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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. YU 1) , Z. LIU 1) , Z. WANG 1)* and H. DOU 2) 1) State Key Laboratory of Automotive Simulation and Control, Jilin University, Changchun 130025, China 2) 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 5.75 o CA 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]

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Page 1: Optimize combustion of compressed natural gas engine by improving in-cylinder flows

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]

Page 2: Optimize combustion of compressed natural gas engine by improving in-cylinder flows

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.

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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.

Page 4: Optimize combustion of compressed natural gas engine by improving in-cylinder flows

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.

Page 5: Optimize combustion of compressed natural gas engine by improving in-cylinder flows

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

Page 6: Optimize combustion of compressed natural gas engine by improving in-cylinder flows

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.

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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.

Page 8: Optimize combustion of compressed natural gas engine by improving in-cylinder flows

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.

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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.

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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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