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ABSTRACT

This paper presents the simulation of in-cylinder stratified

mixture formation, spray motion, combustion and emissions

in a four-stroke and four valves direct injection spark ignition

(DISI) engine with a pent-roof combustion chamber by the

computational fluid dynamics (CFD) code. The Extended

Coherent Flame Combustion Model (ECFM), implemented in

the AVL-Fire codes, was employed. The key parameters of

spray characteristics related to computing settings, such asskew angle, cone angle and flow per pulse width with

experimental measurements were compared.

The numerical analysis is mainly focused on how the tumble

flow ratio and geometry of piston bowls affect the motion of

charge/spray in-cylinder, the formation of stratified mixture

and the combustion and emissions (NO and CO2) for the

wall-guided stratified-charge spark-ignition DISI engine. But

due to the fuel injected during compression stroke, the effect

of intake ports and exhaust ports were not taken into

consideration in this study. It is found that the geometry of

piston bowls has a major effect on the mixture stratification

in-cylinder, the combustion process and others. In addition,

the characteristics of the charge motion and combustion, such

as mean in-cylinder pressure, heat release rate and

accumulated heat release vary as a function of crank angle at

different injection timings and tumble flow ratios, based on

one of two combustion geometries. The results show that the

injection timing and piston bowl shape play very important

roles for the combustion process and mixture stratification.

Further more, the simulation provides an insight into the

interaction of charge flow, fuel spray, piston bowl as well as

combustion.

INTRODUCTION

With the increasing attention on achieving substantia

improvements of fuel economy and reductions of exhaust

emissions, automotive engineers are striving to develop

engines with lower Brake Specific Fuel Consumption

(BSFC), and which can also comply with future stringent

emission requirements. Over the past two decades, many

attempts had been made to develop an internal combustion

engine for automotive applications that combines the best

features of the spark ignition (SI) and the compression

ignition (CI) engines. The objective is to combine the specific

power of the gasoline engine with the efficiency of the diese

engine at part load. Such an engine would exhibit a BSFC

approaching that of the diesel engine, while maintaining the

operating characteristics and specific power output of the SI

engine [1].The direct injection spark ignition (DISI) engines

in theory, have these two merits. On the one hand, the fuel is

injected directly into the combustion chamber in order to

have the mixture clouds with an ignitable composition nearthe spark plug.

In general, the direct injection of the fuel allows for two

distinctly different combustion strategies. The first is to form

the homogeneous charge by early injection during the intake

stroke, due to allowing enough time for the fuel vaporization

and fuel-air mixing. Load control is achieved by appropriate

throttling similar to a multipoint port-fuel-injected (PFI)

engine. The second is to realize stratified charge by late

injection during compression, in which a compact fuel-rich

Stratified Mixture Formation and Combustion

Process for Wall-guided Stratified-charge DISIEngines with Different Piston Bowls by Simulation

2010-01-0595

Published

04/12/2010

Qianwang Fan, Zongjie Hu, Jun Deng and Liguang LiTongji Univ.

Yi You and Jingyan HuGeely Automobile Research Institute

Copyright © 2010 SAE Internationa

转载

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cloud is formed around the spark gap with an overall lean

mixture. In reality, the full potential of the DISI combustion

system is achieved by utilizing both strategies. At high load

the engine should utilize homogeneous charge to maximize

air utilization and to avoid soot formation. A slight gain in

fuel efficiency is still achieved by charge cooling. Stratified

charge is desirable at part load to attain diesel-engine-like

fuel economy. At this mode, pumping losses are minimizedsince engines can operate virtually unthrottled.

During the last several years, the DISI engine has been

greatly evolved, and various types of DISI productions have

been put into the market. Moreover, a lot of investigations

have been achieved by experiments and CFD simulations.

Especially, with the development of computers and numerical

computation methods, in the automotive industry, three-

dimensional, CFD combustion simulation is increasingly

becoming an effective tool for engine design and

development. In the past, the application of transient CFD in-

cylinder analysis has mainly been focused on the intake and

compression strokes, calculating in-cylinder transient tumble

curves for an assessment of probable combustion stability,

and on air-fuel mixture formation for combustibility [3, 4].

Luca Olmo [5] presented that the in-cylinder flow and

combustion analysis of a high performance four valves DISI

engine has been successfully carried out for three different

design configurations, adopting the ECFM combustion model

implemented in the STAR-CD code. Jianwen Yi et al [6]

studied the interaction between in-cylinder flow and injected

fuel spray in DISI engines, the results showed that the major

effect of intake flow on the fuel spray is that the induced flow

tends to make the spray collapse. Sungjun Kim et al [7]investigated the air-fuel mixture formation and combustion

characteristics in spray-guided DISI engine employing a Star-

CD code. The results showed the enhanced tumble flow can

deteriorate the mixture distribution and decrease the burning

rate. Moreover, G. Fontana et al [8] studied the effect of the

different combustion chambers on engine performances and

emissions of a small gasoline engine, employing numerical

and experimental methods.

Although previously a lot of investigations had been

performed, including stratified-charge formation [9,10] and

the effect of various parameters on engine performances and

emissions [11,12] etc. The objectives of this paper are toinvestigate the effects of tumble flow, injection timing and

geometry of piston bowl on the motion of in-cylinder charge-

spray, formation of stratified mixture, combustion and

emissions (NO and CO2) for the wall-guided stratified-charge

spark-ignition DISI engine, when gasoline fuel is injected

during late compression stroke. The aim of this investigation

is to provide some supports for DISI engine design and

parameter calibration.

ENGINE SPECIFICATIONS

Engine SpecificationsEach cylinder of the four-cylinder engine which is employed

has a pentroof head, four valves (two intake valves and two

exhaust valves), a centrally mounted spark plug, a piston with

alternative bowls (shown in Fig.1), and an six-hole injector

mounted under intake port at the inclination angle of 28

degrees with respect to the cylinder head, whose holes are not

located evenly (shown in Fig.2). The Fig.1 shows two

pistons, piston A and piston B. In addition, in order to adap

each piston bowl shape, the cylinder head was also modified

slightly, and the injector installation exists in difference

namely, for piston A, the maximal included angle between

two nozzle holes is located near the side of the cylinder head

but for piston B, one located near the side of piston top. Fo

the injector, the spray characteristics related to computation

later, were given based on the experimental data, which are

shown in Fig 3 and Table 1. In addition, the engine

specification is summarized in Table 1 as well.

(a). Scheme of piston A

(b). Scheme of piston B

Fig. 1. The scheme of piston bowl shape

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Fig. 2. The scheme of nozzle holes' distribution

Fig.3. the scheme of spray event in ambient conditions

Table 1. Engine specification and injector parameters

Note: 0 Degree Crank Angle is assigned in TDC intake

stroke

BASIC THEORY AND MODEL

The CFD calculations were performed with AVL Fire codes

The codes solve ensemble-averaged equations for

momentum, energy, species concentration and mass. The

calculations start from the time of intake valve close (IVC

220°CA ATDC) to the time of exhaust valve open (EVO

510°CA ATDC), calculating step lengths are 0.2°CA during

fuel injection and combustion, others are 0.5°CA. During the

whole process, the flow motion event was computed using

the turbulence model, the spray and combustion events

were performed using Discrete Droplet Model (DDM) and

Extended Coherent Flame Model (ECFM). In additionequations of energy, motion and mass were solved using

Upwind Scheme, Central Differencing Scheme and

MINMOD Relaxed Scheme [13,14,16], respectively. The

theories associated with the models were to be simply

introduced below, for details, referring to the literatures

[13,14,15,16].

MeshIn the study, the unstructured grids were adopted, while the

piston was modeled by the dynamic mesh models. The entire

process simulating piston movement was achieved by the

dynamic grids, which were completed using FAME Meshing

(FM) and FAME Engine Plus (FEP) tools. Both of models, a

most, were meshed up to 300,000 grids, including refining

parts. Specific grid models are given in Fig.4 and Fig.5.




Fig.4. The scheme of piston A

Fig.5. The scheme of piston B

Initial Conditions and Boundary Conditions

SettingThe initial conditions and boundary conditions are provided

resulting from computational values of GT-Power code, and

some referred experimental values. Table.2 shows essential

values of initial and boundary conditions.

Table.2. Initial Conditions and Boundary Conditions

Setting

Note:* Tumble flow ratio is defined as tumble flow

velocity divided by engine revolved velocity; Ignition

duration is 0.0003 seconds, equaling to 3.6°CA at the

speed of 2000 r/min.

Turbulent flow modelThe k-ε model is used as turbulence model. Here, the k

equation [13, 15] and ε equation [13, 15] are expressed

respectively, as follows:

(1)

Where, is the turbulent kinetic energy, is the tensor o

strain rate associated to fluctuating flow; is the tensor o

strain rate associated to mean flow; is kinematic viscosity

coefficient.

(2)

Where, is the density, is the dissipation rate, is the

Prandtl number of , is the effective viscosity

coefficient, , are the constant, is the turbulen

kinetic energy production term.

Combustion modelThe Extended Coherent Flame Model (ECFM) [13, 15] has

been mainly developed in order to describe combustion in

DISI engines. This model is fully coupled to the spray model

and enables stratified combustion modeling including EGR

effects and NO formation. The model relies on a conditiona

unburnt/burnt description of the thermochemical properties of

the gas. The ECFM contains all the features of the standard

CFM and the improvements of the MCFM. Differences to the

other coherent flame models are described in the reference

[13, 15].

For turbulent combustion phenomena, the ECFM model leads

to the calculation of the mean fuel reaction rate. Hence, this

model uses a 2-step chemistry mechanism [13] for the fue

conversion like:

(3)

(4)




whereafter, some basic equations are introduced below.

The mean turbulent fuel reaction rate [13] is computed as the

product of the flame surface density Σ and the laminar

burning velocity SL via:

(5)

Where, is the stoichiometric coefficients of species in

the reaction r, while for the reactants these coefficients are

negative and for the products positive, respectively. is

the reaction rate for reaction (3) and (4). is the mean

laminar fuel consumption rate described earlier. is a

function depending on the equivalence ratio , number of

carbon and hydrogen atoms, respectively.

From the previous sections it is obvious that the extended

CFM can be closed if the local properties of the burnt and

unburnt gases are known. Hence, in each computational cell

two concentrations are calculated: a concentration in the

unburnt gases and a concentration in the burnt gases,

respectively. Hence two additional transport equations have

to be introduced, one for the unburnt fuel mass fraction and

one for the unburnt oxygen mass fraction. Below the two

transport equations [13] are given.

(6)

(7)

Additionally, a transport equation for the unburnt gas

enthalpy is also introduced as shown below.

(8)

Where, in the equations (6) (7) (8), is density, is the

fuel mass fraction in the fresh gas, is the source term

for the unburnt fuel mass fraction in case of spray

applications, is the source term in case of evaporation

of the liquid fuel, is the mass fraction, is the fresh

gas enthalpy.

Spray modelSpray simulations involve multi-phase flow phenomena and

as such require the numerical solution of conservation

equations for the gas and the liquid phase simultaneously

With respect to the liquid phase, practically all spray

calculations in the engineering environment today are based

on a statistical method referred to as the Discrete Droplet

Method (DDM) [13, 15].

Movement equation

(9)

Energy equation (heat transfer)

(10)

Mass equation (evaporation)

(11)

Where, is the Particle velocity vector [m/s], is the

Particle diameter [m], is the droplet density [kg/m3],

is the gas density [kg/m3], is the gas velocity [m/s], is

the droplet velocity [m/s]; is the Particle mass [kg],

is the Specific heat of liquid [J/(kg.K)], is the drople

temperature [K], is the particle surface temperature [K]

is the Nusselt number, is the Particle far-field

temperature [K], λ is the Thermal conductivity [W/(m.K)].

RESULTS AND DISCUSSION

The effect of piston bowl on combustion

and emissionsWith regard to the effect of piston bowl geometry on

combustion and emissions, the simulations were performed a

the injection timing of 310 °CA ATDC and the ignition




timing of 345 °CA ATDC, by varying tumble flow ratio for

two different piston bowl shapes.

Interaction between flow and spray

Fig.6 shows the interaction between charge flow and spray

inside combustion chamber for the stratified-charge direct-

injection gasoline engine. When the spray is induced into

cylinder, in-cylinder single vortex is transformed two ones

around the spray for each piston. Before the spray impinges

the piston bowl surface, many droplets change their

directions and move perpendicularly to the spray axis toward

the top of the combustion chamber. The reason for this

drastic change in the flow direction is the strong influence of

the in-cylinder airflow on the smaller droplets [17]. Then, at

the end of injection, the spray head direction is changed along

piston bowl surface for each piston, due to the effect of piston

bowl profile and tumble flow. As far as two pistons are

concerned, the piston B seems to be more suitable for mixture

stratification at the injection timing of 310 °CA ATDC.

<figure 6 here>

In addition, different tumble intensity causes the spray

velocity difference along the piston bowl contour for two

piston bowls. Moreover, there exists a distinct spray plume

structure inside combustion chamber with two different

pistons, mainly resulting from asymmetric distribution of

holes, and as mentioned before, the nozzle holes are located

differently around the axis of the hole where the injector is

installed.

In-cylinder mixture strength

The mixture formation has a significant impact on

combustion and emissions. Therefore, it is quite important to

study in-cylinder fuel-air distribution. The references [3, 4, 5,

10] presented the results of mixture formation. In those

investigations, the spatial distribution and temporal history of

mixture formation were studied. Besides, mixture strength is

designated as equivalence ratio.

For an engine operation in the stratified charge mode, the

spatial distribution and temporal development of the mixture

strength are shown in Fig. 7. At the end of fuel injection (324

°CA ATDC), the thickest mixture exists inside the down-

right side of combustion chamber, due mainly to theincomplete evaporated spray droplets. As the piston moves

up, the thick mixture also reaches near the spark plug,

resulting mainly from the interaction among spay, piston

bowl profile and tumble flow. Regardless of piston A or

piston B, the tumble flow plays an important role in mixture

formation. In addition, Fig.7 also shows that fuel-air mixture

is stratified by two types of piston bowls, due to the

interaction between spray and piston top. However, compared

to the piston A, the piston B is much fitter to stratified-charge

mixture formation, when fuel is injected inside the cylinder at

the time of 310 °CA ATDC. At the same time, as shown in

Fig.8(d), it is also acquired, for the piston B under tumble

flow ratio (TR) of 1, that the appropriate mixture around the

spark gap is formed, resulting mainly from the effect o

intensified tumble flow.

<figure 7 here>

Combustion Characteristics

Fig.8 shows the spatial distribution and temporal history of

in-cylinder temperature throughout the combustion chamber

Firstly, it shows that the spatial distribution of in-cylinder

temperature for the piston bowl B is superior to that of the

piston bowl A, which is the result of favorable stratified

mixture formation for the piston bowl B. Furthermore, it can

be also found that tumble intensity has a strong influence on

the in-cylinder temperature. Lastly, it implies that the

stratified mixture leads to in-cylinder temperature

stratification throughout the combustion chamber, especially

for the piston bowl B, it is more obvious as the result ofappropriate stratified mixture and stable combustion.

<figure 8 here>

Fig.9. Effect of piston bowl on combustion




Fig.10. Effect of piston bowl on mean in-cylinder

temperature at various tumble flow ratios

Fig.9 shows the comparisons of mean in-cylinder pressure,

rate of heat release and accumulated heat release between the

piston A and the piston B under different tumble flow ratio

(TR). It can be found that the mean pressure is some slightly

different between the piston A and the piston B. However, the

strong tumble flow intensity leads to mean pressure increase

due to stronger tumble flow improving mixture formation and

accelerating combustion process, but tumble flow is too

strong to form proper air-fuel mixture in the position of spark

plug, so tumble intensity is enhanced properly.

Furthermore, as far as the accumulated heat release in Fig.9concerned, compared to the piston B, the piston A is sensitive

to tumble flow intensity, due to piston A shape resulting in

the worse air-fuel mixture with the TR of 0.5. Hence it is

clearly effective for the piston A with the TR of 1.

For the piston B, the rate of heat release and the accumulated

heat release are higher than that of the piston A, regardless of

tumble intensity. Moreover, the strong tumble flow causes

them to rise greatly. The main reasons are that the fast-

burning period becomes shorter, and that the location of area

center in the curve of heat release rate is closer to the top

dead center (TDC) and the combustion process is remarkably

improved [18].

Fig.10 shows the effect of piston bowl shape on mean in-

cylinder temperature under different tumble flow ratio (TR).

It shows that the mean in-cylinder temperature for the piston

B is higher than that of the piston A, as the result that, when

compared to the piston A, the spatial distribution of air-fuel

mixtures for the piston B is better throughout the whole

combustion chamber, and acquiring a stable, stratified-charge

combustion. In addition, the tumble flow motion not only

influences the stratification mixture formation, but also

affects flame kernel formation and flame propagation, hence

with the tumble flow ratio of 1, in-cylinder combustion

process is satisfying, due mainly to appropriate tumble flow

intensity improving mixture formation and accelerating

robust combustion.

For the piston A, when the tumble flow ratio is the value of

0.5, there exist two distinct peak values of mean in-cylindertemperature. The second peak value arises resulting mainly

from the strong post-combustion.

Emissions

Fig.11 shows the effect of piston bowl types on CO2

emissions at various tumble flow ratios. Regardless of tumble

flow ratio, for the piston B, CO2 emissions, are generated

more than that of the piston A, due to mainly appropriate

stratified-charge mixture formation and relatively complete

combustion. On the other hand, regardless of piston bow

shape, the stronger tumble flow promotes more appropriate

stratified-charge mixture formation and more adequatecombustion, leading to more CO2 emissions generation.

Combined with the accumulated heat release in Fig.9, it is

higher for the piston A than that of the piston bowl B with the

TR of 1. However, it shows that carbon dioxide is less for the

piston A, as its piston bowl shape resulting in different air

fuel mixture distribution. Although it is clearly effective for

air-fuel mixture formation of the piston bowl A to enhance

tumble flow, but compared to the piston B, air-fuel mixture

stratification is worse for the piston A, so a lot of carbon

monoxide emission is generated. Besides, the unburn

hydrocarbon is more for the piston B due to better air-fuel

mixture stratification resulting in quenching in the position

far from the spark plug.

Fig.12 shows the effect of piston bowl types on NO

emissions formation at various tumble flow ratios. Generally

speaking, the NOx formation strongly depends on the

combustion temperature and combustion duration at high

temperature. Likewise, for piston B, NO emissions are

generated more than that of the piston A, regardless of tumble

flow ratio, due to mainly the piston B fitting more to

stratified-charge mixture formation and robust combustion

On the other hand, for each piston bowl, the stronger tumble

flow is formed, the more amount of NO emissions isgenerated. Moreover, for piston A at tumble flow ratio of 0.5

the amount of NO emissions equals nearly to zero, as the

result of lower combustion temperature.




Fig.11. Effect of piston bowl on CO2 emissions

Fig.12. Effect of piston bowl on NO emissions

The effect of tumble flow on combustion

and emissions for piston BIt is known that in-cylinder airflow fields have the prominent

influence on combustion and emission characteristics.

Therefore, the section presents the effect of the tumbleintensity on mean in-cylinder pressure, the rate of heat release

and the accumulated heat release at injection timing at 278

°CA ATDC for the piston B.

Fig.13 shows the effect of different tumble flow ratios on

combustion characteristics, including in-cylinder mean

pressure, rate of heat release and accumulated heat release.

The strong tumble flow contributes to stratified-charge

formation at early stage of compression stroke and to

turbulence intensity near TDC during compression stroke.

However, as tumble flow ratio increases, heat release rate is

enhanced, and the phase is advanced, as the result that the

stronger tumble flow causes beneficial air-fuel mixture

around the spark gap, and better stratified-charge mixture

throughout combustion chamber. In addition, accumulated

heat release also increases, but there are slightly distinct

variations for mean in-cylinder pressure.

Fig.14 shows the effect of different tumble flow ratios on

mean in-cylinder temperature. It shows that mean in-cylinder

temperature dramatically increases as tumble flow is

enhanced. In addition, the inflexion of mean temperature is

advanced as well as it is possibly attributed to tumble flow

improving in-cylinder mixture distribution and enhancing

turbulence flow near top dead center (TDC). Finally, all

accelerate the rate of heat release rate.

Fig. 13. Effect of tumble flow on combustion




Fig.14. Effect of tumble flow on mean in-cylinder

temperature

Fig. 15. Effect of tumble flow ratio on carbon dioxide

Fig. 16. Effect of tumble flow ratio on nitrous oxide

The CO2 and NO emissions are given in Fig.15 and Fig.16

respectively. In Fig.15, it can be found that, increasing

tumble flow ratio, the starting of combustion is slightly

advanced, and the amount of CO2 rises dramatically as well

due mainly to appropriate stratified-charge mixture caused by

tumble flow.

The NO formation depends strongly on in-cylinder

temperature and combustion duration at high temperature

The results shown in Fig.16 keep in accordance with the

principle. The stronger tumble flow promotes the bette

stratified-charge formation, leading to a stable combustion

generation.

The effect of injection timing on

combustion and in-cylinder emissions for

piston bowl B at TR=0.5

Combustion

The start of injection (SOI) is an important parameter because

it affects the combustion characteristics and exhaus

emissions of the engine. To achieve the higher degree o

stratification of the air-fuel mixture, the accurate control o

the quantity and timing of fuel injection is necessary

Therefore, it is important to investigate the effect of injection

timing on combustion and emission characteristics later. Inthe investigation, regarding to the piston B, combustion

characteristics and emissions are given afterwards.




Fig.17. Effect of injection timing on combustion

Fig.18. Effect of injection timing on mean temperature

Fig.17 shows the mean in-cylinder pressure, heat release rate

and accumulated heat release characteristics. As injection

timing is retarded, in-cylinder combustion characteristics

don't vary monotonously, in other words, there exists an

optimization value of injection timing. In the calculatingmodel, the injection timing of 298°CA ATDC is better,

relative to that of 278 °CA ATDC and 310°CA ATDC.

Considering the fact that stratified combustion dramatically

increases the combustion pressure, it can be concluded that

the proper delay of injection timing can lead to the desired

stratification of the mixture.

Fig.18 shows the effect of injection timing on mean in-

cylinder temperature. Considering the fact that stratified

combustion dramatically improves combustion process at part

load, it can be concluded that the delay of injection timing

can lead to the desired stratification of the mixture. Hence, it

can be found that, the appropriate delay of injection timing

causes mean in-cylinder temperature to rise.

Emissions

In this part of the study, the exhaust emissions such as CO2

NO, were compared for the three injection timings at the

different crank angles. The relationships between each

emission and injection timing were separately discussed

Some observations between the emissions were also given in

this paragraph.

Fig.19. Effect of injection timing on carbon dioxide

Fig.20. Effect of injection timing on nitrous oxide

CO2 and NO Emissions are given in Fig.19 and Fig.20

respectively. For the injection timing of 298°CA ATDC, it is

suitable to the piston bowl shape at ignition timing of

345°CA ATDC. Therefore, CO2 emissions generate

relatively high, due to adequate combustion.




In general, at the temperatures over 1800 K, NO generation

rate increases rapidly with increasing combustion

temperature. Therefore, the peak combustion temperature

must be kept below 1800 K to prevent the rapid NO

formation. At the injection timing of 298 °CA ATDC, NO

emissions generate relatively high, as the result of adequate

combustion leading to a higher temperature and combustion

period at high temperature. At the injection timing of 278°CAATDC, because the stratified-charge mixture is formed

unsuitably, especially at the time of ignition, combustion

temperature can't reach the level of lots of NO generated,

resulted the NO emissions at the level of nearly to zero.

In addition, other literatures [19] present some information,

which fuel stratification within the cylinder produces local

equivalence ratio different from the global equivalence ratio.

Therefore, the local fuel enrichment will result in relatively

high local temperatures. Thus, the high local temperatures

provide the energy required to form NO at a rapid rate. Fuel

stratification can be achieved by retarded SOI timing in a

direct injection system.

CONCLUSIONS

The simulating results of this study were summarized as

follows:

The piston bowl shape is a key parameter, which can promote

the stratified-charge formation and combustion process in the

SCDI gasoline engine. In this case, compared to the piston A,

the piston B is more suitable for the stratified combustion at

the injection timing of 310°CA ATDC, subsequently, the

ignition timing of 345°CA ATDC.

In general, tumble flow affects the degree of mixture

stratification, flame kernel formation and flame diffusion.

The stronger tumble flow exists in the combustion chamber,

the better stratified-charge mixture is formed, and however,

tumble flow is excessively strong resulting in blowing out the

flame kernel, misfiring and quenching and so on.

To achieve high degree of stratification of the charge, the

accurate control of injection timing is necessary. Advancing

or retarding injection timing excessively, deteriorate mixture

formation, therefore, there is an optimization value. For the

investigated piston B, the injection timing of 298°CA ATDCis more appropriate at ignition timing of 345°CA ATDC with

the tumble flow ratio of 0.5, compared to that of 278°CA

ATDC and 310°CA ATDC.

ACKNOWLEDGMENT

The authors would like to appreciate the funding of this study

from Geely and Tongji Automotive Research Institute.

CONTACT

Dr Liguang Li, Prof.

School of Automotive Studies

Tongji University, Shanghai, 201804, P.R. China.

[email protected]

Tel: +86 21 69583817

Fax: +86 21 69589978

Qianwang Fan

PhD Candidate

[email protected]

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ABBREVIATION AND DEFINITION

ATDC

After Top Dead Center

BSFC

Brake Specific Fuel Consumption

BC

Boundary Condition

CA

Crank Angle

CFD

Computational Fluid Dynamics

CO2

Carbon dioxide

DISI

Direct Injection Spark Ignition

DI

Direct Injection

ECFM

Extended Coherent Flame Model

FM

FAME Meshing

FEP

FAME Meshing Plus

DISI

Direct Injection Spark Ignition

NO

Nitrogen oxide

PFI

Port Fuel Injection

SCDI

Stratified Charge Direct Injection

SOI

Start of Injection

TR

Tumble flow ratio








Fig.6. Interaction between charge-flow and spray




Fig.7. Mixture strength on different sections through the combustion chamber at various crank angles (the symbol “•” denotes

spark plug position.)




Fig.8. Spatial distribution and temporal history of in-cylinder temperature




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ISSN 0148-7191

doi:10.4271/2010-01-0595

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