International journal of automotive technology, tập 10, số 3, 2009

However, looking at cases with O2-16% and O2-12%, a similar trend can be observed where soot formation or flame luminosity peaks are always lower at higher injection pressures because of

EVALUATION OF LOW-TEMPERATURE DIESEL COMBUSTION REGIMES WITH n-HEPTANE FUEL IN A CONSTANT-VOLUME CHAMBER

U B AZIMOV 1) , K S KIM 1)* , D S JEONG 2) and Y G LEE 2)

1)Department of Mechanical Design Engineering, Chonnam National University, Cheonnam 550-749, Korea2)Korea Institute of Machinery and Materials, Eco-Machinery Research Division, 171 Jang-dong,

Yuseong-gu, Daejeon 305-343, Korea(Received 31 March 2008; Revised 24 December 2008)

ABSTRACT– The concept of Low Temperature Combustion (LTC) has been advancing rapidly because it may reduce emissions of NOx and soot simultaneously Various LTC regimes that yield specific emissions have been investigated by a great number of experiments To accelerate the evaluation of the spray combustion characteristics of LTC, to identify the soot formation threshold in LTC, and to implement the LTC concept in real diesel engines, LTC is modeled and simulated However, since the physics of LTC is rather complex, it has been a challenge to precisely compute LTC regimes by applying the available diesel combustion models and considering all spatial and temporal characteristics as well as local properties of LTC In this paper, LTC regimes in a constant-volume chamber with n-Heptane fuel were simulated using the ECFM3Z model implemented in a commercial STAR-CD code The simulations were performed for different ambient gas O 2 concentrations, ambient gas temperatures and injection pressures The simulation results showed very good agreement with available experimental data, including similar trends in autoignition and flame evolution In the selected range of ambient temperatures and O 2 concentrations, soot and NOx emissions were simultaneously reduced

KEY WORDS : Low-temperature combustion, ECFM3Z model, STAR-CD, Autoignition, Soot, NOx

1 INTRODUCTION

International regulations ratified in recent years have

im-posed more stringent limits on pollutant emissions and fuel

consumption in internal combustion engines To comply

with these regulations and reduce diesel NOx and soot

emissions, new combustion concepts and technologies are

being developed aggressively (Workshop 2006; Kimura et

al., 1999; Kawamoto et al., 2004; Pickett and Siebers,

2004a) As one technology, homogeneous charge

com-pression ignition (HCCI) and conventional diesel-based

Low-Temperature Combustion (LTC) concepts show great

potential in reducing NOx and soot emissions

simultane-ously The LTC concept is a better candidate because it

allows easier auto-ignition control and it can be applied to

conventional diesel engines with minimal design

modifi-cations However, the differences in chemistry and

bustion between this concept and conventional diesel

com-bustion must be investigated to determine their effects on

spray combustion characteristics as well as emissions

(Beatrice et al., 2007)

LTC processes are investigated by computer modeling

and simulation, which provide better understanding of the

combustion process of new combustion concepts Different

low-temperature combustion regimes can be evaluated

rapidly at low cost However, the chemical kinetics ved in this concept do not allow the use of classical dieselauto-ignition and combustion models based on oversimpli-fied representations of combustion chemistry In theseclassical models, the time in the reactive zone is usuallyconsidered much smaller than the diffusion time of fuel andair towards the flame region

invol-This work analyzes the effect of various parameters such

as the ambient gas O2 concentration, ambient gas ature, and fuel injection pressure on the evolution of dieselflames and emission formation in low-temperature com-bustion regimes The results are obtained by simulatingLTC conditions with the n-Heptane fuel and ECMF3Zmodel used by the STAR-CD code First, the features ofvarious combustion models are compared with respect toLTC Then, the simulation results are presented in com-parison with available experimental data

temper-2 ANALYSIS FORMULATION

To mitigate the formation of NOx, diesel combustion mustoccur at low temperatures (Yu and Shahed, 1981), but lowcombustion temperature can generally lead to soot formation.The soot, however, can be avoided by initiating combus-tion at an equivalence ratio below 2 and flame temperatureunder 1800K (Kamimoto and Bae, 1988; Akihama et al.,2001; Kitamura et al., 2003) Diesel diffusion flames can

*Corresponding author. e-mail: sngkim@chonnam.ac.kr

266 U B AZIMOV, K S KIM, D S JEONG and Y G LEE

have complete combustion at temperatures in the range of

1500~1600 K, where NOx formation is very low (Pickett,

2005) Therefore, there is a trend towards the development

of low-temperature combustion strategies for diesel engines

The initial premixed burn of classical diesel combustion is

an example of this type of low-temperature combustion,

and if the mixture is lean enough, soot will not form during

the low-temperature combustion reaction

Low-temperature combustion in diesel engines consists

of fuel injection in which the fuel is allowed to vaporize

and mix with the ambient gas before combustion occurs A

high level of Exhaust Gas Recirculation (EGR) is usually

used to reduce the combustion temperature, and heat

release is controlled by the chemical reaction kinetics of

the mixture (Aceves and Flowers, 2004) This introduces

new variables due to the factors that are not present in

traditional diffusion-burn diesel combustion, where

com-bustion starts in a cetane number-based time delay after the

start of the fuel injection

As shown in Figure 1, after the fuel injection, fuel

evaporation occurs as the hot air is entrained into the fuel

jet and mixes During the fuel evaporation, chemistry

becomes active and entrainment continues until ignition

occurs Once ignition occurs, it is assumed that no more air

mixes into the core of the fuel jet because the oxygen is

consumed in the outer layers of the jet Consequently,

non-sooting and low NOx combustion is realized at equivalence

ratios below 2.0 and flame temperatures less than 1800 K

In LTC, with the increase of EGR, the auto-ignition

delay period is increased and fuel-air premixing is

impro-ved Although liquid fuel penetrates much further into the

chamber, the higher energy released from premixed

reac-tions contributes to the intense evaporation of liquid fuel

(Higgins et al 2000; Idicheria and Pickett, 2005) Since the

fuel and air are very well mixed, the amount of oxygen

around the fuel molecules is sufficient to prevent pyrolysis

and soot formation throughout the jet cross-section

Inaccurate predictions of alternative diesel combustion

regimes often originate from the fact that many numerical

approaches use the Magnussen eddy break-up concept

(Magnussen and Hjertager, 1976), in which the complexity

of the combustion chemical reactions is eliminated with a

fast chemistry limit As the diesel combustion progresses,

there is a full spectrum of important chemical and

turbu-lence time scales ranging between the limits of slow,

distributed chemistry and turbulent, mixing-controlled, fast

chemistry Both mixing and chemical time scales are

crucial to the diesel modeling In LTC, introducing

finite-rate chemistry is important for accufinite-rately predicting

pollu-tant formation To improve the accuracy of predictions

while modeling diesel spray combustion, unified

combus-tion models have been built to account for all types

com-bustion modes simultaneously

Abraham et al (1985) suggested replacing the

controll-ing time scale in the Magnussen model by the slowest time

scale of the mixing time and the chemical time Kong et al

(1995) proposed an extended characteristic-time modelbased on Abraham et al (1985) which accounts for chemi-cal and turbulence time scales simultaneously This modelwas combined with the Shell ignition model to simulate theoverall combustion processes in a diesel engine In thiscombined model, the initiation of combustion relies onlaminar chemistry, and turbulence starts to have an influ-ence on combustion only after combustion events havealready been observed, similar to the Magnussen model.Even if premixed and non-premixed combustions are takeninto account in this model, the non-mixture of the specieswithin a computational cell is only represented by themixing time-scale, which does not account for the mixinghistory Consequently, the transition between chemicallycontrolled and mixing-controlled combustion needs to bemonitored by an empirical function This model does notaccount for flame propagation combustion These modifi-cations improve the eddy break-up model only to a minorextent because only the time scales from the limiting ends

of the diesel combustion time scale spectrum are included

A two-zone flamelet combustion model was developed

by Chen et al (2000) Based on the classical flameletmodel by Peters 1986, this model (in which the reactionsoccur in wrinkled turbulent flames that can be considered

as a collection of laminar flamelets) suggests that each cellFigure 1 Conceptual scheme of low-temperature combustion

is divided by the flame front into two zones: the unburned

zone and the burned zone The unburned zone consists of

air, fuel vapor, and residual gases, and the burned zone

contains combustion products The unburned zone is further

divided into two regions: the segregated region and the

fully mixed region The combustion is decoupled as two

sequential events: mixing and burning However, the use of

flamelet models requires a separation between chemistry

and turbulence time scales, even though the chemical

reactions are still in the fast chemistry limit

Another approach called EPFM (Eulerian Particle

Flame-let Model) was developed by Hasse et al (2000) This

model is an extension of the RIF (Representative

Inter-active Flamelet) concept by Pitsch et al (1995) for a single

representative flamelet The EPFM model can be used to

solve multiple unsteady flamelets in a flamelet code, and

simultaneously solve the Navier-Stokes equations in a CFD

code The CFD code solves the three-dimensional

equa-tions for the flow, turbulence, enthalpy, mixture fraction

and its variance The flamelet parameters are calculated

from the turbulence and mixture field, and are then passed

to the flamelet code The EPFM model assumes the

intro-duction of different marker particles, which are associated

with different flamelet histories depending on the path a

particle takes through the turbulent flow field This model

allows the representation of autoignition and diffusion

flames, but represents the mixing and combustion, which

are very local phenomena, in an averaged way since the

flamelets are based on the averaged properties over all

parts of the domain This model does not account for flame

propagation and computational cost increases with the

number of flamelets involved

The CMC (Conditional Moment Closure) approach,

which was independently developed by Klimenko(1990,

1993), and Bilger (1993), is considered to one of the more

advanced models for turbulent reacting flows In this

approach, the mixture fraction Z is not represented solely

by its mean value and fluctuations like in most models;

instead, the Z-space is discretized, and combustion and

mixing processes are solved for different values of Z The

main concept behind CMC is to find how the reactive

scalars (e.g temperature, species mass fractions) depend on

the mixture fraction CMC can be applied for infinitely fast

and finite rate chemistry The CMC model calculates

conditional moments at a fixed location within the flow

field using modeled transport equations for the conditional

moments of the reactive scalars with no assumptions on the

small-scale structure of the reaction zones or on the relative

timescales of chemistry and turbulence This approach is

very promising but computational cost still remains

unaccep-table for industrial applications

More recently, a new, flame surface density approach

was proposed to model auto-ignition and diffusion flames

It considers the dimension of mixing, represented by the

mean mixture fraction and its fluctuation, and the

dimen-sion of progress of reaction, represented by the mean

pro-gress variable and its fluctuation (Pope, 1988; Candel andPoinsot, 1990; Bray et al., 2005) This approach is based onthe Coherent Flame Model (CFM), which describes therate of fuel consumption per unit volume as the product ofthe flame surface density (i.e the flame surface per unitvolume) and the local flame speed at which it consumes themixture This approach supposes that the chemical reaction

of fuel oxidation occurs in a very thin layer This layerseparates the burned and unburned gases and propagatestoward the fresh mixture of fuel, oxygen and dilutant TheCFM model was extended first to the ECFM model, andwas specifically adapted to model combustion with perfect-

ly or partially mixed mixtures and to simulate the tion processes in direct injection spark ignition engines(Colin et al., 2003) Then ECFM was adapted to accountfor unmixed or diffusion combustion, and the three-zonedescription of the mixing state was added This new com-bustion model, called ECFM3Z (3-Zones Extended CoherentFlame Model), can therefore be seen as a simplified CMC-type model, which discretizes the mixture fraction space byonly three points Therefore, this model was selected to bethe most appropriate for simulating LTC regimes because itcan reflect the real physics of LTC, it relies on flameletlibraries, and it is less computationally demanding than theCMC and EPFM models

combus-3 ECFM3Z MODEL CONCEPT The ECFM3Z model was briefly presented by Beard et al.(2003) and was described in detail by Colin and Benkenida(2004) This model describes the unburned/burned gaszones based on the flame surface density equation In order

to account for diffusion flames and mixing processes, eachcomputational cell is split into three mixing zones: a pure

or unmixed fuel zone, a pure air plus EGR zone or unmixedair and EGR zone, and a mixed zone, containing fuel, airand EGR This structure can account for the three maincombustion modes: auto-ignition, flame propagation anddiffusion flame as encountered in LTC It is based on twodimensions: the mixing state description and the reactionprogress description The mixing state description is repre-sented by the Probability Density Function (PDF) of themixture fraction

(1)where is the average value of the mixture fraction inthe mixed zone The first δ function corresponds to theunmixed air region, the second one to the mixed region andthe third one to the unmixed fuel region In this structure,space Z is discretized by only three points The mixingmodel can reflect the transference of unmixed fuel andunmixed air into the mixed region The reaction progressdescription is represented by the progress variable

-268 U B AZIMOV, K S KIM, D S JEONG and Y G LEE

where is the mass fraction of the fuel present in the

unburned gases, and is the mass fraction of fuel before

the onset of combustion (fuel tracer) is constant in

space and time for perfectly mixed charges In practical

applications varies in space and time because of the

imperfect mixing of the charge In addition, a transport

equation is solved to obtain the Favre average mass

den-sities of the chemical species of the fuel, O2, N2, NO, CO2,

CO, H2, H2O, O, H, N, OH and of the soot inside the

computational cell containing the three mixing zones A

detailed description and specific features of the ECFM3Z

model are given in Colin and Benkenida, 2004; Colin et al.,

2005; Reveille et al., 2006; Knop and Jay, 2006; Priesching

et al., 2007; Shi et al., 2007

4 SIMULATION PROCEDURE

The simulation was conducted using the STAR-CD

com-mercial CFD code in a three-dimensional computation grid

The ECFM3Z model with appropriate adjustments was

incorporated into STAR-CD The computational grid assumes

a cylinder-shaped constant volume chamber of 80 mm in

diameter and 80 mm in length The discretization of space

(number of cells) and time (time steps) are set after the

Courant number (STAR-CD Methodology, 2006) In

addi-tion, the complete spray combustion duration was adjusted

to match that of the experiment The mesh resolution was

set to achieve good agreement between the simulation

results and experimental results for the penetration of

non-reacting and non-reacting fuel jets

The fuel was injected with spray characteristics adjusted

according to the spray characteristics assumed in the

ex-periments In the spray model, atomization proceeded

according to the Reitz-Diwakar model and the fuel droplets

were formed according to the Reitz-Diwakar breakup

model This atomization model assumed that the liquid

emerges from the nozzle as a jet, waves form on the jet’s

surface, and then the waves are amplified and the liquid is

eventually broken up into droplets by aerodynamic forces

caused by the high relative velocity between the liquid and

the gas (Reitz, 1987) To apply this model, a semi-cone

angle must be known and given as part of the input data

Based on this angle, the initial droplet velocity is

determin-ed This angle was determined from experiments

perform-ed using the same common-rail spray characteristics and

ambient gas conditions as those mentioned in this paper

(Jeong, 2003)

The autoignition in the present simulation was controlled

by the double-delay autoignition model This autoignition

model was developed to consider the effect of cool flames,

which are characterized by a weak increase in temperature

after an initial delay, followed by a slowing of the reaction

rates until the second delay After this second delay, the

reaction rate increases rapidly, and the main autoignition

takes over This model makes use of pre-computed tables

containing the results of complex chemistry calculations of

the autoignition of n-heptane (Curran et al., 1998; Subramanian,2007) The tables give values for the two delays and thesedelays are functions of pressure, temperature, equivalenceratio and EGR For emission simulation, the 3-step Zeldovichmodel and ERC model were used for NOx and soot emissioncalculations, respectively The simulation conditions arelisted in Tables 1 and 2 The ambient gas temperature,ambient gas content, ambient gas pressure, fuel injectionpressure, injection duration and single-hole injector orificeparameters correspond to those of the experiment

5 RESULTS AND DISCUSSION5.1 Evaluation Approach for LTC Regimes The present paper numerically evaluates the LTC regime of

DI diesel combustion Since LTC differs from conventionaldiesel combustion, it is necessary to use a model that isuniversally applicable to both conventional and alternativediesel combustion applications Soot and NOx emissionscomputed for particular conditions, and the entire combus-tion event was evaluated to understand the physics of thecombustion as well as the relations among the operatingparameters For this purpose, the parameter called “Com-bustion Factor” was introduced:

(3)where, is the mass fraction of fuel in the burnt gases,and is the fuel tracer

This parameter is considered as an indicator of thecombustion mode (premixed vs diffusion) in the complexLTC process It is extremely difficult to differentiate thediesel LTC process into certain modes because premixed,partially premixed and diffusion modes occur simultane-ously However, it might be possible to map the com-bustion event and see which combustion mode prevails and

how it would change with the change of a certain operating

parameter The combustion factor varies from 0 to 1 If this

parameter approaches 1, the combustion is considered

premixed, and if it approaches 0, the combustion is

con-sidered as diffusion flame mode To validate this approach,

the original ECFM3Z model was invoked (Colin and

Benkenida, 2004) As mentioned earlier, the ECFM3Z

model consists of 3 zones that allow computation of all 3

combustion modes, autoignition, premixed flame and

diffusion flame In the ECFM3Z model, the transport

equa-tions are solved to obtain the Favre average mass densities

of chemical species as well as the fuel In the mixed zone,

the fuel is divided into two parts: the fuel present in the

fresh gases and the fuel present in the burned gases This

division is necessary because the fuel in the fresh gases will

be consumed by autoignition and the premixed flame,

while the fuel in the burned zone will be consumed by the

diffusion flame During the combustion event, if there is

any fuel in the burned gases, this fuel will be consumed and

post-oxidized by the source of the Magnussen EBU model

or diffusion combustion model Therefore, it would be

possible to evaluate the extent of diffusion combustion within

the entire combustion event and to estimate the magnitude

of soot formation, accepting the fact that soot formation

can be avoided during the premixed mode To evaluate the

entire range of the combustion regimes, the combustion

factor was normalized using the expression below to obtain

results for all the conditions under the same scale:

(4)5.2 Auto-ignition and Flame Development

As mentioned earlier, the double-delay autoignition model

was used within the ECFM3Z model to simulate the

autoignitions of various LTC regimes Figure 2 shows the

autoignition delay mapping for all conditions mentioned in

this paper This figure shows that the first autoignition

delay moderately changes from that of the main

autoigni-tion Also, it is seen that with the decrease of O2

concent-ration or the increase of equivalence ratio and decrease of

the ambient gas temperature, the autoignition delay slightly

increases According to the low-temperature reaction

mechanism, the first autoignition of hydrocarbon fuel is

largely associated with the decomposition of the

keto-hydroperoxide species at temperatures between 800 and

850 K, and the end of the first autoignition occurs when the

temperature reaches NTC zones The start of the first

autoignition is determined by the time needed for the air/

fuel mixture to reach the decomposition temperature The

figure shows that the first autoignition delay periods at

920 K and 870 K are almost similar, because the ambient

gas temperatures are high enough to immediately initiate

the decomposition of the mixture However, the first

auto-ignition delay at 820 K is noticeably different, probably

due to longer time required for the air/fuel mixture to reach

the decomposition temperature In addition, probably at

820 K, there was no significant decomposition of fuel ing evaporation, but some portion of fuel did decomposeduring evaporation for the high temperature case Thisassumption is in agreement with other presented work(Curran et al., 1998; Wang and Rutland, 2005)

dur-After the first autoignition, there is a period of veryslight temperature increase due to “cool flame” chemistry.This period is an ignition delay between the initial fueldecomposition and very rapid temperature rise The timeinterval between the first autoignition and the main auto-ignition is much greater at 820 K for similar O2 conditionscompared to those of the other two cases This difference isdue to both the retarded first autoignition and the furtherretarded reaction progress with the decrease of O2 concent-ration In the ECFM3Z model, defining the occurrences ofthe first and main autoignitions is straightforward becausetheir computed values are automatically stored in the post-processing file Figure 3 compares spray combustion simu-lation results for various O2 concentrations and ambient gastemperatures with experimental data

The results presented in Figure 3 indicate good ment with data from the experiments in terms of spray andcombustion development The liquid fuel pattern, as well

agree-as the spatial distribution of the flame, matches well thepattern obtained from experiments

Figure 3 does not provide any information on the start ofautoignition, but only depicts the comparison in sprayflame development between simulation and experiment.Nevertheless, based on Figures 2 and 3, the effect of vari-ation of charge composition showing the longer autoigni-tion delay periods and decreased flame temperatures wasevident for the diluted charge at lower O2 concentration.5.3 Flame Temperature

The ambient gas temperature has a small influence onflame temperature and NOx formation but has great effect

on the fuel/air equivalence ratio The ambient gas

270 U B AZIMOV, K S KIM, D S JEONG and Y G LEE

ature can only have an effect on autoignition initiation

because of the reaction of fuel with oxygen in the

high-temperature environment And then, as the flame

propa-gates, the combustion is controlled by the O2 concentration

and injection pressure Figure 4(a), (b), (c) shows a similar

pattern in the temperature evolution for various

com-bustion regimes with a change in ambient gas temperature

Although the autoignition delay becomes longer with

decreasing ambient gas temperature, the maximum values

of the flame temperature at conditions with similar O2

content are nearly the same Note that the flame

temper-ature is slightly increased by an increase in injection

pressure, especially during the initial stage of flame

propa-gation and temperature increase This may be explained by

better evaporation and mixing at higher injection pressures

(Gill et al., 2005) As the fuel is well-mixed and distributed,

higher energy is released, and flame temperature is increased

5.4 NOx

The three-step Zeldovich mechanism was used to compute

the NOx emissions, which are generally believed to depend

on only the flame temperature Figure 5(a), (b), (c) shows

that with the decrease of O2 concentration, NOx emissions

are gradually reduced, especially for the conditions of O2

-12%, which corresponds to about EGR-60%, where NOx is

reduced to almost zero These results are in agreement with

experimental and numerical data, indicating that NOx

formation can be avoided at reduced flame temperatures

with decreased oxygen concentration in the ambient gas

(Heywood, 1989; Abd-Alla, 2002; Egnell, 2000; Wagner et

al., 2003; Alriksson and Denbratt, 2006) Also, the effect of

higher injection pressure on NOx formation, as on the flametemperature discussed previously, is indicated A similartrend of increased NOx formation with higher injectionpressures was shown by Henein et al., 2006

5.5 Equivalence RatioAkihama et al (2001) have shown that in addition to thenotable decrease in NOx emissions, soot formation can also

be avoided by producing combustion at flame temperaturesless than 1800 K Kamimoto and Bae (1988) proposed thatsoot formation could be avoided by producing combustion

at equivalence ratios below 2 Non-sooting combustion hasalso been demonstrated at higher temperatures and higherequivalence ratios by entraining sufficient oxygen into thejet NOx formation occurs at high temperatures, but NOx isreduced to N2 under fuel-rich conditions and thus, NOxemissions are decreased with higher equivalence ratios

In the case of low-temperature combustion, soot emissionreduction appears to be related to an increase in ignitiondelay, which is due to the reduced O2 concentration, whichprovides more time for mixing before combustion and apossible decrease in the equivalence ratio of the ignitingfuel-ambient gas mixture

Figure 6 shows equivalence ratio distribution at the time

of main ignition for each combustion regime The deeperthe fuel penetrates into the chamber, the more diluted andmixed it becomes as O2 concentration and ambient temper-ature decrease The equivalence ratio at the jet's leadingpart, the zone where ignition is supposed to occur and wherethe premixed burn occurs, is about 2

The maximum equivalence ratio value corresponds toFigure 3 Spray combustion development at Pinj-90 MPa

the core of the injected fuel and is about 6, and the

stoichio-metric values are along the jet periphery, which is in

agree-ment with the experiagree-mental and numerical results obtained

by other researchers for similar conditions (Idicheria and

Pickett, 2007)

5.6 Soot

The computed soot data were compared with experimental

data of flame luminosity This direct comparison may not

provide quantitative information about soot formation

be-cause soot is determined by directly solving the transport

equations, but flame luminosity is experimentally related to

soot concentration and local flame temperature

It is accepted that the flame luminosity can be

interpret-ed as a qualitative indicator of in-cylinder soot formation

(Siebers et al., 2002; Mueller and Martin, 2002; Choi et al.,

2004; Kim et al., 2007)

In Figure 7(a) and (b), both the soot curves and the flameluminosity curves have a similar pattern After sootformation and flame luminosity reaches their peaks, sootoxidation will dominate and flame luminosity will decrease.The soot oxidation process is clearly seen to be slower thanthe soot formation process in both (a) and (b) A similartrend was observed for other cases for different O2 concent-rations and ambient gas temperatures, as in Figures 8(a),(b) and 9(a), (b) In general, higher flame luminositypeaks were found for lower injection pressures Higherluminosity peaks with higher injection pressure for the case

of O2-21% and ambient temperature 920 K in Figure 7 areprobably due to a higher local equivalence ratio at higherambient gas temperatures

Although a higher injection pressure is believed to tribute to better mixing (Pickett and Siebers, 2004b), there

con-is probably still not enough time to ensure sufficient

Figure 4 Flame temperature variation with the change of

O2 concentration, ambient gas temperature and injection

pressure

Figure 5 NOx variation with the change of O2 ration, ambient gas temperature and injection pressure

concent-272 U B AZIMOV, K S KIM, D S JEONG and Y G LEE

mixing of the charge to decrease local equivalence ratio

and prevent soot formation However, looking at cases with

O2-16% and O2-12%, a similar trend can be observed

where soot formation or flame luminosity peaks are always

lower at higher injection pressures because of a longer

ignition delay

5.7 Combustion Factor

Based on the results mentioned above and applying the

combustion factor concept described earlier, the LTCregimes are evaluated to determine the best scenario ofcombustion in terms of soot and NOx emissions Note that

Figure 6 Equivalence ratio at the time of ignition at Pinj-90

MPa

Figure 7 Comparison of calculated soot and measured

flame luminosity at Tamb-920 K

Figure 8 Comparison of calculated soot and measuredflame luminosity at Tamb-870 K

Figure 9 Comparison of calculated soot and measuredflame luminosity at Tamb-820 K

the resulting curves of the combustion factor are a relative

representation of one regime against the other within the

framework of the conditions considered in this paper,

solely mapping from the worst-case and the best-case

scenario

According to a number of experimental and numerical

research results, simultaneous soot and NOx reduction can

be achieved in premixed combustion with EGR and

fuel-lean mixtures (Kamimoto and Bae, 1988; Dec, 1997;

Kimura et al., 1999; Akihama et al., 2001) Figures 10, 11

and 12 show the results for combustion progress At the

instant of ignition, combustion starts as premixed mode

To confirm the assumptions stated above regarding the

combustion factor, the quantity of fuel present in theburned gases and the values of the equivalence ratio at theinstances when the soot formation is at its maximum weredetermined For brevity, only the case with ambient temper-ature of 870 K is presented in Figure 13 As mentionedearlier, the fuel in the burned gases is consumed by thediffusion flame This fact can be utilized to determine the

Figure 10 Combustion factor at Tamb-920 K

Figure 11 Combustion factor at Tamb-870 K

Figure 12 Combustion factor at Tamb-820 K

Figure 13 Computed equivalence ratio and the fraction offuel in burned gases at Tamb-870 K

Figure 14 Pressure rise comparison for different lation conditions

Figure 15 Apparent heat release rate for different lation conditions

simu-274 U B AZIMOV, K S KIM, D S JEONG and Y G LEE

extent of the diffusion mode This figure shows that the

mass of fuel contained in the burned gas region decreases

with decreasing O2 concentration

Figure 11 shows that the diffusion mode lessens with

decreasing O2, and the entire combustion event for each

condition tends to move towards premixed mode On the

other hand, the equivalence ratio increases with decreasing

O2, because of the richer mixture

Figures 14 and 15 show the pressure rise and heat

release rate plotted as a function of time for different O2

concentrations and ambient gas temperatures These plots

show the increase in ignition delay with decreasing O2

concentration For example, the ignition delay for O2-12%,

Tamb-820 K is 1.659 ms whereas that for O2-21%, Tamb-920

K is 0.819 ms In-chamber pressure decreases with

decreasing O2 concentration, possibly because of slower

chemical kinetics due to the lack of oxygen However, the

time available for premixing is longer because of the

increased ignition delay These results are in agreement

with the previous work of other researchers (Chen et al.,

2003; Kook et al., 2005)

An interesting trend was observed in the effect of

ambient gas temperature on the heat release rate As shown

in Figure 15, heat release greatly increases when the delay

for the latter case is longer, it most likely causes improved

mixing, resulting in a higher peak of heat release In addition,

if two conditions are compared, one with O2-12% and

injection pressure 90 MPa, and the other with O2-12% and

injection pressure 135 MPa, it is seen that more energy is

released in the latter case because of the improved mixing

induced by higher injection pressures These results are in

agreement with previous research (Sugiyama et al., 1994)

The results above can be summarized to support the

argument that the diffusion mode can have a great

influ-ence on soot and NOx formation during diesel combustion

Furthermore, the combustion factor can serve as an

indicator of the quality of the diesel combustion process in

terms of soot and NOx formation Soot and equivalence

ratio decrease with increasing injection pressure This

assumption supports the fact that mixing plays an

impor-tant role in soot formation However, with the decrease of

O2 concentration, soot gradually decreases but equivalence

ratio slightly increases The equivalence ratio increases due

to the lack of oxygen entrainment as O2 decreases, and soot

decreases most likely due to the lower flame temperature

with reduced oxygen content Akihama et al (2001),

show-ed that soot formation can be suppressshow-ed at temperatures

below 1700 K Theoretical analysis of reaction rates

per-formed by Jacobs and Assanis (2007), suggested that soot

formation is insensitive to equivalence ratio at

temper-atures below 1500 K At such low tempertemper-atures, the reactions

forming soot particles from PAH (Polycyclic Aromatic

Hydrocarbons) do not progress, even if rich combustion

occurs Therefore, the ECFM3Z model seems to have good

predictive capabilities for evaluating various diesel LTC

regimes Note that the distinct separation of the different

ignition/combustion modes makes the ECFM3Z modeluniversally and specifically applicable for accurately simulat-ing conventional as well as alternative diesel combustionregimes

6 CONCLUSION

In this paper, a description of a new approach to evaluateLTC regimes using the ECFM3Z model was presented.The computed results, which were compared with availableexperimental data, showed that this model was able toaccurately predict autoignition and combustion for bothconventional and LTC regimes The specific major findingsare summarized as follows:

(1) The double-delay autoignition model, as a part of theECFM3Z model, was shown to be a good predictor ofautoignition delay for the entire range of conditions inthis paper Computed results were compared with ex-perimental data and good agreement was observed.(2) Spray combustion evolution at various levels of O2

concentration and ambient temperature matched verywell with the experimental data, for spatial as well astemporal jet flame development

(3) Ambient gas temperature and fuel injection pressurehad a minor effect on flame temperature Flame temper-ature values for different conditions were almost thesame when ambient gas temperature increased by 50K.Flame temperature increased a little at higher injectionpressures However, oxygen concentration had a greateffect and flame temperature considerably decreasedwith decreasing O2 concentration in the ambient gas.(4) Consequently, NOx level greatly decreased, for it de-pended only on flame temperature For all the condi-tions with 12% oxygen concentration, NOx level de-creased to almost zero because the flame temperaturewas about 1900K With higher injection pressure, NOxlevel slightly increased because of the slight increase oftemperature

(5) Equivalence ratio at the time of ignition stabilized ataround 2, as the fuel jet penetrated further into thecombustion chamber However, for the conditions O2-21%, 920 K and O2-21%, the 870 K equivalence ratiowas higher because at normal oxygen concentrationsand higher ambient gas temperatures, the ignition delayperiod is very short and there is not sufficient time forthe injected fuel to mix with the ambient gas and todilute before the start of combustion

(6) Computed soot data were compared with experimentaldata of flame luminosity Similar patterns and trends insoot formation were observed for all conditions Sootlevel decreased with decreasing O2 concentration andincreasing fuel injection pressure However, at 21%oxygen and 920K ambient gas temperature, the sootlevel was higher at higher fuel injection pressures Thismay be explained by the insufficient fuel-ambient gasmixing time, as well as the higher flame temperature

(7) Finally, the parameter called combustion factor was

computed and analyzed to evaluate the progress of

combustion as well as the development of premixed

and diffusion modes It showed reasonable correlation

with NOx and soot formation data The fuel mass in the

burned gases decreased with decreasing O2

concent-ration, and therefore, the combustion factor leaned

towards the premixed mode This means that with the

decrease of O2 concentration, mixing of fuel with

am-bient gas improved and the fuel was mainly consumed

in the premixed zone Furthermore, the average

equi-valence ratio at the time steps with the highest soot

formation level increased with decreasing O2

concent-ration Heat release peaks were higher for the cases

with lower O2 concentrations and ambient gas

temper-atures, indicating that more fuel was consumed during

the premixed mode Therefore, the combustion factor,

together with NOx and soot data, can be used with the

ECFM3Z model to predict a trend in both conventional

and alternative combustion regimes to determine the

NOx-soot trade-offs Further research is required to

investigate a broader range of combustion regimes,

taking into account different types of fuel, the real

geo-metry of a diesel combustion chamber, high ambient

gas temperatures and injection pressures, various

am-bient gas densities and EGR conditions

ACKNOWLEDGEMENT− This work was a part of the project

“Development of Partial Zero Emission Technology for Future

Vehicle” funded by Korean Ministry of Commerce, Industry and

Energy The authors would like to gratefully acknowledge its

financial support

REFERENCES

Abd-Alla, G H (2002) Using exhaust gas recirculation in

internal combustion engines: a review Energy Conversion

and Management, 43, 1027−1042

Abraham, J., Bracco, F V and Reitz, R D (1985)

Com-parison of computed and measured premixed charge

engine combustion Combustion and Flame, 60, 309

Aceves, S M and Flowers, D L (2004) A detailed chemical

kinetic analysis of low temperature non-sooting diesel

combustion SAE Paper No 2005-01-0923

Akihama, K., Takatori, Y., Inagaki, K., Sasaki, S and Dean,

A M (2001) Mechanism of the smokeless rich diesel

combustion by reducing temperature SAE Paper No.

2001-01-0655

Alriksson, M and Denbratt, I (2006) Low temperature

combustion in a heavy duty diesel engine using high

levels of EGR SAE Paper No 2006-01-0075

Beard, P., Colin, O and Miche, M (2003) Improved

model-ing of DI diesel engines usmodel-ing sub-grid descriptions of

spray and combustion SAE Paper No 2003-01-0008

Beatrice, C., Avolio, G., Bertoli, C., Del Giacomo, N.,

Guido, C and Migliaccio, M (2007) Critical aspects on

the control in the low temperature combustion systemsfor high performance DI diesel engines Oil & Gas Sci Technology 62, 4, 471−482

Bilger, R W (1993) Conditional moment closure for bulent reacting flow Phys Fluids A5, 2, 436

tur-Bray, K., Domingo, P and Vervisch, L (2005) Role of theprogress variable in models for partially premixed turbu-lent combustion Combustion and Flame, 141, 431−437.Candel, S M and Poinsot, T J (1990) Flame stretch andthe balance equation for the flame area Combustion Sci and Technology, 70, 1−15

Chen, C., Bardsley, M E A and Johns, R J R (2000).Two-zone flamelet combustion model SAE Paper No.2000-01-2810

Chen, R., Milovanovic, N., Turner, J and Blundell, D (2003).The thermal effect of internal exhaust gas recirculation

on controlled auto ignition SAE Paper No 2003-01-0751.Choi, D., Miles, P., Yun, H and Reitz, R D (2004) Aparametric studies of low-temperature, late-injectioncombustion in an HSDI diesel engine Int Symp Di- agnostics and Modeling of Combustion in Internal Combustion Engines, COMODIA

Colin, O and Benkenida, A (2004) The 3-zones extendedcoherent flame model (ECFM3Z) for computing premiz-ed/diffusion combustion Oil & Gas Sci Technology 59,

6, 593−609

Colin, O., Pires da Cruz, A and Jay, S (2005) Detailedchemistry-based auto-ignition model including low temper-ature phenomena applied to 3-D engine calculations.Proc Combustion Institute, 30, 2649−2656

Colin, O., Benkenida, A and Angelberger, C (2003) 3Dmodeling of mixing, ignition and combustion phenomena

in highly stratified gasoline engines Oil & Gas Sci Technology 58, 1, 47−62

Curran, H J., Gaffuri, P., Pitz, W J and Westbrook, C K.(1998) A comprehensive modeling study of n-heptaneoxidation Combustion and Flame, 114, 149−177.Dec, J E (1997) A conceptual model of DI diesel com-bustion based on laser-sheet imaging SAE Paper No.970873

Egnell, R (2000) The influence of EGR on heat releaserate and NO formation in DI diesel engine SAE Paper

No 2000-01-1807

Gil, K., Marriner, C., Sison, K and Zhao, H (2005) cylinder studies of multiple diesel fuel injection in a singlecylinder optical engine SAE Paper No 2005-01-0915.Hasse, C., Bikas, G and Peters, N (2000) Modeling DI-diesel combustion using the eulerian particle flameletmodel (EPFM) SAE Paper No 2000-01-2934

In-Henein, N A., Bhattacharyya, A., Schipper, J., Kastury, A.and Bryzik, W (2006) Effect of injection pressure andswirl motion on diesel engine-out emissions in conv-entional and advanced combustion regimes SAE Paper

276 U B AZIMOV, K S KIM, D S JEONG and Y G LEE

ignition and premixed-burn behavior SAE Paper No.

2000-01-0940

Idicheria, C A and Pickett, L M (2005) Soot formation

in diesel combustion under high-EGR conditions SAE

Paper No 2005-01-3834

Idicheria, C A and Pickett, L M (2007) Effect of EGR

on diesel premixed-burn equivalence ratio Proc

Com-bustion Institute, 31, 2931- 2938

Ito, T., Hosaka, T., Senda, J and Fujimoto, H (2004)

Numerical investigation of soot formation in diesel jet

flame using detailed kinetic model Int Symp Diagnostics

and Modeling of Combustion in Internal Combustion

Engines, COMODIA

Jacobs, T J and Assanis, D N (2007) The attainment of

premixed compression ignition low-temperature

com-bustion in a compression ignition direct injection engine

Proc Combustion Institute, 31, 2913−2920

Jeong, B C (2003) Study on the Spray Characteristics of

Common-rail Injection System M S Thesis Yosu National

University South Korea 25−27

Kamimoto, T and Bae, M (1988) High combustion

temperature for the reduction of particulate in diesel

engines SAE Paper No 880423

Kawamoto, K., Araki, T., Shinzawa, M., Kimura, S., Koide

S and Shibuya, M (2004) Combination of combustion

concept and fuel propertiy for ultra-clean DI diesel SAE

Paper No 2004-01-1868

Kim, K S., Beschieru, V., Jeong, D S and Lee, Y (2007)

Experimental investigation and comparison of spray and

combustion characteristics of GTL and diesel fuels Int.

J Automotive Technology 8, 3, 275−281

Kimura, S., Aoki, O., Ogawa, H., Muranaka, S and Enomoto,

Y (1999) New combustion concept for ultra-clean and

high-efficiency small DI diesel engines SAE Paper No.

1999-01-3681

Klimenko, A (1990) Multicomponent diffusion of various

scalars in turbulent flow Fluid Dynamics, 25, 327−333

Klimenko, A (1993) Conditional moment closure and

fluctuation of scalar dissipation Fluid Dynamics, 28,

630−637

Knop, V and Jay, S (2006) Latest developments in gasoline

auto-ignition modeling applied to an optical CAI engine

Oil & Gas Sci Technology 61, 1, 121−137

Kong, S C., Han, Z and Reitz, R D (1995) The

develop-ment and application of a diesel ignition and combustion

model for multidimensional engine simulation SAE Paper

No 950278

Kook, S., Bae, C., Miles, P C., Choi, D and Pickett, L M

(2005) The influence of charge dilution and injection

timing on low-temperature diesel combustion and

emissions SAE Paper No 2005-01-3837

Magnussen, B F and Hjertager, B H (1976) On

mathe-matical modeling of turbulent combustion with special

emphasis on soot formation and combustion 16th Symp.

Combustion, The Combustion Institute, 719−729

Mueller, C J and Martin, G C (2002) Effects of

oxy-genated compounds on combustion and soot evolution in

a DI diesel engine: Broadband natural luminosity ing SAE Paper No 2002-01-1631

imag-Peters, N (1986) Laminar flamelet concepts in turbulentcombustion 21st Symp Combustion The Combustion Institute

Pickett, L M (2004) Low flame temperature limits formixing-controlled diesel combustion Proc Combustion Institute, 30, 2727−2735

Pickett, L M and Siebers, D (2004a) Non-sooting, lowflame temperature mixing-controlled DI diesel combustion.SAE Paper No 2004-01-1399

Pickett, L M and Siebers, D (2004b) Soot in diesel fueljets: Effect of ambient temperature, ambient density, andinjection pressure Combustion and Flame,138, 114−135.Pitsch, H., Wan, Y P and Peters, N (1995) Numericalinvestigation of soot formation and oxidation underdiesel engine conditions SAE Paper No 952357.Pope, S B (1988) The evolution of surfaces in turbulence.Int J Eng Sci 26, 5, 445−469

Priesching, P., Ramusch, G., Ruetz, J and Tatschl, R.(2007) 3D-CFD modeling of conventional and alter-native diesel combustion and pollutant formation-Avalidation study. SAE Paper No 2007-01-1907 Reitz, R D (1987) Modeling atomization processes inhigh-pressure vaporizing sprays Atomization and Spray Technology, 3, 309−337

Reveille, B., Kleemann, A and Jay, S (2006) Towardseven cleaner diesel engines: Contribution of 3D CFDtools Oil & Gas Sci Technology 61, 1, 43−56

Shi, X., Li, G and Zhou, L (2007) DI diesel enginecombustion modeling based on ECFM-3Z model SAE Paper No 2007-01-4138

STAR-CD 3.26 Methodology (2006) CD-Adapco Inc.Subramanian, G (2007) Modeling engine turbulent auto-ignition using tabulated detailed chemistry SAE Paper

No 2007-01-0150

Sugiyama, G., Ryu, H and Kobayashi, S (1994) tational simulation of diesel combustion with high pre-ssure fuel injection Int Symp Diagnostics and Modeling

Compu-of Combustion in Internal Combustion Engines, COMODIA.Wagner, R M., Green, J B., Dam, T Q., Edwards, K D.and Storey, J M (2003) Simultaneous low engine-outNOx and particulate matter with highly diluted dieselcombustion SAE Paper No 2003-01-0262

Wang, Y and Rutland, C J (2005) Effect of temperatureand equivalence ratio on the ignition of n-heptane fuelspray in turbulent flow Proc Combustion Institute,30,

893−900

Workshop (2006) Basic Research Needs for Clean and Efficient Combustion of 21st Century Transportation Fuels DOE Report USA

Yu, R C and Shahed, S M (1981) Effects of injectiontiming and exhaust gas recirculation on emissions from a

DI diesel engine SAE Paper No 811234

EFFECTS OF INTAKE FLOW ON THE SPRAY STRUCTURE

OF A MULTI-HOLE INJECTOR IN A DISI ENGINE

S KIM 1)* , J M NOURI 2) , Y YAN 2) and C ARCOUMANIS 2)

1)Department of Automotive Mechanical Engineering, Silla University, Busan 617-736, Korea

2)School of Engineering and Mathematical Sciences, The City University, Northampton Square, London, EC1V 0HB, UK

(Received 13 June 2008; Revised 20 December 2008)

ABSTRACT− The spray characteristics of a 6-hole injector were examined in a single cylinder optical direct injection spark ignition engine The effects of injection timing, in-cylinder charge motion, fuel injection pressure, and coolant temperature were investigated using the 2-dimensional Mie scattering technique It was confirmed that the in-cylinder charge motion played a major role in the fuel spray distribution during the induction stroke while injection timing had to be carefully considered at high injection pressures during the compression stroke to prevent spray impingement on the piston.

KEY WORDS : Mie scattering, Intake swirl, Spray structure, Multi-hole injector, Direct injection, Gasoline engine

1 INTRODUCTION

Direct-injection spark ignition (DISI) engines offer the best

promise for simultaneous reduction of fuel consumption

and exhaust emissions in gasoline engines Several DISI

engine models have emerged into the international market

and have highlighted the potential benefits on both fuel

economy and pollutant reduction Most of these engines

are based on the wall-guided combustion design concept

(Wirth et al., 1998; Nouri and Whitelaw, 2002) These

“first generation” injection systems, with swirl pressure

atomizers have been shown lower fuel consumption by up

to 20% in the case of stratified, overall-lean part-load

operation, but showed no significant improvements in HC

and NOx emissions (Fraidl et al., 1996) The key success in

DISI engines is in preparing the right amount of stratified

fuel mixture under part-load operation when the fuel is

injected late in the compression stroke; the goal is to

quickly transport the fuel/air mixture towards the spark

plug with no impingement on surfaces and to achieve

complete evaporation of the droplets in the short time

available between the end of injection and start of ignition

Most recent studies have focused on an alternative strategy

to the wall- and air-guided mode of mixture preparation for

producing stratified fuel mixture preparation for producing

stratified fuel mixtures, the so-called spray-guided using a

new generation fuel injection system with either central or

side fuel injection (Wirth et al., 2004; Shim et al., 2008)

The major advantage of this configuration is that it makes

use of the injection process to ensure that a stable

combustible mixture reaches the spark plug at the time of

ignition which, in turn, depends strongly on the spraycharacteristics and, in particular, its cycle-to-cycle stabilitywhich otherwise may even cause a misfire Thus, to utilizethe full benefit of DISI technology, knowledge of thetemporal evolution of the spray structure, its tip penetrationand distribution of the droplet velocities and diameters as afunction of nozzle design, and injection and chamberpressures is a prerequisite It should be mentioned that theswirl pressure atomiser (first generation) was found to beunsuitable for the new concept of mixture preparation due

to demonstrated spray cone angle instability with backpressure, leading to a complete collapse of the spraystructure when fuel was injected during the compressionstroke (Li et al., 2004; Nouri and Whitelaw, 2006).Recently, a number of injector manufacturers havedesigned new high-pressure multi-hole injectors and out-wards opening piezo injectors, referred to as ‘second-generation’ systems, based on the expectation that theyproduce stable fuel sprays with fine fuel droplets indepen-dent of the time of fuel injection (Wirth et al., 2004) Multi-hole injectors have been studied because of their potentialfor achieving good fuel stratification, thus extending thelean limit further (Preussner et al., 1998) They also offerthe highest possible flexibility in adapting the spray patternlayout to a particular combustion chamber design Theinvestigations of (Ortmann et al., 2001; Lippert et al.,2004; Mitroglou et al., 2006, 2007) on multi-hole injectorsfor gasoline engines confirmed the improved stability ofthe spray at elevated chamber pressures relative to that ofswirl injectors Also, enhanced air entrainment has beenobserved as a result of an enlarged surface area produced

by separated spray jets, enhanced flexibility to direct thesprays towards the proximity of the spark plug and improved

*Corresponding author. e-mail: sskim@silla.ac.kr

278 S KIM, J M NOURI, Y YAN and C ARCOUMANIS

matching with the injector, generated spray and combustion

chamber design Recently, a series of detailed experimental

investigations in a high pressure chamber and a DISI

engine have been carried out and reported in (Mitroglou et

al., 2005, 2006, 2007) regarding gasoline spray

characteri-stics and mixture distribution The constant high pressure

chamber was equipped with a high-pressure multi-hole

injector at injection and chamber pressures up to 20 MPa

and 1.2 MPa, respectively The test results in the constant

chamber confirmed that the overall spray angle relative to

the axis of the injector was independent of injection and

chamber pressure The effects of injection and chamber

pressure on droplet velocities and diameter were also

quantified From the experimental results in the engine, for

late fuel injection during the compression stroke, aiming at

stratified overall lean mixtures, the elevated in-cylinder gas

pressure/density reduces spray penetration and produces a

more compact spray that can more easily be directed

towards the spark plug In addition, the investigations of

(Birth et al., 2006; Nouri et al., 2007) identified the

complex nature of the in-nozzle flow and, in particular, the

development of different types of cavitation that can

influence the stability of the emerging jet sprays

Mixture preparation in direct injection engines is one of

the most important processes in ensuring a successful DISI

combustion system (Zhao et al., 1997) Preparing the

desired mixture inside the combustion chamber over the

full range of engine operating conditions is quite difficult,

as the fuel/air mixing process is influenced by many time

dependant variables In this study, the spray characteristics

generated by a high pressure multi-hole injector have been

examined as a function of injection timing, in-cylinder air

charge motion, coolant temperature, and injection pressure

using the Mie scattering technique The engine

configu-ration and experimental techniques for the present

experi-ments are described in the following section, the results are

presented and discussed in section 3, and the paper ends

with a summary of the most important findings

2 EXPERIMENTAL SET-UP

2.1 Engine Design

The single cylinder research engine used in this study was

designed for optical measurements and, as such, it offers

good optical access It includes a 4-valve modern pent roof

cylinder head designed to allow spray guided operation

The optical engine set up is shown in Figure 1 (a)~(d), and

the engine configuration details are summarized in Table 1

As shown in Figure 1(a), downstream of the throttle valve,

there is a second valve installed at the inlet of one of the

ports, named the Swirl Control Valve (SCV) When this

valve is closed, in-cylinder swirl is generated The position

of this valve can be varied manually from fully open to

closed using an external gauge controller Without SCV,

this cylinder head was designed to generate high tumble

flows; and the TVRo (steady flow tumbling vortex) values

measured were 1.38 for 1000 rpm in a steady flow rig test(Karaiskos, 2005)

Optical access to the combustion chamber was providedfrom the side (vertical images) via a fused silica cylinderliner (Figure 1(b) and 1(c)) As shown in Figure 1(b), thereare also two quartz windows on both sides of the cylinderhead to provide access to the pent roof area The pistoncrown has a flat design so that an optical window can befitted to obtain horizontal images The injector and sparkplug are oriented longitudinally, as shown in Figure 1(d);the line of the spark plug and injector is in the middle of thepent roof, between the intake and exhaust valves All testswere carried out without combustion, and with the enginebeing motored Identification of the engine cycle and crankangle position was achieved by an optical pick up sensormounted on the exhaust camshaft, and a crankshaft encoder(Muirhead Vactric), which produced 1440 pulses perrevolution, thus resulting in a resolution of 0.25ºCA.Engine control was achieved by using an advanced timercard (NI PCI-6602) with in-house software (Labview),which controlled injection and ignition

The prototype injector used in the present experimentshad been designed and manufactured by Bosch specificallyfor DISI engines, and it is a high pressure six hole injectorwith the holes symmetrically arranged on the periphery of

an imaginary circle, as shown in Figure 2 The detailedspecifications of the multi-hole injector, which operateswith injection pressures up to 20 MPa, are described in

Figure 1 Engine set up: (a) Schematic of engine set-up; (b)Optical access arrangement; (c) Front view optical access(d) Cylinder head configuration

Table 1 Test engine specifications

Cylinder head Pentroof Ports Tumble/SwirlBore×Stroke

(mm) 83×92 In Vavletiming 6

oBTDC/

50oABDCCompression

ratio 10.5 Ex Valve timing 50

oBBDC/

6oATDC

Table 2 In view of the very short time available for fuel

atomization and vaporization in DISI engines, particularly

in the case of injection late in the compression stroke, the

electromechanical response of the injector becomes an

important consideration Therefore, prior to the acquisition

of spray images for the characterization of the spray

development during the injection period, the injection

delay, defined as the time between the rising edge of the

triggering signal and first appearance of liquid at the nozzle

exit, was quantified Throughout preliminary injection

testing, the injection delay time was found to be 0.7 ms,

and during convenient spray image acquisition, the real

injection delay time was adjusted to 0.2 ms with a 0.5 ms

delay in the triggering signal

2.2 Mie Scattering System

The optical set-up was used for capturing Mie scattering

images The illumination of the spray was achieved by

means of a xenon flash light directed via a couple of optical

fibers to the area of interest Qualitative and quantitative

information about the spray was extracted from

high-resolution forward illuminated images recorded with a

non-intensified 12 bit CCD PCO SensiCam camera, offering a

resolution of 1024×1240 pixels and low readout noise, in

conjunction with a Nikkor telescopic zoom lens (75~300

mm 1/4.5~5.6) Image acquisition timing is controlled by

the engine control system, which is equipped with two

general purpose triggers An active cycle frequency of

image acquisition and fuel injection was set in such a way

as to allow sufficient time (15 s) for the xenon flash light to

recharge fully

To explore the spray pattern of a high pressure 6-hole

injector, a variety of different operating modes and

condi-tions were tested, as shown in Table 3 The injection

duration was kept at 1 ms for all of the test conditions

Spray imaging was repeated three times for each time step

of each test case The spray cone angle and penetration,obtained from the Mie images, are defined and provided inFigure 2; images taken in the A-A plane view were used toobtain the spray cone angle For the investigation of in-cylinder spray characteristics, the injected sprays of the 6-hole injector were visualized in the B-B plane view atengine speeds of 1000 rpm There is a 0.2 ms delaybetween the injection trigger signal and first appearance of

a spray Therefore, spray evolution images were capturedfrom 0.3 ms ASOI to 1.1 ms ASOI at 0.1 ms time intervals

3 RESULT AND DISCUSSION3.1 Early Injection for Homogeneous Stoichiometric Opera-tion

Multi-hole injectors are known to have stable spray tures under various operating conditions The overall spraycone angle remains close to the nominal design value withincreasing chamber pressure; thus, early and late injectionduring an engine’s cycle appear to have almost identicalspray shape, affecting only the spray’s penetration in thecombustion chamber Homogeneous operation dictatesearly injection of the fuel during the induction stroke Aselection of early injection timing includes injection of fuel

struc-at ATDC 60oCA, 90oCA, and 120oCA

From the previous LDV measurement of in-cylinderflow under ‘SCV open’ or tumble flow condition (Kariskos,2005), it was realized that high velocities were generatedduring the intake process, rising to a maximum betweenATDC 60oCA and 120oCA, and then decreasing in response

to the piston motion During this period, the incoming highvelocity annular air-jet flows were directed axially towardsthe down-going piston and radially towards the exhaust.The results also showed that the generated swirl flow wasneither strong nor well defined with respect to cylinderaxis The injected spray pattern during the intake strokewith ‘SCV open’ can be strongly affected by the tumblemotion and its variation will result from the turbulence ofthe swirl motion

Evolution of the spray pattern at different injectiontimings of ATDC 60oCA, 90oCA, and 120oCA, with theSCV fully closed (maximum swirl), a fuel injection pressure

of 7 MPa, and a coolant temperature of 40oC is displayed inFigure 3 As shown, there are two distinct features in thespray structures; one is that the multiple spray plumes (jets)

Figure 2 Injector nozzle and spray view

Table 2 Specifications of multi-hole injector

Hole

Hole

diameter 140µm L/D ratio 2.14

Manufacturer BOSCH Production type Gasoline DISIProto type

Table 3 Experimental conditions

SCV position Open/Close Injection pressure 7 MPa/12 MPaCoolant

temperature 40ºC/90ºC Fuel ISO OctaneIntake air

temperature ~20ºC Operating mode Homogeneous/Stratified

280 S KIM, J M NOURI, Y YAN and C ARCOUMANIS

from the multi-hole nozzle cannot be discriminated, and

the second is a clear tilt of the overall spray towards the

exhaust side and down the same as that of the incoming

annular air jet trajectory The merging or smearing of the

spray plumes takes place as soon as the fuel plumes are

generated from the nozzle This is because the plumes are

subjected to a strong intake flow with high tumble and

swirl velocities, and high turbulence As a result, the

smaller and slower droplets are dispersed rapidly under a

highly turbulent and swirling flow, causing the separated

injected fuel plumes to smear together It is also clear from

the images that the tilt of the overall spray is in the

direction of the intake cross-flow These effects are more

evident when the elapsed time goes over 0.7 ms ASOI, the

whole spray is now inclined downstream and furthermore,

the fuel droplets of the tip edge start to be separated from

the main plume jet towards the cross-flow direction; the

latter effect may be a result of high swirling and turbulence

The extent of the separation increases with elapsed time

after the start of injection, and those of ATDC 90oCA and

120oCA SOI are more pronounced than that of ATDC

60oCA

At 0.9~1.1 ms ASOI, the spray tilt is even more

recogni-zable, with downstream injected fuel droplets largely

distri-buted in the cross-flow direction This phenomenon

repre-sents the promotion of the injected fuel distribution through

the combustion chamber The swirl flow activates the

spatial advantage of the multi-hole nozzle to accommodate

the homogeneous charge mixture At 1.1 ms of elapsed

time, a small portion of the separated fuel droplets reaches

the cylinder wall, which is undesirable The spray evolution

with tumble flow, with a fuel injection pressure of 7 MPa

and a coolant temperature of 40oC at the start of injection at

ATDC 60oCA, 90oCA, and 120oCA, is displayed in Figure

4 Similar to the spray pattern under swirl flow, and for the

same reasons, the multiple spray plumes cannot be guished The injected fuel spray plumes cannot avoid thestrong influence of the incoming air cross flow duringintake valve opening due to the injector position in thecylinder head Generally, the tumble flow does not deflectthe spray pattern as strongly as the swirl flow, and there is

distin-no fuel droplet separation phedistin-nomedistin-non; the latter indicates

no impingement on the liner The larger spray deflectionand droplets separation with swirl flow, as seen in Figure 3,clearly suggest the presence of centrifugal force acting onthe fuel droplets away from the center of the cylinder.Overall comparison with the spray patterns under theflow of Figure 3 indicates that the spatial distribution of theinjected fuel spray under tumble flow is apparently lessthan that of swirl flow, especially over the elapsed time of0.7 ms In addition, the tilt of the overall spray in thedirection of the intake cross-flow is not as much as theswirl Therefore, for a well distributed, homogenized andstoichiometric mixture, it is more important for swirl flow

to be generated in the cylinder than tumble flow 3.2 Late Injection for Stratified Lean Operation ModeThe concept of stratification needs to be clarified according

to the engine design At the time of ignition, an ignitablemixture cloud should be around the vicinity of the sparkplug This mixture cloud could be slightly rich in fuel locally,while the remaining volume of the combustion chamber isoccupied by air The size of the mixture cloud increaseswith increasing engine load, and the load is controlledquantitatively by the amount of fuel injection The mostcommon technique to achieve mixture stratification is byinjecting the fuel during the compression stroke, and afterthe closure of the inlet valve In this study, three injectiontimings during the compression stroke have been selectedATDC 270oCA, 285oCA, and 300oCA, which were definedFigure 3 Mie images during the intake stroke under swirl

flow, Pinj.=7 MPa, and Tcoolant=40ºC Figure 4 Mie images during the intake stroke under tumbleflow, Pinj.=7 MPa, and Tcoolant=40ºC

as medium and late injection timings During this period,

tumble motion still existed, but swirl flow decayed and at

ATDC 300oCA, the turbulence intensity increased linearly

across the cylinder while the weak main flow moved

towards the exhaust valve area These tumbling/swirl

velo-city values are much smaller than those of early induction,

which may suggest that the injected spray pattern during

the compression stroke may be less affected by the tumble

motion The evolution of the spray pattern at the start of

injection at ATDC 270oCA, 285oCA, and 300oCA, with the

SCV fully closed (swirl), fuel injection pressure of 7 MPa,

and coolant temperature of 40oC, is displayed in Figure 5

Not like the spray pattern of the intake stroke, the multiple

spray plumes from a multi-hole nozzle can clearly be

discriminated As mentioned before, the axial and swirl

mean velocities, and also the turbulence level, were not so

large as to overcome the spray plume momentum, and

therefore there is much less deformation and dispersion of

fuel droplets Until an elapsed time of 0.9 ms ASOI, the

spray plume patterns were similar regardless of SOI timing

However, when the elapsed time exceeds 0.9 ms ASOI, the

front shape of the tip of spray plumes can no longer

maintain its straight penetration, and is distorted slightly

perhaps due to the RMS component of swirl flow With

respect to the start of injection timing, the growth of spray

penetration is restricted by the upward moving piston and

higher chamber pressure The spray penetration of ATDC

300oCA SOI was strongly affected, and a shorter spray

penetration can be observed

The evolution of the spray pattern under conditions of

tumble flow, fuel injection pressure of 7 MPa, and coolant

temperature of 40oC is displayed in Figure 6 Similar to the

spray pattern under swirl flow, the whole spray pattern was

kept straight regardless of SOI With respect to the start of

injection timing, growth of the spray plumes maintains its

straight penetration, unlike that of the swirl flow.From the spray pattern of late injection during the com-pression stroke, it can be argued that the spray shape andpenetration were affected by the RMS component of in-cylinder flow and piston movement In particular, the spraypenetration of the latest start of injection is strongly restricted

by the upward moving piston

3.3 Temperature Effect on Spray Droplet VaporizationSince the Mie scattering technique is based on scatteredlight by liquid droplets only the remaining non-yet-vaporizedspray could be captured More specifically, assuming thatthe base spray image for characterizing evaporation would

be at the lowest available temperature, then the nation of images taken at the base and at a higher temper-ature would provide important qualitative information onthe relative percentage of liquid already vaporized, as wassuggested by (Mitroglou, 2005) The principle of this ap-proach is shown schematically in Figure 7, and the out-come would represent the probability density function ofthe liquid fuel droplets that are most likely to be evapo-rated

combi-Figure 5 Mie images during the compression stroke under

swirl flow, Pinj.=7 MPa, and Tcoolant=40ºC Figure 6 Mie images during the compression stroke undertumble flow, Pinj.=7 MPa, and Tcoolant=40ºC

Figure 7 Mie image processing for vaporizing region

282 S KIM, J M NOURI, Y YAN and C ARCOUMANIS

Figure 8 illustrates the temperature effect on spray droplet

vaporization for sprays injected at 7 MPa and 12 MPa into

the cylinder In general, the results show that a small

amount of liquid fuel is vaporized for a temperature rise

from 40oC to 90oC, this is perhaps expected since the

boiling temperature of the fuel (isooctane) is 102o~105oC

@0.1 MPa; similar results was reported (Mitroglou, 2005)

for the same increase in temperature It is also evident that

the amount of vaporize fuel is slightly more with higher

injection pressure probably due to minor improvements in

atomization and efficacy A more specific analysis is

need-ed to quantify the effect of a coolant temperature of 90oC in

spray vaporization relative to 40oC For example, taking

plane Mie images rather than surface images will help

considerably, along with taking extra images at

temper-atures above the fuel boiling point Overall, the present

results show that only small amounts of liquid are expected

to vaporize during the injection, and this would most likely

happen around the edges of the individual fuel spray jets,

away from the injector exit

3.4 Spray Penetration and Cone Angle

The spray penetration and spray cone angle at differentinjection timing, injection pressure, and coolant temperature

at 1000 rpm are plotted in Figure 9 and Figure 10 Thespray penetration of 60oCA SOI and Tcoolant=40oC is shown

in Figure 9(a) The penetration is affected by fuel pressure so that at initial stage till 0.5 ms ASOI, the spraypenetration at 7 MPa is a little greater than that of 12 MPa,mainly because the mechanical operational delay time ofthe injector at 12 MPa is longer However, the injected fueldroplets had a substantial momentum, as a result of thehigher fuel pressure, and consequently, penetrated furtherinto the cylinder than those injected at a lower injectionpressure as the time ASOI increases

injection-From the elapsed time of 0.6 ms ASOI, the spray tion at 12 MPa becomes greater than that at 7 MPa Thepenetration continues to increase until 0.8 ms ASOI, andfrom 0.8 ms ASOI onward, the penetration stops at about

penetra-40 mm due to loss of droplet momentum The spraypenetration of 300oCA SOI under swirl and tumble floware shown in Figure 9(b) and Figure 9(c), respectively Thespray penetration during the compression stroke has asimilar trend to that of the intake stroke The penetration isalso affected by fuel injection pressure But additionally, it

is strongly affected by the chamber pressure (movingpiston), which causes a maximum penetration of 35 mm,shorter than that of the intake stroke After 0.9~1.0 msASOI, the spray tip starts to impinge on the piston Theinjected fuel of high pressure reaches the piston earlier thanthat of lower pressure Therefore, it is necessary to carefullyconsider the extent of fuel impingement according to the fuelpressure But, the temperature effect on the spray penetration

is small and not as noticeable as the fuel pressure Theplane (A-A), where the overall spray angle was calculated,

is shown in Figure 2, and the angle was measured betweenthe extreme edges of the two outer jet sprays near theinjector tip, where the effects of the cross-flow wasminimum Figure 10 shows the spray cone angle duringintake and compression stroke, and at different injectionpressures and coolant temperatures

The results showed that the overall spray angle remainedconstant and almost independent of injection pressure,chamber pressure, and coolant temperature There is also asmall and gradual reduction in the overall spray cone angle

Figure 8 Effect of the coolant temperature on fuel

vapori-zation at ATDC 300°CA SOI under tumble flow

Figure 9 Spray penetration during intake and compression strokes

with the elapsed time ASOI, which is similar for all

conditions tested, making the overall spray cone angle

smaller than that of the nominal value This can be related

to the complex flow structure inside the nozzle hole,

especially in the presence of different types of cavitation,

depending on pressure differences across the nozzle due to

the opening of the needle In particular, there is a geometric

cavitation that forms on the upper part of the nozzle, and

can affect the trajectory of the exiting fuel jets by forcing

them downwards

4 CONCLUSION

Spray characteristics of a high pressure 6-hole multi-hole

injector were investigated in an optical engine using Mie

scattering The results were obtained at an engine speed of

1000 rpm, and the effects of injection timing, in-cylinder

charge motion, coolant temperature, and injected fuel

pressure were investigated The most important findings

are summarized below:

(1) To obtain a homogeneous and stoichiometric mixture,

in-cylinder swirl proved to be far more effective than

tumble flow during the intake stroke The results showed

a clear shift of the spray jets in the direction of the intake

cross-flow

(2) The spray pattern of late injection during the

compre-ssion stroke was little affected by tumble and swirl

cross-flow However, the effect of increased chamber

pressure due to piston movement was considerable in

limiting the spray jet penetration

(3) The effect of coolant temperature on fuel droplets

vaporization was found to be small when the

temper-ature was raised from 40oC to 90oC

(4) Fuel pressure promotes spray penetration although,

during the compression stroke, it is strongly affected by

the upward moving piston causing an increase in the air

density in the cylinder

(5) The overall spray cone angle was found to be constant

and almost independent of injection pressure, chamber

pressure, and coolant temperature A gradual reduction

in the overall spray angle was also found with elapsed

time after the start of injection, which can be related to

the development of cavitation in the nozzle holes

ACKNOWLEDGEMENT− This work was supported by the Korea Research Foundation Grant (KRF-2005-013-D00009) And the authors would like to thank Dr N Mitroglou for his contribution to this research programme and Mr Tom Fleming and Mr Jim Ford for their valuable technical support during the course of this work.

REFERENCESBirth, I G., M Rechs, M U., Spicher, U and S Bernhardt,

S (2006) Experimental investigation of the in-nozzleflow of valve covered orifice nozzle for gasoline directinjection 7th Int Symp Internal Combustion Diagnostics,

59−78, Kurhaus Baden-Baden

Fraidl, G K., Piock, W F and Wirth, M (1996) Gasolinedirect injection actual trends and future strategies for injec-tion and combustion systems SAE Paper No 960465, 95−

111

Honda, T., Kawamoto, M., Katashiba, H., Sumida, M.,Fukutomi, M and Kawajiri, K (2004) A study of mixtureformation and combustion for spray guided DISI SAE Paper No 2004-01-0046

Karaiskos, I E (2005) Spray Structure and Mixture bution in a Direct Injection Gasoline Engine. Ph D.Dissertation, University of London

Distri-Li, T., Nishida, K and Hiroyasu, H (2004) tion of initial spray from a D.I gasoline injector by holo-graphy and laser diffraction method Int J Atomization and Sprays, 14, 477−494

Characteriza-Lippert, A M., El Tahry, S., Huebler, M S., Parrish, S E.,Inoue, H., Noyori, T., Nakama, K and Abe, T (2004).Development and optimisation of a small-displacementspark-ignition direct-injection engine-stratified operation.SAE Paper No 2004-01-0033

Mitroglou, N (2005) Multi-Hole Injectors for Injection Gasoline Engines Ph D Dissertation TheCity University

Direct-Mitroglou, N., Arcoumanis, C., Mori, K and Motoyama,

Y (2005) Mixture distribution in a multi-valve spark ignition engine equipped with high-pressure multi-hole injectors ICOLAD 2005, 27−40

twin-Mitroglou, N., Nouri, J M., Gavaises, M and Arcoumanis,

C (2006) Flow and spray caracteristics in spray-guidedFigure 10 Spray cone angle during intake and compression strokes

284 S KIM, J M NOURI, Y YAN and C ARCOUMANIS

direct injection engines J Engine Research 7, 3, 255−

270

Mitroglou, N., Nouri, J M., Yan, Y., Gavaises, M and

Arcoumanis, C (2007) Spray structure generated by

multi-hole injectors for gasoline direct injection engines

SAE Paper No 2007-01-1417

Nouri, J M and Whitelaw, J H (2002) Effect of chamber

pressure on the spray structure froma swirl pressure

atomiser for direct injection gasoline engines 1st Int.

Conf Optical Diagnostics, ICOLAD, 1, 121−129

Nouri, J M and Whitelaw, J H (2006) Impingement of

gasoline sprays on angled plates Int J Atomization and

Sprays 16, 6, 705−726

Nouri, J M., Mitroglou, N., Yan, Y and Arcoumanis, C

(2007) Internal flow and cavitation in a multi-hole

injector for gasoline direct injection engines SAE Paper

No 2007-01-1405

Nouri, J M and Whitelaw, J H (2007) Impingement of

gasoline sprays on angled plates Int J Atomization and

Sprays 17, 6, 1−20

Ortmann, R., Arndt, S., Raimann, J., Grzeszik, R and

Wurfel, G (2001) Methods and analysis of fuel injection,

mixture preparation and charge stratification in different

direct-injected SI engines. SAE Paper No 2001-01-0970

Preussner, C., Doring, Fehler, S and Kampmann, S (1998)

GDI: interaction between mixture preparation combution

system, and injector performacne SAE Paper No 980498.Shim, Y S., Choi, G M and Kim, D J (2008) Numericalmodeling of hollow-cone fuel atomization, vaporizationand wall impingement processes under high ambienttemperatures Int J Aumotive Technology 9, 3, 267−

275

Skosberg, M., Dahlander, P., Lindgren, R and Denbratt, I.(2005) Effects of injector parameters on mixture formationfor multi-hole nozzles in a spray-guided gasoline DIengine SAE Paper No 2005-01-0097

Wirth, M., Piock, W F., Fraidl, G K K., Schoeggi, P andWinklhofer, E (1998) Gasoline DI engines the completesystem approach by interaction of advanced developmenttools SAE Paper No 980492

Wirth, M., Zimmermann, D., Friedfeldt, R., Caine, J., Schamel,A., Davies, M., Peirce, G., Storch, A., Ries-Müller, K.,Gansert, K P., Pilgram, G., Ortmann, R., Würfel, G andGerhardt, J (2004) A cost optimised gasoline sprayguided direct injection system for improved fueleconomy,seminar on fuel economy and enginedownsizing Institu- tion of Mechanical Engineers, One Birdcage Walk,London, 13 May 2004

Zhao, F., Lai, M and Harrington, D L (1997) A review ofmixture preparation and combustion control strategiesfor SIDI gasoline engines SAE Paper No 970627

INFLUENCE OF INJECTION PARAMETERS ON THE TRANSITION FROM PCCI COMBUSTION TO DIFFUSION COMBUSTION

IN A SMALL-BORE HSDI DIESEL ENGINE

T FANG 1)* , R E COVERDILL 2) , C.-F F LEE 2) and R A WHITE 2)

1)Department of Mechanical and Aerospace Engineering, North Carolina State University,

3182 Broughton Hall-Campus Box 7910, 2601 Stinson Drive, Raleigh, NC 27695, USA

2)Department of Mechanical Science and Engineering, University of Illinois at Urbana-Champaign,

1206 West Green Street, Urbana, IL 61801, USA

(Received 11 June 2008; Revised 13 October 2008)

ABSTRACT− In this paper, the influence of injection parameters on the transition from Premixed Charge Combustion Ignition (PCCI) combustion to conventional diesel combustion was investigated in an optically accessible High-Speed Direct- Injection (HSDI) diesel engine using multiple injection strategies The heat release characteristics were analyzed using in- cylinder pressure for different operating conditions The whole cycle combustion process was visualized with a high-speed video camera by simultaneously capturing the natural flame luminosity from both the bottom of the optical piston and the side window, showing the three dimensional combustion structure within the combustion chamber Eight operating conditions were selected to address the influences of injection pressure, injection timing, and fuel quantity of the first injection on the development of second injection combustion For some cases with early first injection timing and a small fuel quantity, no liquid fuel is found when luminous flame points appear, which shows that premixed combustion occurs for these cases However, with the increase of first injection fuel quantity and retardation of the first injection timing, the combustion mode transitions from PCCI combustion to diffusion flame combustion, with liquid fuel being injected into the hot flame The observed combustion phenomena are mainly determined by the ambient temperature and pressure at the start of the second injection event The start-of-injection ambient conditions are greatly influenced by the first injection timing, fuel quantity, and injection pressure Small fuel quantity and early injection timing of the first injection event and high injection pressure are preferable for low sooting combustion

KEY WORDS : HSDI diesel engine, Conventional diesel combustion, PCCI combustion

1 INTRODUCTION

Because of the increasing threat of limited fossil fuel

resources and the worldwide concern of environmental

issues, emissions regulations for current engines are becoming

increasingly more stringent Direct Injection (DI) diesel

engines are attractive power sources due to their superior

fuel economy and excellent reliability However, oxides of

nitrogen (NOx) and Particulate Matter (PM) must be

reduced for diesel engines to meet the stricter emissions

standards New techniques and combustion concepts have

been developed to solve the problems

Homogeneous Charge Compression Ignition (HCCI)

com-bustion is a promising technique that provides a unique

approach to simultaneously reduce NOx and PM emissions

while maintaining high thermal efficiency For HCCI

com-bustion, a premixed or ideally homogeneous air-fuel

mix-ture auto-ignites due to compression; it is a bulk

combus-tion, eliminating local high temperature regions

Conse-quently, the NOx emissions are extremely low comparedwith conventional diesel combustion and Spark Ignition(SI) combustion In addition, because the air-fuel mixture

is premixed, there is no locally rich region, so soot and PMare also greatly reduced Early studies of the HCCI com-bustion mode were carried out in two-stroke engines(Onishi et al., 1979; Noguchi et al., 1979) and in four-stroke engines (Najt and Foster, 1983; Thring, 1989) byusing heavy Exhaust Gas Recirculation (EGR) It wasshown that, in the HCCI combustion mode, the ignitionprocess is controlled by low temperature (950 K) hydro-carbon oxidation kinetics, while the energy release process

is controlled by high temperature (above 1000K) carbon oxidation

hydro-Multiple injection strategies have been reported forsimultaneous reduction of NOx and PM in both large bore

DI diesel engines (Nehmer et al., 1994; Tow et al., 1994;Han et al., 1996) and small-bore high-speed DI dieselengines (Zhang, 1999; Tanaka et al., 2002; Chen, 2000).Several studies (Nehmer and Reitz, 1994; Tow et al., 1994;Han et al., 1996) have shown that pulsed injections may

*Corresponding author. e-mail: tfang2@ncsu.edu

286 T FANG, R E COVERDILL, C.-F F LEE and R A WHITE

provide a method to reduce PM emissions and allow for the

reduction of NOx from controlled pressure rise Late

injec-tion in double or triple injecinjec-tion strategies can promote the

particulate oxidation process Reduced soot emissions are

due to the fact that soot-producing rich regions are not

replenished when the injection pulse is terminated and

restarted The combustion mode in these studies can be

categorized as conventional diesel combustion In addition

to the studies in heavy duty DI diesel engines,

investi-gations have also been performed in small-bore HSDI diesel

engines using multiple injection strategies The effects of

pilot injection on the combustion process were studied

experimentally by Zhang (1999) and Tanaka et al. (2002)

Simultaneous reduction of combustion noise and emissions

is possible by decreasing the influence of pilot burned gas

through minimizing the fuel quantity and advancing the

injection timing of the pilot injection Simultaneous

reduc-tion of NOx and PM by using multiple injecreduc-tions was

implemented in a small diesel engine (Chen, 2000) by

optimizing the combinations of EGR rate, pilot timing and

quantity, main timing, and dwell between the main and

pilot injections Post injection was shown to be effective in

reducing PM due to the improved particulate oxidation

process In the results of these papers, the combustion

modes were not limited to conventional diesel combustion

Some evidence of the PCCI combustion mode can be

found in the heat release rate curves Some results (Tanaka

mode using double injections, as discussed in the following

sections

Hashizume et al (1998) proposed an HCCI solution for

higher load operating conditions The combustion is named

MULtiple stage DIesel Combustion (MULDIC) on the basis

of the PREmixed lean DIesel Combustion (PREDIC)

con-cept (Takeda et al., 1996) The first stage is premixed lean

combustion (PREDIC), and the second stage is diffusion

combustion under high temperature and low oxygen

condi-tions Smoke and NOx were reduced by MULDIC even at

an excess air ratio of 1.4 Further studies on the MULDIC

concept were done in the same research group by Akagawa

et al (1999) In this study, they developed a new pintle type

injector for reduced fuel penetration, especially for the

early injection The top-land crevice volume, namely the

wall quenching volume, was also reduced The results

showed reductions of THC and CO emissions At the same

time, NOx and smoke can also be reduced at high load

conditions

A multi-pulse injection strategy was used by Su et al

(2003) in their study of HCCI combustion in an HSDI

diesel engine They used multiple short injection pulses for

the early injection or followed by a main injection near Top

Dead Center (TDC) When the load is less than 9.3 bar

Indicated Mean Effective Pressure(IMEP), reductions of

both smoke and NOx were obtained HCCI combustion in

a small bore HSDI diesel engine was also investigated

using early multiple short injection pulses during the

com-pression stroke by Helmanteland Denbratt (2004) In order

to decrease the fuel wall impingement, a small-includedangle injector was used The results showed a dramaticreduction of NOx Smoke emissions also showed a signifi-cant reduction, while HC and CO emissions substantiallyincreased

Investigations by Hasegawa and Yanagihara (2003) ployed two injections in the HCCI combustion mode,referred as UNIform BUlky combustion System (UNIBUS).The first injection was used as an early injection for earlyfuel mixing and to advance the changing of fuel to lowerhydrocarbons, while the second injection was used as anignition trigger Bulk combustion was observed in thecombustion chamber Low NOx and smoke were possible

em-in both em-injections usem-ing this combustion concept Anothertwo-stage diesel fuel injection HCCI combustion study wasdone in a single cylinder small diesel engine (Kook andBae, 2004) A large fraction of fuel was injected earlyduring the compression stroke or even during the inductionstroke A second injection with a small amount of fuel wasinjected near the compression TDC to ignite all of the air-fuel mixture The experimental results showed that thesecond injection could only be used as a combustion triggerfor low intake air temperature The first injection timingshould be advanced earlier than 100 CAD BTDC toachieve homogeneous and non-luminous combustion NOxwas greatly reduced using this injection strategy HC, CO,and fuel consumption were higher than in conventionaldiesel combustion

Conceptually speaking, HCCI combustion is an idealoperation mode for low emission diesel engines (Choi et

al., 2004) However, in a real diesel engine, it is quite cult to homogeneously mix air and fuel using in-cylinderdirect injection strategies, even with a very early injectionduring the suction stroke (Swami Nathan et al., 2007) Aheterogenous premixed charge often occurs under theseinjection strategies Mixture heterogeneity often exists, evenfor very early in-cylinder injection timings In general, Pre-mixed Charge Compression Ignition (PCCI) combustion is

diffi-a more diffi-accurdiffi-ate terminology for these conditions thdiffi-an

“HCCI” PCCI only requires a premixed charge, and themixture is not required to be homogeneous “PCCI” is abroader concept than “HCCI” Most of the above mention-

ed combustion modes are types of PCCI combustion

In these previous studies, the combustion processes wereoften visualized through an optical engine with modifiedpiston geometries The replacement of the true piston shapechanges the flow field into which the fuel is injected In thiswork, the investigation uses an optical engine with a realisticpiston geometry Among the current operating conditions, atransition from PCCI combustion mode to conventionaldiesel combustion mode was seen for the second maininjection The influential factors such as injection pressure,injection timing, and injection fuel quantities are studiedand the effects of the first injection parameters on thecombustion mode for the second injection are addressed

2 OPTICAL ENGINE AND FACILITY

A single-cylinder DIATA research engine supplied by Ford

Motor Company was modified into the optical engine used

for the current experimentation Key aspects of the DIATA

engine are listed in Table 1 Optical access to the

com-bustion chamber was attained through the side window or

through the fused silica piston top The optical engine

design maintains the geometry of the ports and combustion

chamber of the original engine A complete description of

the optical engine can be found in a previous publication

(Mathews et al., 2002) A Bosch common-rail electronic

injection system was used, and was capable of injection

pressures up to 1350 bar A valve covered orifice injector

with six 0.124 mm holes placed symmetrically in the nozzle

tip and a spray cone angle of 150 degrees were used The

injector was fitted with a needle lift sensor monitoring the

needle operation throughout injection A Phantom v7.0

high-speed digital video camera was used to capture the natural

flame emission for the whole cycle National Instruments

LabView version 6.0 was used as the data acquisition and

timing control software An optical shaft encoder with 0.25

crank angle resolution was used to provide the time basis

The engine temperatures and pressures were monitored

through a multifunction data acquisition board

3 ENGINE OPERATING CONDITIONS

The results presented in this paper are based on operating

conditions considered typical for this engine Intake

temper-atures and pressures were increased to match the TDC

conditions of the metal engine with the same geometry and

operating conditions The operating conditions are

summari-zed in Table 2 The fuel quantities of the first injections

were calibrated and injected at given injection timings The

main injection pulse durations were adjusted to match the

load for all of the cases to be 5.0 bar IMEP The injectiontiming of the main injection was set at TDC for all of thecases The fuel used was a low-sulfur European Diesel fuel,selected properties of which are shown in Table 3 Due tothe extensive optical access provided by the optical DIATAengine, 3-D like combustion imaging was feasible (Fang et

al., 2005, 2006, 2007, 2008; Miles, 2000) Combustionimages were obtained using the high-speed video camera

by setting the operating frame rate at 12000 frames persecond with the resolution at 512×256 to capture theimages from the bottom and side For all of the cases, theexposure time was 2 ms

4 RESULTS AND DISCUSSIONS4.1 In-cylinder Pressure and Heat Release AnalysisThe optical engine was warmed up by circulating heatedcoolant and lubricating oil to simulate a warm engineenvironment The engine operated in skip fire mode inorder to reduce the heat load of the quartz piston, with oneinjection cycle followed by 12 motoring cycles Pressuredata were recorded and saved to the computer for postprocessing

Pressure traces for the eight cases are shown in Figures1a and 2a In the plots, 360 CAD corresponds to thecompression Top Dead Center (TDC) It is seen from thefigures that high injection pressure results in faster com-bustion, and thus more rapid pressure increase, due tobetter fuel spray atomization and mixing Combustion noise,which is directly relevant to pressure rise rate, will behigher for the higher injection pressure cases Some knock-like combustion behaviors are seen for the high injection

Table 1 Specifications of the single cylinder DIATA research

Intake valve diameter 24 mm

Maximum valve lift 7.30/7.67 mm (Intake/Exhaust)

Intake valve opening 13 CAD ATDC

(at 1 mm valve lift)Intake valve closing 20 CAD ABDC

(at 1 mm valve lift)

Ex valve opening 33 CAD BBDC

(at 1 mm valve lift)Exhaust valve closing 18 CAD BTDC

(at 1 mm valve lift)

Table 2 Summary of engine operating conditions

Casenumber pressure[bar]Rail

First timing[CAD ATDC]

Pilot quantity[mm3]

Main tion timing[CAD ATDC]

injec-IMEP[bar]

288 T FANG, R E COVERDILL, C.-F F LEE and R A WHITE

pressure cases Higher in-cylinder pressure peaks are seen

for Cases 5~8 than for Cases 1~4 Earlier first injection

timing and small fuel quantity lead to a longer ignition

delay for the second injection Long ignition delay for the

second injection results in higher pressure increase rate and

therefore higher combustion noise

The heat release rates are illustrated in Figures 1(b) and

2(b) Results of the heat release rates support the

obser-vations in the in-cylinder pressure plots Ignition delays are

seen to be shorter for higher first fuel quantity and later first

injection timing Narrower and higher heat release peaks

are observed for the higher injection pressure cases, ing more rapid combustion and concentrated heat releaseprocesses, while flatter and broader heat release ratepatterns are seen for the lower injection pressure cases Forthe second injection combustion, it is seen that some of thehigh injection pressure cases are close to PCCI combus-tion However, for the low injection pressure cases, diffu-sion combustion becomes more apparent, with increasingfirst injection fuel quantity and decreasing retarding firstinjection timing The combustion mode transition fromdiffusion combustion to PCCI combustion is observed inFigure 1 In-cylinder pressures (a), heat release rates (b),

show-and cumulative heat release (c) for Cases 1~4 Figure 2 In-cylinder pressures (a), heat release rates (b),and cumulative heat release (c) for Cases 5~8

the cases investigated.

The cumulative heat release curves for these cases are

illustrated in Figures 1(c) and 2(c) The released heats of

each injection are shown in Table 4 The total heat release

for the low injection pressure cases is about 386~388

joules, while for high injection pressure cases, it is about

371~379 joules Since the work outputs for these cases are

similar based on the load match, lower energy input

indi-cates higher cycle efficiency Because the high injection

pressure cases have fast burns more like constant-volume

combustion than the low injection pressure cases, the cycle

thermal efficiency for high injection pressure is generally

higher than that of the low injection pressure cases The

released heats for the first injection show large differences

for different first injection parameters Table 4 shows that

early injection timing and higher injection pressure result

in less complete combustion for the first injection, which

leads to lower in-cylinder temperature and pressure at the

start of the second injection The combustion mode of the

second injection depends on the ambient temperature and

pressure at the start of injection Therefore, the combustion

mode of the second injection greatly depends on the first

injection parameters

The in-cylinder temperature is estimated based on the

intake condition and in-cylinder pressure using the ideal

gas law Some critical temperatures for the eight cases are

shown in Figure 3 It is found that higher injection pressure

results in lower in-cylinder temperature at the start of the

second injection The temperature is lower for early first

injection timing and smaller first fuel quantity These

observations are consistent with the heat release analysis

The in-cylinder temperature at the start of the second

injec-tion event influences the igniinjec-tion delay A high temperature

leads to a short ignition delay, which causes the overlap of

the liquid jet with the hot flame, namely typical diffusion

combustion For the lower ambient temperature and higher

injection pressure of the second injection, the combustion

mode is close to PCCI combustion with little evidence of

diffusion combustion With the increase of ambient

temper-ature, the diffusion flame becomes more pronounced, as

shown in the heat release rate of the low injection pressure

cases The maximum in-cylinder bulk temperatures are

quite similar for all of the cases Slightly higher valuescan be found for high injection pressures due to thefaster burning process (Van Gerpen et al., 1985; Kobayashi

et al., 1992)

4.2 Flame LuminosityThe combustion process was visualized using the high-speed video camera described in Section 2 The bottomview combustion images were used to compute the flameluminosity by summing the pixel values For each case, 5sets of combustion movies were taken and the flameluminosities were obtained by averaging the 5 sets of data

Table 4 Released heats for the eight operating conditions

Figure 3 Estimated bulk in-cylinder temperature

Figure 4 Flame luminosity time history for the eight cases

290 T FANG, R E COVERDILL, C.-F F LEE and R A WHITE

The time histories of flame luminosity for the eight cases

are illustrated in Figure 4 For low injection pressure, it is

found that the flame luminosity curves for the four cases

are similar, with long tails These long tails are believed to

be due to a slow post soot oxidation process Similar

obser-vations for high injection pressure cases can be seen for

Cases 6 and 8, while Case 5 and 7 have different flame

luminosity characteristics The flame luminosity is mainly

dependent on temperature and soot concentration Under

similar temperature distributions, higher flame luminosity

indicates higher soot formation Cases 5 and 7 have much

lower flame luminosity than other cases, showing low

sooting and/or low temperature combustion for these two

cases The combustion duration is longer for low injection

pressure cases because of reduced mixing and the slow post

oxidation process

The derivatives of flame luminosity are shown in Figure

5 and, to some extent, show the soot formation rate and

oxidation rate during the combustion process In general,

higher positive peaks can be seen in the plots for the high

injection pressure cases, showing faster combustion and

soot formation processes The duration for the positive

flame luminosity increase rate is shorter for the high

injection pressure cases However, the negative peaks are

much higher for the high injection pressure cases,

indicat-ing a higher soot oxidation rate for these cases The

negative peak timings are quite close to the maximum

in-cylinder temperature timings This confirms that a highertemperature results in a higher soot oxidation rate It isinteresting to note that Case 5, namely the typical PCCIcombustion case, shows a lower increasing rate andrelatively lower oxidation rate, which indicates that pre-mixed combustion has a lower soot formation rate with lessfuel rich regions

Using the same logic as in a previous study (Fang et al.,2005), the Flame Luminosity Parameter can be defined asthe ratio of the average flame luminosity over the releasedheat for the main injection The Flame Luminosity Para-meters (FLPs) are shown in Figure 6 A lower FLP valueindicates lower sooting or lower combustion temperature

If the soot concentration and temperature are similar, alower FLP value also indicates higher combustion effici-ency Lower values indicate better combustion performancefor that injection event From the results, it is apparent thathigher injection pressure benefits the combustion perfor-mance Earlier first injection timing also greatly influencesthe value of FLP A combination of high injection pressureand earlier first injection timing with small fuel quantityresults in a PCCI combustion process with the best per-formance

4.3 Flame Spatial Fluctuation (FSF) and Flame geneity (FNH)

Non-homo-Based on the definitions of FSF and FNH (Fang et al.,2005), both parameters were computed for all the cases.For each case, 5 sets of combustion images were used toobtain an averaged value For completeness, the definitions

of the two parameters are also listed below The firstparameter is defined as the Flame Spatial Fluctuation (FSF)

as follows:

(1)where I i,j is the captured flame radiation intensity at pixelposition (i,j) and is the mean flame radiation intensityfor an image at a certain crank angle The flame non-homogeneity (FNH) is defined as the sum of the length of

spatial gradients for the images over all of the pixels:

(2)where and are the partial differentiation in

the x and y direction, respectively The difference in these

two parameters has been discussed by Fang et al. (2005)

The FSF and FNH for the eight cases are plotted in Figures

7 and 8, respectively

Figure 7 shows that the diffusion dominant combustion

process has higher FSF values, showing highly substantial

fluctuations in flame spatial distributions It is also noted

that the high injection pressure diffusion combustion cases,

namely Cases 6 and 8, have slightly higher FSF values,

indicating that high injection pressure might lead to a more

fluctuating diffusion flame On the other hand, the

pre-mixed dominant combustion process has lower FSF values

due to a more uniform flame distribution for the cases such

as Case 5 and Case 7 Compared with flame luminosity, the

FSF peak timing is later than the flame luminosity peak

timing, which shows that high flame luminosity does not

imply high flame fluctuation or non-uniformity The reason

for this can be explained as follows by referencing the

combustion images in a later section It has been shown

previously (Mathews et al., 2003) that for multiple

injec-tion strategies, the combusinjec-tion flame fills the entire

com-bustion chamber, including the squish region at the early

stage of combustion, and the late cycle combustion flame is

mainly in the bowl region The same observation can be

seen in the combustion images discussed in the followingsection At an early stage, flame fills the whole field ofview, which will lead to a higher value of flame luminosity

by summing up the pixel values But for FSF, a moredistributed flame structure results in a smaller value Withcontinued combustion, the flame is more concentrated inthe bowl region and has a donut shape structure Althoughthe flame luminosity is reduced at this time, a moreconcentrated high flame intensity donut shaped regionleads to a larger value of FSF Such a characteristic of theflame development process can be clearly illustrated by thedefined FSF

The FNH time histories for the eight cases are depicted

in Figure 8 Trends similar to the FSF can be seen fordifferent injection parameters Higher injection pressureslead to lower FNH values The diffusion dominant com-bustion cases have higher FNH values, indicating moreheterogeneity, while premixed dominant combustion caseshave lower values, which indicates more homogeneouscombustion An obvious difference of FNH from FSF isthe peak timing It can be seen that the FNH peak timing isoften earlier than the FSF peak timing Compared with theflame luminosity results, it is seen that the FNH peaktiming is close to the flame luminosity peak timing Thepeaks of FNH are mainly due to the jet structure or flameedges in the combustion images The combination of flamelocal intensity and jet structure determined the FNH values.Because the later cycle combustion has no jet structure, theFNH value will be reduced for late cycle combustion

Figure 7 FSF time history for the eight cases Figure 8 FNH time history for the eight cases

292 T FANG, R E COVERDILL, C.-F F LEE and R A WHITE

Premixed combustion has little jet structure as well as low

flame intensity, and therefore has a lower value of FNH

Based on the above discussions, some characteristic

tim-ings can be found from the time histories of the discussed

parameters, as shown in Figure 9 Because the relationships

of the peak flame luminosity timing with the peak FSF and

FNH timings have been discussed before, this study will

focus on other timings One interesting timing difference is

between the peak estimated in-cylinder temperature timing

and the peak flame luminosity timing It is found that the

peak flame luminosity timing is always earlier than the

peak bulk temperature timing, where the reason is that the

in-cylinder bulk temperature depends mainly on the heat

release process and expansion work Heat is still being

released during the soot oxidation process, and the

in-cylinder temperature can increase as long as the expansion

work is less than the heat release, which is often the case at

the early stage of expansion On the other hand, the flame

luminosity is determined by temperature and soot

concen-tration Under similar local temperatures, flame luminosity

peaks often occur at the highest soot concentration, with

the soot formation rate balanced by the soot oxidation rate

When the soot oxidation rate is higher than the formation

rate, flame luminosity will be reduced Therefore, the peak

in-cylinder temperature often occurs later than the peak

flame luminosity timing

Another important timing difference is the end of

injec-tion timing and the timing of early flame appearance The

difference of these two timings gives an overlap time,

which might determine the combustion mode If the

over-lap time is less than zero, there will be purely premixed

combustion Otherwise, the combustion mode will be some

kind of diffusion flame The overlap times of liquid spray

and early flame are illustrated in Figure 10 for the eight

cases The results are grouped into two categories,

includ-ing a lower injection pressure group and a higher injection

pressure group The data points were also fitted with a

straight line to estimate the variation trend It is seen that

lower injection pressure results in a larger value of overlap

time, and all the overlap times are positive for these cases,

which indicates diffusion flame combustion However, forhigher injection pressure the overlap time is smaller and iseven negative for Case 5, leading to premixed combustion.The overlap time increases with ambient temperature at thestart of the second injection for both injection pressures.The ambient temperature is mainly dependent on the para-meters of the first injection An earlier first injection withsmall fuel quantity using higher injection pressure is prefer-able for obtaining PCCI combustion Another approach is

to retard the second injection timing after TDC with alower SOI temperature, which is used in UNIBUS com-bustion (Hasegawa and Yanagihara, 2003)

4.4 Combustion ImagesThe digital combustion images obtained using the high-speed video camera were processed using the same color-map and scales to compare different injection parameters.The camera was operated at 12000 frames per second Thisframe rate corresponds to 0.75 CAD intervals between twosequential images at 1500 rpm For each case, 5 com-bustion movies were taken and a typical whole cycle movie

Figure 9 Some critical timings for the eight cases Figure 10 Liquid spray and early flame overlap time versus

estimated in-cylinder temperature at SOI for the eight cases

Figure 11 Combustion images Case 1 All times shown inCAD ASOI in the brackets

was selected for analysis and presentation For each set of

combustion images, 6 images are presented to show the

combustion flames at different times, including two early

flame images (the first two images), one at peak flame

luminosity timing (the third), one at peak FSF timing (the

4th), and two late cycle flame images (the 5th and 6th) In

order to enhance the image contrast, the contrast factor for

the first two images and the sixth image were adjusted

when presenting them in the figures

The combustion images for the eight cases are shown in

Figures 11~18, respectively The first two images in each

figure show the ignition points and early flame

develop-ments The ignition processes are consistent with the

pre-ssure and heat release rate results For most of the cases,

the ignition points are located in the bowl region near the

spray tip areas From the side window images, it is found

that the early flame pockets are located in the near wall

region of the bowl However, for Cases 2 and 4, the

ignition points are located more upstream than in the other

cases because of high ambient temperature at the start of injection This is consistent with previous work with

differ-ent injection pressures for convdiffer-entional diffusion bustion (Minami et al., 1990) Higher first injection fuelquantity and later first injection timing lead to earlier ap-pearances of the flame points in the combustion chamber.Diffusion flame combustion is clearly seen for all of thelow injection pressure cases and for Cases 6 and 8, withliquid fuel being injected into hot flame Some evidencecan also be seen for Case 7 However, an apparent pre-mixed combustion is seen for Case 5 Lower injection pre-ssure cases have stronger diffusion flames than the higherinjection pressure cases The strength of the diffusion flamecorrelates well with the ambient temperature at the start ofthe second injection Therefore, the factor resulting in ahigher ambient temperature at the start of second injectionleads to a stronger diffusion flame Higher first fuel quan-tity and later injection timing cause stronger diffusion flames.For most of the cases, the flame fills the squish region.However, little flame is seen in the squish region for Cases

com-Figure 12 Combustion images Case 2 All times shown in

CAD ASOI in the brackets

Figure 13 Combustion images Case 3 All times shown in

CAD ASOI in the brackets

Figure 14 Combustion images Case 4 All times shown inCAD ASOI in the brackets

Figure 15 Combustion images Case 5 All times shown inCAD ASOI in the brackets

294 T FANG, R E COVERDILL, C.-F F LEE and R A WHITE

5 and 7; most of the flame is in the bowl region, even in the

early stage of combustion In the late combustion period,

most of the flame is in the bowl region with a donut shape

due to the strong swirling motion in the combustion

chamber

For the combustion images at peak flame luminosity

timing, namely the third image, large differences can be

seen for differing injection pressure cases For the low

injection pressure cases, the flame is more widespread than

in the higher injection pressure cases Most of the higher

injection pressure flame is concentrated in the bowl region

For images at the peak FSF value timings, the flame is

similar for all of the cases and has a donut shape It is more

concentrated in a circular region between the injector tip

and bowl wall This clearly explains why the peak flame

luminosity timing is earlier than the peak FSF timing, as

mentioned previously For the higher injection pressure

cases, the difference between these two timings becomes

smaller because both timings have similar flame structures

for most of the cases except Case 8

From the late cycle combustion images (the fifth andsixth images), much weaker late cycle flame intensity isseen in the higher injection pressure cases High injectionpressure results in a higher soot oxidation rate undersimilar conditions This higher soot oxidation rate can beattributed to a higher combustion temperature under higherinjection pressure (Kobayashi et al., 1992) A fast sootoxidation rate results in less soot going out to the exhaust

5 CONCLUSIONS

In the current work, the influences of the injection meters on the combustion mode transition were investi-gated The effects of injection timing, injection fuel quan-tity, and injection pressure were discussed Several para-meters were defined and used to evaluate the flame struc-ture and combustion performance Some observations andconclusions are listed as follows:

para-The first injection parameters affect the ambient ment at the start of the second injection event and influencethe combustion mode for the second injection Highersecond injection SOI temperatures are seen for higher firstinjection fuel quantities, later first injection timing, andlower injection pressures The SOI temperature greatlyinfluences the spray and flame overlap time, which directlydetermines the combustion mode

environ-A small fuel quantity, early injection timing of the firstinjection event, and high injection pressure are preferablefor low sooting and/or low temperature combustion; Combustion visualization results show the transition fromPCCI combustion mode to conventional diesel combustionmode The diffusion flame fills the whole combustionchamber, while for the PCCI mode most flame is confined

in the bowl region Late cycle flames are in the bowl regionfor both combustion modes;

Newly defined parameters, such as FSF and FNH,provide further insights into the combustion structure andflame development process

Figure 17 Combustion images Case 7 All times shown in

CAD ASOI in the brackets

Figure 18 Combustion images Case 8 All times shown inCAD ASOI in the brackets

Figure 16 Combustion images Case 6 All times shown in

CAD ASOI in the brackets

ACKNOWLEDGEMENT− This work was supported in part by

the Department of Energy Grant No DE-FC26-05NT42634, by

Department of Energy GATE Centers of Excellence Grant No.

DE-FG26-05NT42622, and by the Ford Motor Company under

University Research Program We also thank Paul Miles of

Sandia National Laboratories, Evangelos Karvounis and Werner

Willems of Ford for their assistance on the design of the optical

engine and on the setup of the experiments.

REFERENCES

Akagawa, H., Miyamoto, T., Harada, A., Sasaki, S.,

Shimazaki, N., Hashizume, T and Tsujimura, K (1999)

Approaches to solve problems of the premixed lean

diesel combustion SAE Paper No. 1999-01-0183

Chen, S K (2000) Simultaneous reduction of NOx and

particulate emissions by using multiple injections in a

small diesel engine SAE Paper No. 2000-01-3084

Choi, G H., Han, S B and Dibble, R W (2004)

Experi-ental study on homogeneous charge compression ignition

engine operation with exhaust gas recirculation Int J.

Automotive Technology 5, 3, 195−200

Fang, T., Coverdill, R E., Lee, C F and White, R A

(2005) Low temperature combustion within a

small-bore high-speed direct-injection (HSDI) diesel engine

SAE Paper No. 2005-01-0919

Fang, T., Coverdill, R E., Lee, C F and White, R A

(2006) Combustion and soot visualization of low

temper-ature combustion within an HSDI diesel engine using

multiple injection strategy SAE Paper No. 2006-01-0078

Fang, T., Coverdill, R E., Lee, C F and White, R A

(2007) Smokeless combustion within a small-bore HSDI

diesel engine using a narrow angle injector SAE Paper

No. 2007-01-0203, SAE Trans.: J Engines, 116, Section

3, 255−270

Fang, T., Coverdill, R E., Lee, C F and White, R A

(2008) Effects of injection angles on combustion

pro-cesses using multiple injection strategies in an HSDI

diesel engine Fuel 87, 15/16, 3232−3239

Han, Z., Uludogan, A., Hampson, G J and Reitz, R D

(1996) Mechanism of soot and NOx emission reduction

using multiple injection in a diesel engine SAE Paper

No 960633

Hasegawa, R and Yanagihara, Y (2003) HCCI combustion

in DI diesel engine SAE Paper No. 2003-01-0745

Hashizume, T., Miyamoto, T., Akagawa, H and Tsujimura,

K (1998) Combustion and emission characteristics of

multiple stage diesel combustion SAE Paper No 980505

Helmantel, A and Denbratt, I (2004) HCCI operation of a

passenger car common rail DI diesel engine with early

injection of conventional diesel fuel SAE Paper No.

2004-01-0935

Kobayashi, S., Sakai, T., Nakahira, T., Komori, M and

Tsujimura, K (1992) Measuremnt of flame temperature

distribution in DI diesel engine with high pressure fuel

injection SAE Paper No 920692

Kook, S and Bae C (2004) Combustion control using

two-stage diesel fuel injection in a single cylinder PCCIengine SAE Paper No 2004-01-0938

Mathews, W S., Coverdill, R E., Lee, C F and White, R

A (2002) Liquid and vapor fuel distributions in a bore high-speed direct-injection diesel engine SAE Paper

small-No. 2002-01-0266

Mathews, W S., Fang, T., Coverdill, R E., Lee, C F andWhite, R A (2003) Combustion visualization of multi-ple injections within an HSDI diesel engine Proc 3rd Joint Meeting of the US Sections of the Combustion Institute, Chicago, IL

Miles, P C (2000) The Influence of swirl on HSDI dieselcombustion at moderate speed and load SAE Paper No.2000-01-1829

Minami, T., Yamaguchi, I., Shintani, M., Tsujimura, K andSuzuki, T (1990) Analysis of fuel spray characteristicsand combustion phenomena under high pressure fuelinjection SAE Paper No. 900438

Najt, P M and Foster, D E (1983) Compression ignitionhomogeneous charge combustion SAE Paper No 830264.Nehmer, D A and Reitz, R D (1994) Measurement ofthe effect of injection rate and split injections on dieselengine soot and NOx emissions SAE Paper No. 940668.Noguchi, M., Tanaka, Y., Tanaka, T and Takeuchi, Y (1979)

A study on gasoline engine combustion by observation

of intermediate reactive products during combustion.SAE Paper No. 790840

Onishi, S., Hong Jo, S., Shoda, K., Do Jo, P and Kato, S.(1979) Active thermo-atmosphere combustion (ATAC) -

A new combustion process for internal combustionengines SAE Paper No. 790501

Su, W., Lin, T and Pei, Y (2003) A compound technologyfor HCCI combustion in a DI diesel engine with earlyinjection of conventional diesel fuel SAE Paper No.2003-01-0741

Swami Nathan, S., Mallikarjuna, J M and Ramesh, A.(2007) Effect of mixture preparation in a diesel HCCIengine using early in-cylinder injection during the suctionstroke Int J Automotive Technology 8, 5, 543−553.Takeda, Y., Keiichi, N and Keiichi, N (1996) Emissioncharacteristics of premixed lean diesel combustion withextremely early staged fuel injection SAE Paper No.961163

Tanaka, T., Ando, A and Ishizaka, K (2002) Study onpilot injection of DI diesel engine using common railinjection system JSAE Review, 23, 297−302

Thring, R H (1989) Homogeneous charge compressionignition (HCCI) engine SAE Paper No. 892068.Tow, T C., Pierpont, D A and Reitz, R D (1994) Reduc-ing particulate and NOx emissions by using multipleinjections in a heady duty DI diesel engine SAE Paper

No 940897

Van Gerpen, J H., Huang, C W and Borman, G L (1985).The effects of swirl and injection parameters on dieselcombustion and heat transfer SAE Paper No. 850265 Zhang, L (1999) A study of pilot injection in a DI dieselengine SAE Paper No. 1999-01-3493

International Journal of Automotive Technology , Vol 10, No 3, pp 297 − 303 (2009)

297

PHYSICAL-CHEMICAL PROPERTIES OF ETHANOL-DIESEL BLEND FUEL AND ITS EFFECT ON THE PERFORMANCE AND EMISSIONS OF

A TURBOCHARGED DIESEL ENGINE

Z.-Q CHEN 1)* , X.-X MA 2) , S.-T YU 1) , Y.-N GUO 3) and J.-S LIU 3)

1)Institute of Automotive Electronic and Technology, Shanghai Jiaotong University, Shanghai 200240, China

2)National Die & Mold CAD Engineering Research Center, Shanghai Jiaotong University, Shanghai 200030, China

3)Institute of Internal Combustion Engine, Jilin University, Changchun 130012, China

(Received 19 October 2005; Revised 15 June 2008)

ABSTRACT− This paper deals with the main physical-chemical properties of diesel blend and the effects of diesel blends (up to 15% volume) on engine performance (full load torque vs engine speed, BSEC vs torque at 1400 r/min and 2300 r/min, and effect of start of injection angle) and emissions in ECE R49 tests (steady 13 points) using a 6.6 L inline 6-cylinder turbocharged direct injection diesel engine The results show that an increase in ethanol fraction results in decreased viscosity of the blend fuel and very high distillation characteristics in the low temperature range Solvents can improve the solubility of ethanol-diesel blends The engine power was degraded proportional to the ethanol content (10% and 15%) due

ethanol-to the LHV (low heating value) of the blends The higher latent heat of vaporization and lower CN (cetane number) of ethanol, which results from the steady state emissions of CO, HC, and SOF (soluble organic fraction), were much higher in the ECE R49 tests at low loads Soot (solid mass) emissions were improved The particulate matter emissions were significantly increased with higher blend volumes, and NOx emissions slightly increased with higher ethanol volumes By increasing the injection angle properly, the performance parameters of the diesel engine were improved, but NOx emissions were deteriorated slightly.

KEY WORDS : Ethanol-diesel blend fuel, Viscosity, Distillation, Solvent, Exhaust emissions

1 INTRODUCTION

The global fuel crises in the 1970s triggered awareness in

many countries of their vulnerability to oil shortages

Considerable attention was focused on the development of

alternative fuel sources Ethanol, a renewable bio-fuel, has

been widely used as a fuel for SI engines, mainly in Brazil,

or as a gasoline additive for improving octane and better

combustion in the USA and Canada (Kremerand Fachetti,

2000; Poulopoulos et al., 2001).In addition, ethanol has a

high oxygen content, which may help to improve combustion

and reduce particulate emissions in CI engines(Hansen et

al., 2005) However, there are many obstacles to using

ethanol in CI engines Ethanol has limited solubility in

diesel fuel, an extremely low CN, a higher latent heat of

vaporization, and a much lower dynamic viscosity than

diesel fuel

To date, several methods of applying ethanol in CI

engines have been introduced, such as ethanol fumigation,

dual injection, ethanol-diesel fuel blends, and ethanol-diesel

fuel emulsion (Qudais et al., 2000; Caro et al., 2001) In

recent years, a number of studies have investigated

ethanol-diesel blend fuels used in CI engines(He et al., 2003; Lü et

al., 2005) In particular, Akzo Nobel Surface Chemistry(Urban Lofvenberg, 2002) and Lubrizol Corporation(Corkwell

additive which makes it possible to blend ethanol withdiesel to obtain a stable and homogenous fuel Ajav et al.

(1999)performed an experimental study of some performanceparameters on a single cylinder diesel engine using ethanol-diesel blends (up to 20% volume) as fuel and found thatbrake specific fuel consumption increased by up to 9%with an increase of ethanol up to 20% in the blends ascompared to diesel alone, and CO and NOx emissions weredecreased Lü et al. (2004) applied a CN improver in anethanol-diesel blend fuel and experimented on a 4-cylinderhigh-speed DI diesel engine and found that NOx, smoke,and HC emissions were improved, CO emissions wereincreased, and the ignition delay was prolonged with theethanol-diesel blend Li et al. (2005)researched the basicphysical-chemical properties of ethanol-diesel blend fuel(cetane number, viscosity, flash point, and surface tension)and the effect on performance and emissions in a single-cylinder air-aspired DI diesel engine and found that smoke,

CO, and NOx emissions were reduced, but HC emissionsincreased significantly

*Corresponding author. e-mail: chenziqiang@sjtu.edu.cn

In this paper, the main physical-chemical properties

(viscosity vs temperature, distillation, and solubility) were

studied experimentally The effects on the performance and

emissions for a diesel engine fueled with different

ethanol-diesel blend fuels were tested using a 6-cylinder turbocharged

direct injection diesel engine following the ECE R49

emission test procedure (steady 13 points) In addition, the

effect of different start of injection angles on the engine

performance was also investigated

2 EXPERIMENTAL APPARATUS AND TEST

PROCEDURE

In this paper, three properties of ethanol-diesel blend were

studied experimentally: viscosity vs temperature, distillation,

and solubility The viscosity test was done on a VAV2000

dynamic viscosity analyzer following the test procedure of

GB265 (Chinese Standard) The distillation test was done

on a distillation analyzer following the test procedure of

GB/T6536 (Chinese Standard) The solubility test was

per-formed using a QHY-1000 Water Bath Constant Temperature

Shaking bench and a refrigerator, where each test point

must be kept for at least 20 minutes to observe whether the

phases in the blends separate For the viscosity and

distillation experiment, the #–35 diesel fuel was used as

the baseline fuel for the ethanol-diesel blend fuel For the

solubility experiment, #–35 diesel fuel and #10 diesel fuel

were used as the baseline fuels

The engine used in the study was a commercial DI,

water-cooled, 6-cylinder, in-line, turbocharged diesel engine

whose major specifications are shown in Table 1 A

HORIBA MEXA-8220D Gas Analyzer was used to measure

the concentration of nitrogen oxide (NOx), unburned total

hydrocarbon (HC), and carbon monoxide (CO) A diffluent

type of sampling system was used for the analysis and

sampling of particulate emissions The filters used for the

collection of particulate matter were clean filters, which

were baked in an oven for 1.5 h at 230oC and then weighed

using a sensitive digital balance of ±0.1 mg accuracy that

was calibrated before weighing The samples were drawn

through the filter for 1 min, and the difference in mass

between these two stages of weighing was the total mass of

the particulate matter In order to determine the mass of the

soot particles, these filters were baked again in the oven at

230oC for 1.5 h to remove moisture and the volatile matters,

and they were then re-weighed The difference in massbetween these filters and the clean ones was the mass of thesoot particles

The emission tests were performed following the testprocedure of ECE R49 The effects of the start of injectionangle on the engine performance and emissions weredetermined at the maximum torque point (100% load, 1400r/min), and three injection advance angles (6oCA, 9oCAand 12oCA) were tested Commercial #–35 diesel fuel andanalysis-grade anhydrous ethanol (99.7% purity) were used

in this engine test For each ethanol–diesel blend fuel (E10,E15), 1% (by volume) of solvent was added to the blends

to improve the solubility The outside environment ature was −20oC, and the indoor temperature was 10oC

temper-3 BLEND FUEL PROPERTIES3.1 Heating Value of Blend FuelEthanol has a much lower LHV compared with diesel fuel.The LHV of an ethanol-diesel blend can be determined byformula (1) (Cui, 1990):

3.2 Cetane Number of Blend FuelThe cetane number (CN) is an important fuel property fordiesel engines It influences the engine start ability, emissions,peak cylinder pressure, and combustion noise A high CNensures good cold starting ability, low noise, and long enginelife

Table 1 Diesel engine specifications

Rated power (kW)/speed (r/min) 155/2300

Maximum torque (N.m)/speed (r/min) 725/1400

Injection advance angle (oCA) (9±1)

Nozzle number x Orifice diameter (mm) 6×0.24 Figure 1 Heating value percent and Cetane number ofblend fuel versus ethanol fraction.

PHYSICAL-CHEMICAL PROPERTIES OF ETHANOL-DIESEL BLEND FUEL AND ITS EFFECT 299

The CN of ethanol (8) is much lower than that of diesel

fuel The CN of blend fuel can be calculated by formula (2)

(Cui, 1990):

where:

CN, CN1, and CN2 are the CN of the blends, the diesel

fuel, and the ethanol, respectively, and a and b are the

volume percent of diesel fuel and ethanol in the blends,

respectively (%)

The CN of the blend is shown in Figure 1, using a CN of

50 for diesel fuel The CN of the blend decreases linearly

with the increase in ethanol content, which may influence

the combustion performance of the blend fuel in the diesel

engine, especially the period of ignition delay Thus, it is

necessary to add the ignition additive to the blends with a

higher ethanol content (Lü et al., 2004; Caro et al., 2001)

3.3 Viscosity of Blend Fuel

Because the fuel also functions as a lubricant and to

prevent leakage in the diesel engine, the viscosity is an

important evaluation norm for the diesel fuel The relation

between viscosity and temperature (viscosity curve) is an

important property of fuel, especially for alternative fuel

The viscosity test was performed in the temperature range

of 20~100oC, and the experimental results are shown in

Figure 2

The viscosity decreases with increasing ethanol content

in the blends This is the reason that the boiling point of

ethanol is 78.5oC, which is much lower than that of diesel

fuel; when the temperature of the blend fuel is higher than

the boiling point of ethanol, the ethanol is vaporized before

the blend fuel When the temperature increases to 100oC,

the experimental results show that nearly all the ethanol is

vaporized from the blend Thus, it is necessary to prevent

the occurrence of air block and cavitation in the fuel pipe

for diesel engines fueled with ethanol-diesel blends In

addition, the decrease in the viscosity of the blend would

result in increased leakage in the plunger pump and the

reduction of the volume of fuel delivered

3.4 Distillation of Blend FuelDistillation is the most essential property for fuel Thedistillation curve (distillation vs temperature) represents thepercent of light, medium, and heavy distillation components

of a fuel The percent of each component will influence thetiming and intensity of the ignition and combustion process.The distillation characteristics of different blends (E0, E10,E20, E30) are shown in Figure 3

Figure 3 shows that the distillation fraction increasesrapidly with increasing temperature A large differenceexists between the distillation curves when the distillationfraction is lower than 50% This is also a reason that theethanol has a lower vaporization temperature and vaporizesfirst from the blend fuel when it is in the high temperaturerange Overall, the strong vaporization of blend fuels in thelow temperature range could influence the combustion

Figure 2 Comparison of viscosity versus temperature of

blend fuels

Figure 3 Distillation of blend fuels

Figure 4 Solubility of blend fuel

characteristics and the storage security of the blend fuel

3.5 Solubility of Blend Fuel

Ethanol has limited solubility in diesel fuel However,

solvents can improve the solubility of ethanol-diesel blends

(Corkwell et al., 2002) Here, two types of diesel fuel

(#−35 diesel fuel and #10 diesel fuel) were used as the

baseline fuels for the different ethanol-diesel blends In

addition, 1% solvent (by volume) was also added to each

ethanol-diesel blend

Figure 4 shows the solubility curves The region above

the solubility curve is the uniform solution of

ethanol-diesel blend, while the region below the solubility curve is

the phase separation When 1% solvent by volume was

added to each ethanol-diesel blend, the temperature of the

phase separation of each blend was decreased by 3~13oC

Using the solvent in the blends is thus an effective way to

solve the problem of solubility of ethanol-diesel blend In

addition, the blends with small or large ethanol addition are

feasible and valuable

4 RESULTS AND ANALYSIS OF ENGINE

TESTS

4.1 Effects on the Power

Without any modification of engine parameters, the effects

on the power of the diesel engine were investigated Figure

5 shows that the torque decreases greatly with increasing

ethanol content in the blend At the maximum torque, the

torque output of the diesel engine fueled with E10 and E15

fuel decreased by 5% and 10%, respectively, compared

with that of diesel fuel This is due to the LHV of the

ethanol-diesel blends The LHV of E10 fuel is 96.4% that

of diesel fuel, while the LHV of E15 is 94.6% that of diesel

fuel Therefore, under the same volume of fuel delivered,

the torque output of the diesel engine must be decreased

4.2 Effects on the Fuel Economy

Because there is a large difference in the LHV between

ethanol and diesel fuel, the brake specific fuel consumption

(BSFC) is not a proper norm to evaluate the fuel economy

of different blend fuels Instead, the brake specific energyconsumption (BSEC) {defined as BSFC×(Hu)blend fuel}

is used to evaluate the fuel economy of different blendfuels

Figures 6(a) and (b) show the BSEC vs engine load fordifferent ethanol–diesel blends and pure diesel fuel at theFigure 5 Effect on the power of the diesel engine

Figure 6 Effects on the fuel economy

Table 2 ECE R49 test procedure

Condition number Engine speed(r/min) Load percent (%) Test time (min)

PHYSICAL-CHEMICAL PROPERTIES OF ETHANOL-DIESEL BLEND FUEL AND ITS EFFECT 301

load characteristic of 1400 r/min and 2300 r/min The

BSECs of the blend fuels are nearly equal to that of diesel

fuel at the high load condition, and they deteriorate gradually

with increasing ethanol content at medium and low load

conditions, especially at the low load of 1400 r/min This is

the reason that the combustion temperature is relatively

lower, and the air/fuel ratio is relatively higher, at medium

and low load conditions Ethanol has a higher latent heat of

vaporization and lower CN, all of which lead to the bad

combustion characteristics of blend fuels in these conditions

With increasing engine load, the temperature in the cylinder

rises, the air/fuel ratio decreases, and the burning of theblend fuel is improved, so the BSEC of the blend fuel isnearly equal to that of diesel fuel at high load conditions.4.3 Effects on the Emissions

4.3.1 Emission test of ECER49Without any modification of engine parameters, the effects

of different ethanol-diesel blend fuels (E10, E15) on theemissions of the turbo-charged diesel engine were evaluatedfollowing the ECE R49 test procedure (shown in Table 2).The experimental results, including the brake specific hydro-

Table 3 Comparison of brake specific emissions of different blend fuels

Fuel\Brake specific emission

carbon mass (BSHC), the brake specific CO mass (BSCO),

the brake specific oxides of nitrogen mass (BSNOx), the

brake specific particulate mass (BSPM), the brake specific

solid mass (BSSOOT), and the brake specific soluble

organic fraction mass (BSSOF), are shown in Table 3 and

Figure 7(a)~(e)

The increase in the CO level (shown in Table 3) with

increasing ethanol fraction is a result of incomplete

com-bustion of the ethanol-air mixture at low loads (Figure

7(a)) The factor causing combustion deterioration at the

low loads is the lower combustion temperature; in addition,

the higher latent heat of vaporization of ethanol and large

distillation in the low temperature range (Figure 3) would

help to reduce the combustion temperature and result in the

reduction of the CO oxidation reaction rate Thus, a

thickened quench layer created by the cooling effect of

vaporizing ethanol could have played a major role in the

increased CO production at the low loads However, the

combustion temperature increases gradually with the

increase of engine load, which weakens the cooling effect

of vaporizing ethanol, and the higher oxygen content of the

blend fuels would further help to increase the

oxygen-to-fuel ratio in the oxygen-to-fuel-rich regions, all of which would speed

up the CO oxidation reaction rate Thus, the CO emissions

decrease with increasing ethanol fraction at high loads

The increase in the hydrocarbon emissions with the

increase of ethanol fraction (Table 3) is the result of the

ethanol having a lower CN, which deteriorates the

com-bustion performance of the blend fuel Especially at the

low load conditions of the diesel engine (condition No 2,

3, and 12 in Figure 7(b)), the combustion temperatures are

much lower, and the air/fuel ratios are relatively larger, all

of which form the poor combustion conditions of the blend

fuel and deteriorate the BSEC of the blend fuel and the

hydrocarbon emissions Because the main component of

the soluble organic fraction (SOF) is unburned hydrocarbon,the increase in the HC levels results in a high BSSOF(Table 3, Figure 7(c))

The decrease in the BSSOOT level with an increase ofethanol fraction (Table 3, Figure 7(d)) is the result of thehigher oxygen content of the blend fuels, which would help

to increase the oxygen-to-fuel ratio in the fuel-rich regionsand speed up the oxidation reaction of dry smoke It isfound (Table 3) that the increased BSSOF level is higherthan the reduced BSSOOT level, so the particulate emissionsincrease with increasing ethanol fraction in the blend fuel.The NOx emission level increases with an increase ofethanol fraction (Table 3) Figure 7(e) shows that the NOxemission level increases greatly with the increasing ethanolcontent in the blend fuel at the high load conditions This isthe reason that the combustion temperature at the high load

is higher than that at low load; in addition, the blend fuelshave higher oxygen content than the diesel fuel, whichresults in increased NOx production

4.3.2 Effect of injection advance angle

By adjusting the injection advance angle, the effects of thedifferent ethanol-diesel blend fuels (E10, E15) on the per-formance and emissions of the turbo-charged diesel enginewere evaluated experimentally at the maximum torque point(100% load, 1400 r/min) Three injection advance angles(6oCA, 9oCA, and 12oCA) were tested Figure 8 shows thatthe power, BSEC, and the emissions of CO and smoke (Rb)improved with the increase of injection advance angle forthe two blend fuels The HC emissions increase slightly,while the NOx emissions greatly deteriorate This is thereason that the blend fuel has a longer ignition delay timecompared with the diesel fuel (Lü et al., 2004) By increasingthe injection advance timing properly, the premix burningcan be increased compared with that of 9oCA This not only

Figure 8 Effect of different injection advance angle on the performance of the diesel engine

PHYSICAL-CHEMICAL PROPERTIES OF ETHANOL-DIESEL BLEND FUEL AND ITS EFFECT 303

results in the improvement of most of the performance

parameters of the engine but also increases the maximum

burning temperature and results in increased NOx emissions

5 CONCLUSIONS

The following conclusions can be drawn from this study:

(1) The LHV, CN, and viscosity of blend fuels decrease

with increasing ethanol fraction in the blend fuel The

blend fuel has a strong vaporization in the low

temperature range Solvents can improve the solubility

of ethanol-diesel blends

(2) The experimental results show that the LHV results in

the reduction of power of the diesel engine fueled with

ethanol-diesel blend fuel The higher latent heat of

vaporization of ethanol is the main factor that results in

the deterioration of CO emission levels of blend fuels

at low loads The lower CN of ethanol leads to the

deterioration of HC emission levels, SOF emission

levels, particulate emission level, and the BSEC of

blend fuels at low loads The higher oxygen content of

ethanol helps reduce the soot emission level of the

blend and decreases the CO emission level of blend

fuel at high loads, while also increasing NOx emissions

of the blend fuels at high load conditions

(3) With increasing the injection advance angle, the burning

characteristics of ethanol-diesel blend fuels were

improved, and most performance parameters, except

the NOx emissions, were improved

REFERENCES

Ajav, E A., Singh, B and Bhattacharya, T K (1999)

Experimental study of some performance parameters of

a constant speed stationary diesel engine using

ethanol-diesel blends as fuel Biomass and Bioenergy, 17, 357−

365

Caro, P S., Mouloungui, Z., Vaitilingom, G and Berge, J

C (2001) Interest of combining an additive with

diesel-ethanol blend diesel engines Fuel, 80,565−574

Corkwell, K., Akarapanjavit, N., Srithammavong, P.,Schuetzle, D and Han, W (2002) The development ofdiesel/ethanol fuel blends for diesel vehicles: Fuel for-mulation and prosperities The 14th Int Symp Alcohol Fuels (ISAF XIV), Phuket, Thailand

Cui, X C (1990) The Alternative Fuel for Internal bustion Engine Machinery Publisher of China 44−60.Hansen, A C., Zhang, Q and Lyne, P W L (2005) Ethanol-diesel fuel blends−A review Bioresource Technology,

auto-Li, D G., Huang, Z., L , X C., Zhang, W G and Yang, J

G (2005) Phisico-chemical properties of ethanol-dieselblend fuel and its effect on performance and emissions

of diesel engines Renewable Energy, 30,967−976

Lü, X C., Huang, Z., Zhang, W G and Li, D G (2005).Combustion visualization and emissions of a directinjection compression ignition engine fueled with bio-diesohol Int J Automotive Technology 6, 1, 15−21

Lü, X C., Yang, J G., Zhang, W G and Huang, Z (2004).Effect of cetane number improver on heat release rateand emissions of high speed diesel engine fueled withethanol-diesel blend fuel Fuel, 83, 2013−2020

Poulopoulos, S G., Samaras, D P and Philippopoulos, C

J (2001) Regulated and unregulated emissions from aninternal combustion engine operating on ethanol-contain-ing fuels Atmospheric Environment,35, 4399−4406.Qudais, M A., Haddad, O and Qudaisat, M (2000) Theeffect of alcohol fumigation on diesel engine performanceand emissions Energy Conversion & Management, 41,

389−399

Urban Lofvenberg (2002) E-diesel in Europe: A new lable fuel technology The 14th Int Symp Alcohol Fuels (ISAF XIV), Phuket, Thailand

avai-u ê

CHARACTERIZATION OF THE HVAC PERFORMANCE WITH DEFROSTER GRILLERS AND INSTRUMENT PANEL REGISTERS

M F KADER, Y M YOUN, Y D JUN and K B LEE *

Department of Mechanical Engineering, Kongju National University, Chungnam 330-717, Korea

(Received 19 November 2007; Revised 18 November 2008)

ABSTRACT− Improving HVAC performance is of paramount importance to sustain passenger comfort in a car Numerical analysis of a three-dimensional model predicts a detailed description of fluid flow and temperature distribution in the passenger compartment and on the inside surface of the windshield During the winter, the highest temperature is observed on the lower part of the windshield and in the vicinity of the defroster nozzle Defrosting performance was excellent when the injection angle of the defrost nozzle was between 15 to 25 degrees with a standard distance ratio of one During the cooling period, it was found that the temperature and humidity decrease to a comfortable range almost linearly at the initial stage The numerical predictions are in good agreement with the experimental results

KEY WORDS : CFD, Instrument panel (IP), Defroster nozzle, Windshield, HVAC system

1 INTRODUCTION

Computational Fluid Dynamics (CFD) is widely used in

various studies in the field of automobile Heating,

Venti-lating, and Air-Conditioning (HVAC) systems In the summer,

the temperature of the compartment of a parked automobile

increases to an extremely high value under sunny

condi-tions The cooling period during vehicle start up requires

time to reach steady-state conditions During this period,

the conditions are highly non-uniform Another important

aspect is the need for adequate visibility throughout the

windshield, particularly at very low temperatures when ice

usually forms on the windshield Hence, characterization of

HVAC performance is essential to achieve a comfortable

automobile environment and to improve the capacity of

windshield defrost systems by melting ice immediately and

completely from the windshield outer surface

The substantial advancement in the field of CFD has

encouraged a number of researchers to investigate HVAC

performance Numerical simulations of two- and a

three-dimensional airflow in a passenger compartment were

per-formed by Hara et al. (1988) CFD was used to analyze the

effect of four HVAC design parameters on passenger

thermal comfort in a simplified passenger compartment

(Lin et al., 1992) They found that the location of the vents

and the air flow rate had the most influence on passenger

thermal comfort The position of the outlet in the rear of the

car was equally important for the thermal comfort of rear

passengers A study by Ishihara et al. (1991) examined the

airflow inside a one-fourth scale three-dimensional model

A fluidic oscillator device by Stouffer and Sharkitt (1987)was developed to improve the airflow distributions over thewindshield The device was also used as a windshield defrost/defog nozzle with some degree of success Lee et al. (1994)utilized a CFD code, namely ICEM-CFD, to simulate themechanism of windshield de-icing The complete vehicleconfiguration was transformed from CAD and the meshwas created and assembled using a multi-domain approach.The authors demonstrated the capability of the developedmodule by simulating cold room de-icing tests to supple-ment the experimental work Recently, Brewster et al. (1997)used the CFD code STAR-CD to simulate the mechanism

of ice building on the windshield in three-dimensionalform The authors used a non-linear enthalpy-temperaturerelationship to simulate the ice/water layer Melting contourswere predicted every 5 minutes Aroussi and Aghile (2000)used a one-fifth scaled Perspex model of a passengercompartment for experiments with the Particle ImageVelocimetry (PIV) technique A further study by Aroussi et

al. (2003) simulated turbulent fluid flow over and heattransfer through a model of vehicle windshield defrostingand demisting system Furthermore, Park et al. (2006)simulated the flow and temperature field on the interior of

an automobile cabin when the hot air is discharged fromthe defrost nozzle to melt the frost on the windshield glass.Lee et al. (2006, 2007) focused on the temperature dis-tribution characteristics of an automobile interior, bothnumerically and experimentally when operating an HVACsystem in the summer

In the present study, the flow field, temperature andhumidity distribution within a 3D model of a vehicle com-partment and the melting pattern of frost are investigated

*Corresponding author. e-mail: kumbae@kongju.ac.kr