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International Journal of Automotive Technology, Vol. 14, No. 4, pp. 539−549 (2013)
DOI 10.1007/s12239−013−0058−3
Copyright © 2013 KSAE/ 072−04
pISSN 1229−9138/ eISSN 1976−3832
539
OPTIMIZE COMBUSTION OF COMPRESSED NATURAL GAS ENGINE
BY IMPROVING IN-CYLINDER FLOWS
X. YU1), Z. LIU1), Z. WANG1)* and H. DOU2)
1)State Key Laboratory of Automotive Simulation and Control, Jilin University, Changchun 130025, China2)R&D Center, FAW Group Cooperation, Changchun 130011, China
(Received 25 September 2012; Revised 19 November 2012; Accepted 18 December 2012)
ABSTRACT−In order to solve the problem of slow flame propagation in a spark-ignition engine fueled with compressed
natural gas (CNG), the influence of in-cylinder flows on combustion process was investigated in CA6SE3-21E4N CNG-
engine by means of numerical simulation and experiment. The status of in-cylinder flows from intake to expansion stroke was
described by computational fluid dynamic tool, which revealed that the in-cylinder flows were one of the main reasons of slow
burning rate. Therefore, a special-shaped combustion chamber called Cross was used to improve the in-cylinder flows. The
results showed that peak turbulent kinetic energy of Cross was 43.9% higher than that of original combustion chamber called
Cylinder during the late compression period at 1450 rpm 100% load. The combustion parameters, brake specific fuel
consumption (BSFC) and regulated emissions were obtained by means of experiment. At 1450rpm 25%, 50%, 75% and 100%
load conditions, the ignition delay of Cross was longer than that of Cylinder, moreover, the Cross produced averagely 5.75oCA
shorter combustion duration. The BSFC of Cross was on an average of 4.3% reduction at 1450 rpm as well as the HC and CO
emissions were reduced whereas the NOx emissions were significantly increased.
KEY WORDS : CNG engine, In-cylinder flows, Combustion chamber, Combustion, Emissions
NOMENCLATURE
ATDC : after top dead center
BDC : bottom dead center
BSFC : brake specific fuel consumption
BTDC : before top dead center
CAD : crank angle duration
CNG : compressed natural gas
CO : carbon monoxide
COV : coefficient of variation
HC : hydrocarbon
HRR : heat release rate
IMEP : indicated mean effective pressure
NOx : nitrogen oxides
PFI : port fuel injection
TDC : top dead center
TKE : turbulent kinetic energy
WOT : wide open throttle
1. INTRODUCTION
Increasing awareness on energy crisis and environmental
pollution leads to a quest for alternative and clean burning
fuels of petroleum oil in the internal combustion engines.
Compressed natural gas (CNG), regarded as one of the
most promising alternative fuels (Aslam et al., 2006; Ruter
et al., 2012; Yao et al., 2011), is widely used in automotive
engines due to its rich resource and friendly emissions
(Middleton et al., 2008). However, flame propagation
velocity is slow during the combustion process in engines
when fueled with CNG due to its physiochemical
properties (Baratta et al., 2008; Patricia et al., 2011)
(present in Table 1). For this reason, the spark-ignition
CNG-engine always has the disadvantage of poor engine
power output and fuel economy. In order to solve this
problem, the methods like fueled with natural gas-
hydrogen blends (Liu et al., 2008; Bysveen, 2007) and
advance in-cylinder flows in the CNG-engine (Chiodi et
al., 2004; Prasad et al., 2011) are adopted. From the view
of in-cylinder material transport and distribution, it is well
known that the flame propagation velocity depends on in-
cylinder mixture concentration and flows (Cho et al.,
2007). Mixture concentration is controlled by injection
timing in intake multi-point injection CNG engines.
Therefore, flame propagation velocity of CNG engines
only depends on in-cylinder flows when the injection
timing is unchanged (Kaiadi et al., 2010; Ge et al., 2009).
As a result of it, reasonable organized in-cylinder flows can
accelerate the flame propagation velocity so as to increase
burning rate and to improve power performance, fuel
economy and emission performance of CNG-engine*Corresponding author. e-mail: [email protected]
540 X. YU, Z. LIU, Z. WANG and H. DOU
(Reddy and Abraham, 2010; Cho and He, 2008).
The in-cylinder flows are divided into two categories,
which are macrostructure and microstructure flows (Dinler
and Yucel, 2007). Large scale flow structures like swirl and
tumble belong to macrostructure, which will transform into
microstructure at the end; while microstructure, a kind of
small scale flow structures like turbulence, directly decides
the velocity of flame propagation (Abdullah et al., 2011).
Hence based on the two flow structures, there are two ways
to enhance in-cylinder flows in the port fuel injection (PFI)
spark-ignition engines. Specifically, one way is to enhance
large scale flows (swirl or tumble) through the improve-
ment of intake port structures (Jemni and Kantchev, 2011;
Qi et al., 2012); and the other directly enhance turbulence
level or squish intensity in the late compression period by
optimizing combustion chamber geometry (Simpson and
Olsen, 2010). Enhanced flow motions during intake stroke
are dissipated in the late compression stroke that limits its
potentiation with the restriction of engine operation loads.
Therefore, the latter approach is more effective in
comparison.
To analyze the influence of in-cylinder flows on
combustion process, researches are mostly proceeding in
optical engine. Ancimer et al. (2000) used discrete wavelet
transform to analyze the in-cylinder flows, which were
measured by laser Doppler velocimetry in an optical CNG-
engine. They observed that swirl structures would break
before top dead center (TDC) and were instable in the
beginning of expansion. The turbulent kinetic energy
(TKE) increased obviously in the flame-front region. The
heat release in the rapid combustion phase had strong
correlation with the average turbulence intensity. Huang et
al. (2005) studied the evolution processes of the in-cylinder
flows during the intake and compression strokes by using a
particle image velocimeter in motored two-valve single–
cylinder engine. The correlation between in-cylinder flows
and engine performance was analyzed and discussed by
using the quantified nondimensional parameters in their
research. Johansson and Olsson (1995) studied the in-
cylinder flows in a heavy-duty lean burn CNG-engine to
extend its lean burn limit. For an ideal in-cylinder flows
distribution, it was proposed that peak turbulent crank
angle should behind the TDC about 10oCA to acquire
higher heat release rate (HRR). These researches indicate
that, reasonable organized in-cylinder flows and enhance-
ment of turbulent intensity can be effective methods to
improve burning quality in CNG-engine.
Hence in the study aiming to optimize combustion, in-
cylinder flows were reasonably distributed to accelerate the
flame propagation velocity in CNG-engine. It was expect-
ed to achieve an advantageous status of turbulent evolution
history that has less TKE at ignition timing, to form stable
flame kernel and advances gradually to the peak in the
rapid burning phase, which helps the flame propagation
and increases the HRR. During the late phase of combus-
tion, TKE should be maintained to a higher level to reach
complete combustion.
In the previous experimental results of our team (Wang
et al., 2008), the cross combustion chamber is potent to
enhance in-cylinder flows. Given the high cost and
difficulty of optical engine researches, the in-cylinder
flows in different combustion chambers were revealed
throughout the course from intake valve opening to exhaust
valve opening, by means of computational fluid dynamic
tool. The combustion and emissions of CNG-engine were
investigated experimentally and were analyzed from the
view of flows. These results will provide theoretical basis
for realizing the high efficiency clean combustion of CNG-
engine.
2. NUMERICAL MODEL AND EXPERIMENTAL SETUP
2.1. Numerical Simulation Mesh and Boundary Conditions
The numerical CNG-engine model, which was modeled by
computational fluid dynamic tool, is shown in Figure 1. It
has about 370000 cells. In order to alleviate the influence
of pressure fluctuation on the in-cylinder flows in intake
port, a stable pressure box was set on the inlet entrance to
guarantee the stability of inlet gas.
The numerical simulation ran at 1450 rpm 100% load
condition, computed from intake valve opening (-50oCA)
to exhaust valve opening (432oCA), where the TDC of air
intake was defined as 0oCA and the TDC of compression
Table 1. Properties of CNG.
Fuel CNG
Lower heating value (MJ/kg) 50.05
Mixture heating value (MJ/m3) 3.39
Octane number (RON) 130
Cetane number <10
Ignition temp. of atmospheric pressure (oC) 537
Flame propagation velocity (m/s) 34-37
Figure 1. Numerical mesh of the CNG-engine at BDC.
OPTIMIZE COMBUSTION OF COMPRESSED NATURAL GAS ENGINE BY IMPROVING IN-CYLINDER FLOWS 541
was 360oCA. Figure 2 shows the valve lift profiles. The
CNG injection timing was 13oCA and its duration lasted
112oCA. The boundary conditions of simulation were
obtained from experiments and the cylinder wall conditions
were simplified adiabatically. The standard k-ε high
Reynolds number turbulence model and PISO computa-
tional algorithm were adopted. In addition, all simulations
focused on the in-cylinder flows without the involvement
of combustion simulation. Figure 3 shows the comparisons
of experimental and simulated cylinder pressure and HRR
at 1400 rpm 100% load condition. It can be seen that the
predicted results have a good agreement with the measured
data.
2.2. Test Engine and Measuring System
Experiments were conducted on a multi-point injection,
spark ignition, electronic controlled, lean burn CNG
CA6SE3-21E4N engine, which was converted from a
diesel engine. The retrofitted CNG engine maintained the
basic structures such as crankshaft connecting rod system,
cooling system and lubrication system, which in order to
commonly use the parts. But the 17.5 compression ratio in
diesel engine was not suitable for spark-ignited engine.
Therefore, the CNG engine decreased its compression ratio
to 12. The spark plugs were fixed into the fuel injector
location of the original engine. A new turbocharger should
Figure 2. Valve lift profiles.
Figure 3. Experimental and simulated pressure and HRR
curves at 1400 rpm 100% load.
Table 2. Specifications of test engine.
Engine type CA6SE3-21E4N
Bore (mm)×Stroke (mm) 106×125
Displacement volume (L) 6.618
Compression ratio 12
Rated power (kW)/speed (rpm) 155/2300
Peak torque (N·m)/speed (rpm) 700/1400
Injection mode PFI
Ignition system Spark ignition
Figure 4. Schematic of experimental setup.
542 X. YU, Z. LIU, Z. WANG and H. DOU
be matched to the CNG engine. The specifications of this
engine are presented in Table 2, and the layout of the setup
for experiments is shown in Figure 4. The engine was
equipped with a piezo-electric pressure transducer of type
Kistler 6125 to acquire cylinder pressures for the calcula-
tion of burning parameters such as heat release, ignition
delay, combustion duration, etc. Regulated emissions (HC,
CO and NOx) were measured with Horiba MEXA-
7100DEGR exhaust gas analyzer. For measurement of the
engine brake torque as well as the air mass flow and natural
gas flow, CW440-1 eddy current dynamometer, AVL1000
and CMF010 flowmeters were applied, respectively.
2.3. Experimental and Boundary Conditions
It was tested at wide open throttle (WOT) with a speed
range from 1000 rpm to 2300 rpm, namely, the speed
characteristic conditions. The combustion and emission
data were acquired to analyze the influence of in-cylinder
flows on the combustion process; and the ESC 13-mode
test cycle conditions also tested with the chosen operating
conditions of 25%, 50%, 75% and 100% loads at 1450 rpm
to analyze the combustion process and emissions. The
CNG-engine with different geometry of combustion
chambers used the same boundary conditions through the
test. The experimental boundary conditions of the main
operating condition points are shown in Table 3.
2.4. Geometry of Combustion Chambers
To improve the in-cylinder flows, a special-shaped
combustion chamber called Cross was used in the research.
The original combustion chamber called Cylinder was
cylindrical shaped. The schematic of combustion chambers
is shown in Figure 5, which indicates that the cross
combustion chamber adds four ribs around the combustion
chamber wall to enhance turbulent intensity and to
accelerate burning velocity. Both combustion chambers
ensure the compression ratio of CNG-engine being
unchanged.
3. RESULTS AND DISCUSSION
3.1. In-cylinder Flows
3.1.1. Macroscopic discussion of turbulence evolution
history
Figure 6 shows the comparison of turbulent evolution
history from intake to expansion stroke, which are the
numerical simulation results of the two combustion
chambers at 1450 rpm 100% load. As computational cells
have different mass, which provides the basis of average
TKE weighting herein. The TKE calculating equation (Jin,
2008) is shown as following:
(1)
Where the TKEmass,ave is mass averaged TKE, cset is the
investigated cells, i stands for the number of cells, m is the
mass of cell, C is the mass fraction.
As shown in Figure 6, TKE in the two combustion is
large. During the intake stroke, lots of annular vortexes will
break out and increase the TKE for the reason that the inlet
airflow strikes valve lever and the gas flow through the
intake valve produces shear layer. Although the TKE is
large, it cannot be directly utilized to advance combustion
velocity since the decay and dissipation make it difficult to
maintain a high TKE until the late of compression stroke.
The annular vortexes are influenced by geometry of intake
TKEmass ave,Ci *TKEi Ci
mi
mcset
----------=,
i
∑=
Table 3. Experimental boundary at 1450 rpm.
Engine speed(rpm)
Load (%)/torque (N.m)
Ignition timing (deg)
Injectiontiming (deg)
Excessair
ratio
1450
25/175 341 35 1.25
50/350 340 30 1.29
75/520 341.5 26 1.32
100/700 348 13 1.31
Figure 5. Geometry of combustion chambers. Figure 6. Evolution history of in-cylinder TKE.
OPTIMIZE COMBUSTION OF COMPRESSED NATURAL GAS ENGINE BY IMPROVING IN-CYLINDER FLOWS 543
port and valves lift, therefore, as shown in Figure 6, the
TKE of the two combustion chambers is approximately the
same during intake. It can be concluded that the geometry
of combustion chamber does not affect the in-cylinder
flows during the intake stroke. (It is similar to the
conclusions of Prasad et al. (2011)) However, during the
compression period, the decayed and dissipated TKE is
higher in Cross than that in Cylinder, as ribs of Cross
increase the flow velocity around the combustion chamber
wall. As a result of which, the peak TKE of Cross is 78.2%
higher than that of cylinder.
3.1.2. Microscopic discussion of in-cylinder flows
The in-cylinder overall average TKE indicates the
turbulence evolution history macroscopically. However,
the evolution of in-cylinder flows during the late period of
Table 4. Evolutional history of in-cylinder turbulence kinetic energy and velocity vectors from 300 to 396oCA.
CAD
Scale (m2/s2 and m/s)
Cylinder Cross Cylinder Cross
300 TKE
300 velocityvectors
324 TKE
324 velocityvectors
348 TKE
348 velocityvectors
372 TKE
372 velocityvectors
396 TKE
396 velocityvectors
544 X. YU, Z. LIU, Z. WANG and H. DOU
compression has significantly effect on combustion
process. Then it is analyzed in this section.
The view planes called A-A planes and B-B planes are
defined in Table 4 for micro analysis of in-cylinder flows.
The A-A planes cut through the center of intake and
exhaust valves and the B-B planes are diametral planes
through the center of combustion chamber, as shown in
Table 4. Table 4 lists the evolution history of in-cylinder
TKE and velocity vectors for both the cross and cylindrical
combustion chambers, which are in the two planes ranging
from 300 to 396oCA at a constant increment of 24oCA,
which cover the whole combustion process. The A-A
planes listed in Table 4 indicate that the distribution of
TKE in those two combustion chambers is almost the same
trend while at small crank angle of 300, which the cross
combustion chamber with ribs has limited advantage to
TKE. With the crank angle increasing, a higher TKE is
generated around ribs in cross combustion chamber and
diffused into the cylinder center and then raises the in-
cylinder flows all over the combustion chamber. However,
compared with Cross, the TKE of Cylinder is low and even
lower after the decayed process. The B-B planes listed in
Table 4 show that the TKE of cylindrical combustion
chamber is large in the cylinder center but decreasing
toward the direction of cylinder wall; on the contrary, from
cylinder center to its wall, the TKE gradually increases in
cross combustion chamber. As spark plug is located near
the cylinder center, the TKE distribution of cross
combustion chamber is advantage to the flame propagation
and then to the boost of burning speed (Tyagi and Ting,
2005). The main reason is that in-cylinder flows strike the
ribs, making orderly flows translate into turbulent motions
and a large number of turbulence is generated, which
consequently accelerates the flow velocity around the
combustion chamber wall. This phenomenon will improve
the quality of air-fuel mixture. Meanwhile, the higher TKE
extends the flame areas during combustion process so that
the burning speed is accelerated and the HRR is increased.
In Table 4, the velocity vector plots of A-A planes are
the resultant velocity of x axis and z axis, which indicate
the tumble and squish motions. The velocity vector plots of
B-B planes are the resultant velocity of x axis and y axis
that reveal the in-cylinder swirl motions. The A-A velocity
vetors planes show that flow motions of Clinder are more
uniformed than that of Cross. As the circle places show,
cross combustion chamber boosts the squish and tumble
motions. Thess flow motions are effectively decay mean
flow into turbulent flow, which are advantage to the
efficient combustion. As B-B velocity vectors planes show,
flow motions around the ribs of Cross are more turbulent,
which are the main reason for higher TKE near the cylinder
wall in Cross.
In general, the cross combustion chamber generates a
large number of turbulence during the late period of
compression due to its ribs that increase the in-cylinder
TKE and flow motions, and then the flame propagation
velocity as well as the rate of heat release are increased.
3.2. Experimental Combustion Process
For a comprehensive understanding of the test engine,
Figure 7 shows the speed characteristic curves of the CNG-
engine with cylindrical and cross combustion chamber. The
brake specific fuel consumption (BSFC) and exhaust gas
temperature of cross combustion chamber, as shown in
Figure 7, are on an average of 5.2%, 23.7K lower than that
of cylindrical combustion chamber over the speed range,
respectively. The lower BSFC indicates that the CNG-
engine has better combustion quality with cross combustion
chamber; meanwhile, the lower exhaust gas temperature of
cross combustion chamber will reduce the heat load of
engine and then prolongs the engine life.
3.2.1. Average TKE and HRR
The main purpose of the in-cylinder flows research is to
optimize combustion process and to increase HRR.
Nevertheless, the influence of in-cylinder flows on
combustion concentrates on the late period of compression
stroke. For this reason, the correlation between the TKE
distribution of late period of compression and combustion
at 1450 rpm 100% load is analyzed. Figure 8 shows the
comparison of averaged TKE and HRR over the range
from 300 to 432oCA. As shown in Figure 8, the Cross has a
higher TKE than the Cylinder and its crank angle of peak
TKE is close to TDC. The peak TKE crank angle of Cross
presents at 357oCA while that of Cylinder presents at
345oCA. Figure 8 also shows that the crank angle duration
between peak TKE and peak HRR is 19.5oCA of Cross,
which is 10.5oCA shorter than that of Cylinder. As a result,
a delayed peak TKE crank angle and a short duration
between peak TKE and peak HRR will concentrate the heat
release, decrease the combustion duration, accelerate the
burning velocity, improve the constant volume degree and
then increase the heat release; therefore, it can be
concluded that the TKE distribution of Cross is more
advantage to the CNG-engine combustion.
Figure 7. BSFC and exhaust gas temperature versus engine
speed from 1000 to 2300 rpm at WOT.
OPTIMIZE COMBUSTION OF COMPRESSED NATURAL GAS ENGINE BY IMPROVING IN-CYLINDER FLOWS 545
Compared with the HRR curve of Cross, as shown in
Figure 8, the Cylinder HRR is a little higher in the early
period of heat release at 1450 rpm 100% load. Subsequently,
the HRR of Cross gradually increases during the
combustion process and is higher than that of Cylinder in
the later period at the operating condition. The heat release
of Cross is more intensive with its peak 11.1% higher than
that of Cylinder at 1450 rpm 100% load. As to detailedly
analyze the influence of in-cylinder flows on the
combustion process, the experimental burning data with
the two combustion chambers are presented in the
following sections at 1450 rpm.
3.2.2. Ignition delay
The crank angle duration (CAD) called ignition delay is
between the ignition and the start of combustion (Geok et
al., 2009), as shown in Figure 9. The start of combustion is
defined as the crank angle of 10% fuel mass burned.
Shown in Figure 9, at each load of 1450 rpm, the cross
combustion chamber has longer ignition delay than the
cylindrical combustion chamber. The results indicate that
the flame kernel develops more easily and stably in
cylindrical combustion chamber.
Cho and He (2007) have indicated that the ignition delay
period is mainly influenced by air-fuel mixture concentra-
tion and flow motions around the spark plug. In order to
explain the tendency of ignition delay in the two
combustion chambers, the numerical data, which are TKE
and mixture concentration of spark plug cells, are
presented in Figure 10 during the ignition delay period at
1450 rpm 100% load. As the figure shows, the two
combustion chambers have almost the same mixture
concentration in the spark plug but the TKE is higher in
Cross. Due to that the large TKE is disadvantage to the
flame development, the cross combustion chamber has
longer ignition delay period.
3.2.3. Combustion duration and CA50
The above results indicate that appropriate low TKE is
advantage to the flame development. Nevertheless, after
the flame development period, the combustion needs
higher turbulence level to enhance the flame propagation
which consequently increases the heat release. Figure 11
shows the experimental combustion duration and CA50
data comparison of the two combustion chambers at 1450
rpm. The combustion duration is defined as the crank angle
duration between 10% and 90% fuel mass burned
(Einewall et al., 2005). The CA50 is defined as the crank
angle of 50% fuel mass burned (Massey et al., 2009). As
shown in Figure 11, compared with Cylinder, Cross has an
Figure 8. Average TKE and HRR as a function of CAD at
1450 rpm 100% load.
Figure 9. Comparison of ignition delay period at 1450 rpm.
Figure 10. Average TKE and excess air ratio of spark plug
cells at 1450 rpm 100% load.
Figure 11. Combustion duration and CA50 at 1450 rpm.
546 X. YU, Z. LIU, Z. WANG and H. DOU
average of 5.75oCA shorter combustion duration at each
load of 1450 rpm. Moreover, its CA50 averagely moves
forward to 1.83oCA. It indicates that the in-cylinder flows
of cross combustion chamber improve burning rate so that
realize rapid combustion.
3.2.4. Heat release and stablility
As a result in the obove section, the shorter combustion
duration and forward CA50 crank angle degree of Cross
lead to the combustion pressure peak and HRR peak are on
an average of 3.5% and 14.1% higher than that of Cylinder
at 1450 rpm, respectively, as shown in Figure 12.
In order to evaluate the overall performace of the CNG
engine at 1450 rpm, Figure 13 shows the coefficient of
variation (COV) at each load to indicate the cycle-by-cycle
variations. These data is based on 200 combustion cycles
that acquise by combustion analyser. The COV calculating
equation is shown as following:
(2)
Where the σpmi is standard deviation of indicated mean
effective pressure (IMEP), pmi is the arithmetic average of
IMEP.
Figure 13 shows that the cycle-by-cycle variations in
Cross and Cylinder combustion chambers are little
difference. The COV of those two combustion chambers is
much lower than 5%, which indicates the stability
performance in the engine.
Furthermore, Figure 14 and Figure 15 present a
comparison of the numerical simulation data of TKE
distribution of the two combustion chambers to support the
above experimental conclusions. The virtual sampling
points are positioned in the combustion chambers to
present the in-cylinder flows distribution. Figure 14 shows
the schematic of sampling points. The axial sampling
points are placed on the center of axial plane of combustion
chamber with their marked numbers increasing from top to
bottom of the combustion chamber. The radial sampling
points are placed on the central radial plane of combustion
chamber with their marked numbers increasing from the
center to the combustion chamber wall. Due to the different
size and shape of mesh in the two combustion chambers,
the number of sampling points of Cylinder and Cross are
different. But with the same total length and the direction
of sampling points, the comparison of flow distribution is
valid.
As Figure 15 shows, the radial and axial distribution data
of TKE in the two combustion chambers are obtained from
numerical simulation at 364oCA (the crank angle at the
early stage of rapid burning period), 1450 rpm 100% load
condition. Figure 15 (a) shows that the radial TKE of cross
combustion chamber increases from center to combustion
chamber wall, which is caused by ribs around the
combustion chamber wall. The flame developing from the
center to wall make the TKE distribution advantageous to
the exchange between burned and unburned gas in the
region of flame-front, which can be further extended to
accelerate the burning velocity as a consequence. On the
contrary, the TKE distribution of cylindrical combustion
chamber, shown in Figure 15 (a), is apt to cause incomplete
combustion and decrease the thermal efficiency. From
Figure 15 (b), it can be seen that the axial TKE presents
same distribution in both combustion chambers, but it is
overall greater in Cross than in Cylinder. These two
reasons (it can be also acquired in Table 4) indicate the
well-restructured in-cylinder flows of Cross, which
advance the turbulence level and then accelerate burning
velocity. At this point, the experimental combustion
duration conclusions are done in the above section.
COV σpmi pmi⁄=
Figure 12. Peak of combustion pressure and HRR at 1450
rpm.
Figure 13. Cycle-by-cycle variations at 1450 rpm. Figure 14. Schematic of sampling points.
OPTIMIZE COMBUSTION OF COMPRESSED NATURAL GAS ENGINE BY IMPROVING IN-CYLINDER FLOWS 547
3.3. Fuel Consumption and Emissions
The experimental fuel consumption and emission data are
presented in the following sections to demonstrate the
influence of in-cylinder flows on actual CNG-engine.
Figure 16 shows the BSFC comparison at 1450 rpm with
loads from 25% to 100%. The BSFC of cross combustion
chamber is on an average of 4.3% lower than that of
cylindrical combustion chamber. It indicates that the
combustion in cross combustion chamber is more
complete, due to the restructured in-cylinder flows.
Figure 17 shows the comparison of CO, HC and NOx
emissions at 1450 rpm with 25%, 50%, 75% and 100%
load conditions. As shown in Figure 17 (a) and Figure 17
(b), CO and HC emissions are lower in Cross than in
Cylinder. However, the NOx emissions increase in Cross,
reaching the average of 17.7% higher than that of Cylinder,
as shown in Figure 17 (c). The special-shaped combustion
chamber enhances the in-cylinder turbulence level at the
late period of compression. Consequently, it improves
burning process and achieves rapid combustion. Therefore,
the incomplete products of CO and HC are decreased due
to the more complete combustion. Nevertheless, high-
quality combustion will increase the in-cylinder combus-
tion temperature under which NOx emissions are increased
in cross combustion chamber.
4. CONCLUSION
The conclusions of this numerical simulation and
Figure 15. Radial and axial flows distrubution of in-
cylinder TKE at 364oCA for 1450 rpm 100% load.
Figure 16. BSFC at 1450 rpm for the two combustion
chambers.
Figure 17. Exhaust emissions at 1450 rpm for the two
combustion chambers.
548 X. YU, Z. LIU, Z. WANG and H. DOU
experimental research can be summarized as follows:
(1) The in-cylinder flows can be reasonably restructured by
improving the geometry of combustion chamber. During
the intake, the geometry of combustion chamber dose not
influence the in-cylinder flows. However, in the late
period of compression, it has substantial effect.
(2) It indicates that the radial TKE of cross combustion
chamber increases from center to cylinder wall in the
rapid burning period. The axial TKE of the two
combustion chambers presents the same tendency.
(3) During the compression stroke, at 1450 rpm 100% load
condition, the cross combustion chamber has overall
higher TKE than that of cylindrical combustion
chamber, moreover, its TKE peak is 43.9% higher and
the corresponding crank angle is more closer to TDC.
The crank angle duration of cross combustion chamber
between the peak of TKE and HRR is 10.5oCA shorter
than that of cylindrical combustion chamber.
(4) At the experimental speed characteristic conditions,
compared with the cylindrical combustion chamber,
the BSFC and exhaust gas temperature of Cross are on
an average of 5.2% and 23.7K lower, respectively.
(5) The ignition delay of cross combustion chamber is
longer than that of cylindrical combustion chamber at
1450 rpm from 25% to 100% load. Moreover, cross
combustion chamber has an average of 5.75oCA
shorter combustion duration than Cylinder and the
CA50 of Cross moves forward to 1.83oCA. The peak of
combustion pressure and HRR of cross combustion
chamber are averagely 3.5% and 14.1% larger than that
of cylindrical combustion chamber at 1450 rpm,
respectively. The cycle-by-cycle varations in the two
combustion chamber are all lower than 5%.
(6) At 1450 rpm with the load from 25% to 100%, the
BSFC of cross combustion chamber is averagely 4.3%
lower than that of cylindrical combustion chamber. The
CO and HC emissions are all lower in cross
combustion chamber but the NOx emissions increase
by 17.7%.
ACKNOWLEDGMENT−The work was supported by the
Natural Science Foundation of China (Grant NO. 50906033) and
the Ministry of Science and Technology through its 973 National
Key Project of China (2013CB228402).
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