywood [1], Turner et al.

[24].
⎟
⎠
⎞
⎜
⎝
⎛
⎟
⎠
⎞
⎜
⎝
⎛
=
−
T
P
ON 3800
exp
100
68.17
7.1
402.3
τ
(11)

where
τ
is in milliseconds,
p
is absolute pressure in atmospheres,
T
is in kelvin and
ON
is the octane
number of fuel.

2.5 Total friction work for SI engine
The total friction work contains three major components. These components are the
pumping work
,
p
W
,
which is net work per cycle done by the piston on the in-cylinder gases during the inlet and exhaust
strokes;
rubbing

friction work
,
rf
W
, which is the work per cycle dissipated in overcoming the friction
due to relative motion of adjacent components within the engine; and
accessories work
,
a
W
, which is the
work per cycle required to drive the engine accessories, e.g., pumps, fan, generator etc. The total friction
work can be expressed as follows:
tf p rf a
WWWW=+ +
(12)

The data of total motored friction mean effective pressure (TFMEP) for several four stroke cycle, four
cylinder SI engines between 845 and 2000cm
3
displacement, at wide open throttle, as a function of
engine speed [33]

are well correlated by an equation, which is widely used by Heywood [1]

and Abd Alla
[34].
2
( ) 0.97 0.15 0.05
1000 1000
NN
TFMEP bar
⎛⎞ ⎛⎞
=+ +
⎜⎟ ⎜⎟
⎝⎠ ⎝⎠
(13)

where N is revolutions per minute.

Thus, the brake mean effective pressure (BMEP) of the standard engine can be found by the following
expression:
BMEP IMEP TFMEP=−
(14)

and;
Brake Power (BP)
d cyl
R
N
BMEP V n
n
⎛⎞
=×××
⎜⎟
⎝⎠
(15)
International Journal of Energy and Environment (IJEE), Volume 1, Issue 1, 2010, pp.31-52
ISSN 2076-2895 (Print), ISSN 2076-2909 (Online) ©2010 International Energy & Environment Foundation. All rights reserved.

36
where
IMEP
is the indicated mean effective pressure, which is the work delivered to the piston over the
compression and expansion stroke, per cycle per unit displacement volume;
R
n
is the number of
revolutions per cycle (=2 for four stroke cycle);
cyl
n
, is the number of cylinder (=4).

2.6 Species formation
Twelve species are considered in the combustion products in the cylinder and in the exhaust. they are :
H
2
O, H
2
, OH, H, N
2
, NO, N, CO
2
, CO, O
2
, and Ar. Properties of mixture at every time step are calculated
by using polynomial coefficients of internal energy for these 12 species. These species could reach
equilibrium condition if sufficient time is allowed for the reactions to take place under a certain state [1-
2]. The details are given in the book of Benson [2], consisting of 7 governing equations.
The formation of Nitric Oxide in an engine combustion chamber is a non equilibrium process. The rate
kinetics model based on the seven governing equations for NO formation is considered in the power
cycle and along the exhaust pipes based on the theory developed by Lavoie at.el., which is taken from
reference [2]. Carbon monoxide is computed under chemical equilibrium condition and then empirical
adjustment is made for kinetic behaviours based upon experimental results.

3. Validation
The results of the computational model are verified against the experimental data of the gasoline fueled
engine used by Baruah et al. [35], CNG fueled engine by Aslam [26]

and Ma Fanhua [27] as shown in
Figures 1, 2, and 3 respectively. These figures show that the results predicted by the mathematical model
are quite close (within 3%) to the experimental results. The technical data used for power cycle modeling
for different engines are given in Table 1.
Figure 1 shows the comparison of computed in-cylinder pressure vs. crank angle diagram for gasoline
fueled engine with experimental one during power cycle at fuel-air equivalence ratio of 0.967, 1.084, and
1.173. The computed data closely follow the experimental data except in the vicinity of peak pressure.
Figure 2 shows more clearly the comparison of measured and computed peak pressures at different
equivalence ratio. The average difference is around 3%. The discrepancy of this peak pressure is caused
by the cycle to cycle dispersion, and is due to non-homogeneity of fuel mixture supply in the cylinder in
actual case, but theoretical calculations are based on constant homogeneous mixture supply at every
cycle. Great deals of experimental research work by Zerves Efthimis [36], Ceviz MA [37] have been
done on cyclic dispersion. Many authors [38, 39] suggested that the major source of cycle to cycle
variation is due to the variation in the spherical burning area, produced by variation in the position of the
wall contact flame center and variation in the laminar flame speed at the spark center. In Figure 3, the
experimental and predicted results of NO and CO emission concentrations are compared with the
variation of fuel air equivalence ratio.
In Figure 4, the experimental indicated mean effective pressure and corresponding computational results
are compared with the variation of fuel-air equivalence ratio.
In Figure 5, the experimental results of cylinder pressure for CNG fueled engine are compared with the
present simulation model at the equivalence ratio(ER) of 0.7692. It shows that computed curve follows
the experimental pressure curves very closely.
Figure 6, compares the experimental Brake means effective pressures (BMEP) given by Aslam et al. [26]
for a CNG fueled engine and the theoretical ones calculated by using the present model. These figures
show that computational results are reasonably in good agreement with the measured one. The BMEP
maximizes at midrange of engine speed (≈ 3000 RPM). This is due to the friction which increases with
increasing engine speed. Another reason of BMEP loss is due to the longer ignition delay and low flame
speed of CNG fuel. These factors affect the Brake specific fuel consumption (BSFC). The figure shows
that BSFC drops as the engine speed increases in the low speed range and level off at medium speed
range and increases towards the high speed range. This is because of the heat losses to the combustion
chamber which proportionally increases with decrease in speed. On the other hand BSFC increases due
to the friction which increases with increase in engine speed.
It can be seen obviously from these comparisons that the present simulation model is capable to compute
accurately the engine performance of SI engine using gasoline or CNG fuel.



International Journal of Energy and Environment (IJEE), Volume 1, Issue 1, 2010, pp.31-52
ISSN 2076-2895 (Print), ISSN 2076-2909 (Online) ©2010 International Energy & Environment Foundation. All rights reserved.

37
Table 1. Technical data of engines for validation

Gasoline
Engine type (SI) Vauxhall Victor 2000cc (Baruah [34], Benson [2])
Cycle 4 stroke
Cylinder bore 9.53cm
Stroke 6.92cm
Connecting rod length 13.65cm
Compression ratio 8.5
Angle of ignition 33.6
0
bTDC
Valve timing
evo 114.6
0
aTDC
evc 393.4
0
aTDC
ivo 326.6
0
aTDC
ivc 605.4
0
aTDC
Lower calorific value 44MJ/kg


CNG (Test Engine 1)
Engine type (SI) Proton Magma (Aslam [26])
Cycle 4 stroke
Cylinder bore 7.55cm
Stroke 8.2cm
Connecting rod length 13.65cm
Compression ratio 9.2
Angle of ignition 40.00 bTDC
Lower calorific value 47.377MJ/kg


CNG (Test Engine 2)
Engine type (SI) Dongfeng Motor (Ma F [27])
Cycle 4 stroke
Cylinder bore 10.5cm
Stroke 12.0cm
Connecting rod length 19.2cm
Compression ratio 10.5
Angle of ignition 30.00 bTDC
Valve timing
evo 123.65 aTDC
evc 371.65 aTDC
ivo 341.65 aTDC
ivc 577.65 aTDC
Lower calorific value 47.377MJ/kg
International Journal of Energy and Environment (IJEE), Volume 1, Issue 1, 2010, pp.31-52
ISSN 2076-2895 (Print), ISSN 2076-2909 (Online) ©2010 International Energy & Environment Foundation. All rights reserved.

38
0
10
20
30
40
50
60
200 250 300 350 400 450 500 550 600
Crank angle (degree)
Cylinder pressure (bar)
ER=0.967Comp
ER=0.967 Exp.
ER=1.084 Comp.
ER=1.084 Exp.
ER=1.173 Comp.
ER=1.173 Exp.
Gasoline
CR=8.5
WOT 3000 RPM
Spark advance 36.4
0
bTDC
XSP=0.3517

Figure1. Comparison of measured and computed cylinder pressure with crank angle at different
equivalence ratio



30
35
40
45
50
55
60
0.7 0.8 0.9 1 1.1 1.2 1.3 1.4
Fuel-air equivalamce ratio
Maximum cylinder pressure (bar)
Pmax Comp.
Pmax Exp.
Gasoline
CR=8.5
WOT 3000 RPM
Spark advance 36.4
0
bTDC
XSP= 0.3517

Figure 2. Comparison of measured and computed maximum cylinder pressure
International Journal of Energy and Environment (IJEE), Volume 1, Issue 1, 2010, pp.31-52
ISSN 2076-2895 (Print), ISSN 2076-2909 (Online) ©2010 International Energy & Environment Foundation. All rights reserved.

39
0
1000
2000
3000
4000
5000
6000
0.7 0.8 0.9 1 1.1 1.2 1.3 1.4
Fuel air equivalance ratio
NO concentration (ppm)
0
1
2
3
4
5
6
7
8
9
10
CO concentration (V%)
Comp. NO
Exp. NO
Comp. CO
Exp. CO
Gasoline
CR=8.5
cofac=0.5;WOT 3000 RPM
XSP=0.3517
Spark advance 36.4
0
bTDC

Figure 3. Comparison of measured and computed NO and CO concentration at different fuel-air
equivalence ratio



8
9
10
11
12
13
0.7 0.8 0.9 1 1.1 1.2 1.3 1.4
Equivalence ratio
IMEP (bar)
EXP.
COMP.
Gasoline
CR=8.5
XSP=0.3517; WOT 3000 RPM
Spark advance 36.4
0
bTDC

Figure 4. Comparison of computed and experimental IMEP with variation of equivalence ratio
International Journal of Energy and Environment (IJEE), Volume 1, Issue 1, 2010, pp.31-52
ISSN 2076-2895 (Print), ISSN 2076-2909 (Online) ©2010 International Energy & Environment Foundation. All rights reserved.

40
0
15
30
45
60
75
200 250 300 350 400 450 500 550
Crank angle (degree)
Cylinder pressure (bar)
EXP.
COMP.
CNG [ F.Ma
27
]
WOT;Eqiv.ratio=0.7692
CR=10.5,1200RPM
spark angle 30
0
bTDC

Figure 5. Comparison of computed and experimental cylinder pressure with crank angle



0
0.2
0.4
0.6
0.8
1
1.2
500 1500 2500 3500 4500 5500 6500
Speed ( RPM )
BMEP (MPa )
0.2
0.35
0.5
0.65
0.8
0.95
BSFC (kg/kWh)
comp(BMEP)
exp(BMEP)
COMP(BSFC)
EXP(BSFC)
CNG[ Aslam
26
]
WOT;Eqiv.ratio=0.943
CR=9.2
spark angle 36
0
bTDC

Figure 6. Comparison of computed and experimental BMEP and BSFC with the variation of engine
speed


4. Results and discussion
Figure 7 shows the effect on in-cylinder pressure development with crank angle at different compression
ratio, at 3000 RPM wide open throttle and 36
0
bTDC spark timing. Figure illustrates that, with increase
in compression ratio peak pressure increases. This could be more clearly seen in Figure 8. This is
International Journal of Energy and Environment (IJEE), Volume 1, Issue 1, 2010, pp.31-52
ISSN 2076-2895 (Print), ISSN 2076-2909 (Online) ©2010 International Energy & Environment Foundation. All rights reserved.

41
because, the value of pressure and temperature of the mixture during sparking are higher at higher
compression ratios and heat release during combustion further increases the pressure and temperature to
the higher level. In the same time apparent flame speed increases and consequently combustion duration
decreases, this is shown in Figure 9.
The composition of the working mixture influences the rate of combustion and amount of heat evolved.
When the mixture is made leaner, it releases less thermal energy resulting in lower flame temperature
and hence flame speed reduces. With increase in compression ratio, the clearance volume decreases and
a greater portion of spent up gases are exhausted. This means, that there is less dilution of the charge and
charge density is greater during the burning process. In addition to this, the temperature of the charge is
high at higher compression ratio. Since the heat transfer increases with greater charge density and
increased temperature, the flame speed increases. At very lean air /fuel ratio, increasing compression
ratio results no significant change in flame speed. The combustion duration decreases with increase of
compression ratio, because the maximum cycle gas temperature increases resulting in higher apparent
flame speed. These factors greatly affect engine performance.
Figure 10 shows the effect of compression ratio on engine power, IMEP and BMEP, fuel consumption
and thermal efficiency. The increase in compression ratio shows the increase in mean effective pressure
and engine power, which is due to the higher cylinder gas pressure. The same figure also shows, as
compression ratio increases the specific fuel consumption decreases due to improved combustion with
higher peak pressure and temperature and due to more scope of expansion work.
Figure 11 shows the effect of compression ratio on CO and NO concentrations at exhaust-valve. It is
observed from the figure that as the compression ratio increases the concentrations of CO and NO
emission increase by increasing the maximum cycle temperature. In continuation, the effect of
compression ratio on CO and NO concentrations with fuel air equivalence ratio are shown in Figure 12.
In this figure equivalence ratio and spark timing are adjusted for maximum torque for BMEP data. This
also depicts that as compression ratio increases, the concentrations of NO and CO emission level
increase due to increase of cylinder temperature as shown in Figure 8.
It can be observed from Figure 10, higher the compression ratio higher is the indicated thermal
efficiency. The compression ratio is limited by two practical considerations, namely, material strength
and engine knock. Engine heads and blocks have a designed maximum stress, which should not be
exceeded, thus limiting the compression ratio. On the other hand, if maximum temperature exceeds the
auto-ignition temperature of air-fuel mixture, combustion will occur ahead of the flame, knock occurs
due to high rate of pressure rise causes damage to the engine.
Figure 13 shows the onset knock-limited compression ratio as a function of fuel-air equivalence ratio, for
two fuels with different octane number. The highest requirement is for slightly rich mixtures; increasing
richness and leanness, about this point decreases the octane requirement subsequently. Onset knock limit
depends on spark advance as shown in Figure 14: more advance sparking increases the peak pressure of
the cycle and therefore increases the pressure and temperature of the end charge resulting shortens delay
period (induction/auto ignition time) and increases the tendency to knock.
Figure 15 shows the onset knock-limited compression ratio as a function of equivalence ratio and
cylinder diameter. In this figure, XDS value is the ratio of cylinder diameter and fixed stroke length. The
tigure shows, initially onset knock limit decreases from 0.5 to 0.75 XDS and then increases with increase
in XDS value with increase in cylinder diameter. This is because high surface to volume ratio occurs for
smaller diameter engine, resulting in higher percentage of heat loss as shown in Figure 16.
The emission of carbon monoxide and nitric oxide also limits the use of higher compression ratio. The
concentrations of NO and CO emission depend upon compression ratio as well as spark plug location,
which is shown in Figures 17 and 18 respectively. Positioning of spark plug towards centre (XSP=0.125
to 0.5) and increasing compression ratio, NO and CO emission levels increase almost linearly in both the
conditions due to temperature increase inside the cylinder.
Figures 19 and 20 depict the effect of compression ratio and cylinder diameter to stroke length ratio
(XDS) on NO and CO emission levels at exhaust valve. These figures have been plotted for 0.9434
equivalence ratio, 3000 RPM engine speed and 36
0
bTDC spark timing. These figures illustrate that NO
and CO concentration levels increase by increasing XDS value and compression ratio. XDS is the ratio
of cylinder diameter and stroke length, where cylinder diameter is changed and stroke length is kept
constant. Smaller the value of XDS means smaller cylinder diameter, high surface to volume ratio,
resulting in higher percentage of heat loss as shown in Figure 16. Thus low temperature inside the
cylinder implies low emission of NO and CO concentrations.
International Journal of Energy and Environment (IJEE), Volume 1, Issue 1, 2010, pp.31-52
ISSN 2076-2895 (Print), ISSN 2076-2909 (Online) ©2010 International Energy & Environment Foundation. All rights reserved.

42
0
10
20
30
40
50
60
70
80
240 280 320 360 400 440 480 520
Crank angle (degee)
Cylinder pressure (bar)
CR=8
CR=9
CR=9.2
CR=10
CR=11
CR=12
CNG
WOT;Eqiv.ratio=0.943
3000 RPM
spark angle 36
0
bTDC
TDC

Figure 7. Variation of in-cylinder pressure with crank angle at different compression ratio




0
10
20
30
40
50
60
70
80
90
6 7 8 9 10 11 12 13 14
Compression ratio
Max.cylinder pressure (bar)
2560
2600
2640
2680
2720
2760
2800
Max. cylinder temperature (K)
Pmax
Tmax
CNG
WOT;Eqiv.ratio=0.943
3000 RPM;
spark angle 36
0
bTDC

Figure 8. Variation of maximum cylinder pressure and temperature with compression ratio



International Journal of Energy and Environment (IJEE), Volume 1, Issue 1, 2010, pp.31-52
ISSN 2076-2895 (Print), ISSN 2076-2909 (Online) ©2010 International Energy & Environment Foundation. All rights reserved.

43
61
61.5
62
62.5
63
63.5
64
6 7 8 9 10 11 12 13 14
Compression ratio
Combustion duration(degree)
14.7
15
15.3
15.6
15.9
16.2
16.5
16.8
Apparent flame speed (m)
Combustion duration
Apparent flame speed
CNG
WOT;Eqiv.ratio=0.943
3000 RPM
spark angle 36
0
bTDC

Figure 9. Variation of combustion duration and apparent flame speed at different compression
ratio


5
10
15
20
25
30
35
40
6 7 8 9 10 11 12 13 14
Compression ratio
IMEP(bar),BMEP(bar),IP(kW),
BP(kW),ITE(%)
0.15
0.17
0.19
0.21
0.23
0.25
0.27
0.29
ISFC (kg/kWh), BSFC (kg/kWh)
IMEP(bar) BMEP(bar)
IP(kW) BP (kW)
ITE(%) ISFC(kg/kWh)
BSFC(kg/kWh)
CNG
WOT;Eqiv.ratio=0.943
3000 RPM
spark angle 36
0
bTDC

Figure 10. Engine performances vs. compression ratio
International Journal of Energy and Environment (IJEE), Volume 1, Issue 1, 2010, pp.31-52
ISSN 2076-2895 (Print), ISSN 2076-2909 (Online) ©2010 International Energy & Environment Foundation. All rights reserved.

44
2000
2500
3000
3500
4000
4500
5000
67891011121314
Compression ratio
NO concentration (ppm)
0.3
0.34
0.38
0.42
0.46
0.5
CO concentration (%V)
NO
CO
CNG
WOT;Eqiv.ratio=0.943
3000 RPM
spark angle 36
0
bTDC

Figure 11. Variation of NO and CO concentration with compression ratio variation





0
500
1000
1500
2000
2500
3000
3500
0.7 0.8 0.9 1 1.1 1.2 1.3
Fuel-air equivalence ratio
NO concentration (ppm)
0
0.5
1
1.5
2
2.5
3
3.5
4
4.5
5
CO concentration (V%)
9.2 CR 10.0 CR
11.0CR 12.0 CR
14.0 CR 9.2 CR
10.0 CR 11.0CR
12.0 CR 14.0 CR
CNG
Wide open throttle
3000 RPM

Figure 12. Variation of NO and CO concentration vs. equivalence ratio at different CR with MBT
condition