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

Copyright © 2009 KSAE
1229−9138/2009/046−01

International Journal of Automotive Technology, Vol. 10, No. 3, pp. 265−276 (2009)

DOI 10.1007/s12239−009−0031−3

EVALUATION OF LOW-TEMPERATURE DIESEL COMBUSTION REGIMES
WITH n-HEPTANE FUEL IN A CONSTANT-VOLUME CHAMBER
U. B. AZIMOV , K. S. KIM , D. S. JEONG and Y. G. LEE
1)

1)*

2)

2)

Department of Mechanical Design Engineering, Chonnam National University, Cheonnam 550-749, Korea
Korea Institute of Machinery and Materials, Eco-Machinery Research Division, 171 Jang-dong,
Yuseong-gu, Daejeon 305-343, Korea

1)

2)

(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 O2 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 O2 concentrations, soot and NOx emissions were simultaneously reduced.

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

1. INTRODUCTION

rapidly at low cost. However, the chemical kinetics involved in this concept do not allow the use of classical diesel
auto-ignition and combustion models based on oversimplified representations of combustion chemistry. In these
classical models, the time in the reactive zone is usually
considered much smaller than the diffusion time of fuel and
air towards the flame region.
This work analyzes the effect of various parameters such
as the ambient gas O concentration, ambient gas temperature, and fuel injection pressure on the evolution of diesel
flames and emission formation in low-temperature combustion regimes. The results are obtained by simulating
LTC conditions with the n-Heptane fuel and ECMF3Z
model used by the STAR-CD code. First, the features of
various combustion models are compared with respect to
LTC. Then, the simulation results are presented in comparison with available experimental data.

International regulations ratified in recent years have imposed 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 compression ignition (HCCI) and conventional diesel-based
Low-Temperature Combustion (LTC) concepts show great
potential in reducing NOx and soot emissions simultaneously. 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 modifications. However, the differences in chemistry and combustion between this concept and conventional diesel combustion 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

2

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

*Corresponding author. e-mail:
265

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 combustion 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, nonsooting 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 improved. Although liquid fuel penetrates much further into the
chamber, the higher energy released from premixed reactions contributes to the intense evaporation of liquid fuel
(Higgins
. 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 turbulence 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 finiterate chemistry is important for accurately predicting pollutant formation. To improve the accuracy of predictions
while modeling diesel spray combustion, unified combustion models have been built to account for all types combustion modes simultaneously.
Abraham
. (1985) suggested replacing the controlling time scale in the Magnussen model by the slowest time
scale of the mixing time and the chemical time. Kong
.
et al

et al

et al

Figure 1. Conceptual scheme of low-temperature combustion.

(1995) proposed an extended characteristic-time model
based on Abraham
. (1985) which accounts for chemical and turbulence time scales simultaneously. This model
was combined with the Shell ignition model to simulate the
overall combustion processes in a diesel engine. In this
combined model, the initiation of combustion relies on
laminar chemistry, and turbulence starts to have an influence on combustion only after combustion events have
already been observed, similar to the Magnussen model.
Even if premixed and non-premixed combustions are taken
into account in this model, the non-mixture of the species
within a computational cell is only represented by the
mixing time-scale, which does not account for the mixing
history. Consequently, the transition between chemically
controlled and mixing-controlled combustion needs to be
monitored by an empirical function. This model does not
account for flame propagation combustion. These modifications improve the eddy break-up model only to a minor
extent 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
. (2000). Based on the classical flamelet
model by Peters 1986, this model (in which the reactions
occur in wrinkled turbulent flames that can be considered
as a collection of laminar flamelets) suggests that each cell
et al

et al


EVALUATION OF LOW-TEMPERATURE DIESEL COMBUSTION REGIMES WITH n-HEPTANE FUEL

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 Flamelet Model) was developed by Hasse
. (2000). This
model is an extension of the RIF (Representative Interactive Flamelet) concept by Pitsch
(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 equations 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 introduction 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 unacceptable 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 dimension of progress of reaction, represented by the mean proet al

et al.

267

gress variable and its fluctuation (Pope, 1988; Candel and
Poinsot, 1990; Bray ., 2005). This approach is based on
the Coherent Flame Model (CFM), which describes the

rate of fuel consumption per unit volume as the product of
the flame surface density (i.e. the flame surface per unit
volume) and the local flame speed at which it consumes the
mixture. This approach supposes that the chemical reaction
of fuel oxidation occurs in a very thin layer. This layer
separates the burned and unburned gases and propagates
toward the fresh mixture of fuel, oxygen and dilutant. The
CFM model was extended first to the ECFM model, and
was specifically adapted to model combustion with perfectly or partially mixed mixtures and to simulate the combustion processes in direct injection spark ignition engines
(Colin
., 2003). Then ECFM was adapted to account
for unmixed or diffusion combustion, and the three-zone
description of the mixing state was added. This new combustion model, called ECFM3Z (3-Zones Extended Coherent
Flame Model), can therefore be seen as a simplified CMCtype model, which discretizes the mixture fraction space by
only three points. Therefore, this model was selected to be
the most appropriate for simulating LTC regimes because it
can reflect the real physics of LTC, it relies on flamelet
libraries, and it is less computationally demanding than the
CMC and EPFM models.
et al

et al

3. ECFM3Z MODEL CONCEPT

The ECFM3Z model was briefly presented by Beard
.
(2003) and was described in detail by Colin and Benkenida
(2004). This model describes the unburned/burned gas
zones based on the flame surface density equation. In order

to account for diffusion flames and mixing processes, each
computational cell is split into three mixing zones: a pure
or unmixed fuel zone, a pure air plus EGR zone or unmixed
air and EGR zone, and a mixed zone, containing fuel, air
and EGR. This structure can account for the three main
combustion modes: auto-ignition, flame propagation and
diffusion flame as encountered in LTC. It is based on two
dimensions: the mixing state description and the reaction
progress description. The mixing state description is represented by the Probability Density Function (PDF) of the
mixture fraction
P( Z )=a δ ( Z )+bδ ( Z – ZM )+cδ ( Z – 1 )
(1)
M
where Z is the average value of the mixture fraction in
the mixed zone. The first δ function corresponds to the
unmixed air region, the second one to the mixed region and
the third one to the unmixed fuel region. In this structure,
space is discretized by only three points. The mixing
model can reflect the transference of unmixed fuel and
unmixed air into the mixed region. The reaction progress
description is represented by the progress variable
et al

Z

Y˜ uFu
m -u =1− --------c˜ =1− ----m
Y˜ TFu

(2)

268

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

where Y˜ uFu is the mass fraction of the fuel present in the
unburned gases, and Y˜ TFu is the mass fraction of fuel before
the onset of combustion (fuel tracer). Y˜ TFu is constant in
space and time for perfectly mixed charges. In practical
applications Y˜ TFu 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 densities 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 .,
2005; Reveille
., 2006; Knop and Jay, 2006; Priesching
., 2007; Shi
., 2007.
et al

et al

et al

et al

4. SIMULATION PROCEDURE

The simulation was conducted using the STAR-CD commercial 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 addition, 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 nonreacting and reacting fuel jets.
The fuel was injected with spray characteristics adjusted
according to the spray characteristics assumed in the experiments. 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 determined. This angle was determined from experiments performed 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 ., 1998; Subramanian,
2007). The tables give values for the two delays and these
delays are functions of pressure, temperature, equivalence
ratio and EGR. For emission simulation, the 3-step Zeldovich
model and ERC model were used for NOx and soot emission
calculations, respectively. The simulation conditions are
listed in Tables 1 and 2. The ambient gas temperature,
ambient gas content, ambient gas pressure, fuel injection
pressure, injection duration and single-hole injector orifice
parameters correspond to those of the experiment.
et al

5. RESULTS AND DISCUSSION
5.1. Evaluation Approach for LTC Regimes
The present paper numerically evaluates the LTC regime of
DI diesel combustion. Since LTC differs from conventional
diesel combustion, it is necessary to use a model that is
universally applicable to both conventional and alternative
diesel combustion applications. Soot and NOx emissions
computed for particular conditions, and the entire combustion event was evaluated to understand the physics of the
combustion as well as the relations among the operating
parameters. For this purpose, the parameter called “Combustion Factor” was introduced:
Y˜ Fb
(3)
ψ =1− --------Y˜ TFu
where, Y˜ Fb is the mass fraction of fuel in the burnt gases,

and Y˜ TFu is the fuel tracer
This parameter is considered as an indicator of the
combustion mode (premixed vs. diffusion) in the complex
LTC process. It is extremely difficult to differentiate the
diesel LTC process into certain modes because premixed,
partially premixed and diffusion modes occur simultaneously. However, it might be possible to map the combustion event and see which combustion mode prevails and
Table 1. Simulation conditions.
Fuel
Ambient gas temperature [K]
Ambient gas pressure [MPa]
Rail pressure [MPa]
Injection duration [ms]
Injected fuel mass
Nozzle hole

n-Heptane
820, 870, 920
4
90, 135
1.2
8.4 mg, 10.3 mg
dn=0.163 mm, ln/dn=5.52

Table 2. Ambient gas content.
Molecular percentage
N2
CO2
O2
21.0
79.0

0
16.0
81.0
3.0
12.0
81.0
7.0

MW
28.84
29.12
29.6


EVALUATION OF LOW-TEMPERATURE DIESEL COMBUSTION REGIMES WITH n-HEPTANE FUEL
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 considered 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 equations 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:

ψ – ψ minψ n=------------------1–ψ
min

(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 autoignition. Also, it is seen that with the decrease of O2 concentration 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 ketohydroperoxide 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 autoignition delay at 820 K is noticeably different, probably
due to longer time required for the air/fuel mixture to reach

269

Figure 2. Autoignition delay mapping.
the decomposition temperature. In addition, probably at
820 K, there was no significant decomposition of fuel during evaporation, but some portion of fuel did decompose
during evaporation for the high temperature case. This
assumption is in agreement with other presented work
(Curran
., 1998; Wang and Rutland, 2005).
After the first autoignition, there is a period of very
slight temperature increase due to “cool flame” chemistry.
This period is an ignition delay between the initial fuel
decomposition and very rapid temperature rise. The time
interval between the first autoignition and the main autoignition is much greater at 820 K for similar O2 conditions
compared to those of the other two cases. This difference is
due to both the retarded first autoignition and the further
retarded reaction progress with the decrease of O2 concentration. In the ECFM3Z model, defining the occurrences of
the first and main autoignitions is straightforward because
their computed values are automatically stored in the postprocessing file. Figure 3 compares spray combustion simulation results for various O2 concentrations and ambient gas
temperatures with experimental data.
The results presented in Figure 3 indicate good agreement with data from the experiments in terms of spray and
combustion development. The liquid fuel pattern, as well

as the spatial distribution of the flame, matches well the
pattern obtained from experiments.
Figure 3 does not provide any information on the start of
autoignition, but only depicts the comparison in spray
flame development between simulation and experiment.
Nevertheless, based on Figures 2 and 3, the effect of variation of charge composition showing the longer autoignition delay periods and decreased flame temperatures was
evident for the diluted charge at lower O2 concentration.
et al

5.3. Flame Temperature
The ambient gas temperature has a small influence on
flame temperature and NOx formation but has great effect
on the fuel/air equivalence ratio. The ambient gas temper-


270

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

Figure 3. Spray combustion development at P -90 MPa.
inj

ature can only have an effect on autoignition initiation
because of the reaction of fuel with oxygen in the hightemperature environment. And then, as the flame propagates, 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 combustion 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 temperature is slightly increased by an increase in injection

pressure, especially during the initial stage of flame propagation and temperature increase. This may be explained by
better evaporation and mixing at higher injection pressures
(Gill
., 2005). As the fuel is well-mixed and distributed,
higher energy is released, and flame temperature is increased.
et al

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 O212%, 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 flame
temperature discussed previously, is indicated. A similar
trend of increased NOx formation with higher injection
pressures was shown by Henein
., 2006.
et al

5.5. Equivalence Ratio
Akihama
. (2001) have shown that in addition to the

notable decrease in NOx emissions, soot formation can also
be avoided by producing combustion at flame temperatures
less than 1800 K. Kamimoto and Bae (1988) proposed that
soot formation could be avoided by producing combustion
at equivalence ratios below 2. Non-sooting combustion has
also been demonstrated at higher temperatures and higher
equivalence ratios by entraining sufficient oxygen into the
jet. NOx formation occurs at high temperatures, but NOx is
reduced to N2 under fuel-rich conditions and thus, NOx
emissions are decreased with higher equivalence ratios.
In the case of low-temperature combustion, soot emission
reduction appears to be related to an increase in ignition
delay, which is due to the reduced O2 concentration, which
provides more time for mixing before combustion and a
possible decrease in the equivalence ratio of the igniting
fuel-ambient gas mixture.
Figure 6 shows equivalence ratio distribution at the time
of main ignition for each combustion regime. The deeper
the fuel penetrates into the chamber, the more diluted and
mixed it becomes as O2 concentration and ambient temperature decrease. The equivalence ratio at the jet's leading
part, the zone where ignition is supposed to occur and where
the premixed burn occurs, is about 2.
The maximum equivalence ratio value corresponds to
et al


EVALUATION OF LOW-TEMPERATURE DIESEL COMBUSTION REGIMES WITH n-HEPTANE FUEL

Figure 4. Flame temperature variation with the change of
O2 concentration, ambient gas temperature and injection

pressure.
the core of the injected fuel and is about 6, and the stoichiometric values are along the jet periphery, which is in agreement with the experimental 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 because 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 interpreted as a qualitative indicator of in-cylinder soot formation
(Siebers
, 2002; Mueller and Martin, 2002; Choi
.,
2004; Kim
., 2007).
et al.

et al

et al

271

Figure 5. NOx variation with the change of O2 concentration, ambient gas temperature and injection pressure.
In Figure 7(a) and (b), both the soot curves and the flame
luminosity curves have a similar pattern. After soot
formation and flame luminosity reaches their peaks, soot
oxidation will dominate and flame luminosity will decrease.
The soot oxidation process is clearly seen to be slower than

the soot formation process in both (a) and (b). A similar
trend was observed for other cases for different O2 concentrations and ambient gas temperatures, as in Figures 8(a),
(b) and 9(a), (b). In general, higher flame luminosity
peaks were found for lower injection pressures. Higher
luminosity peaks with higher injection pressure for the case
of O2-21% and ambient temperature 920 K in Figure 7 are
probably due to a higher local equivalence ratio at higher
ambient gas temperatures.
Although a higher injection pressure is believed to contribute to better mixing (Pickett and Siebers, 2004b), there
is probably still not enough time to ensure sufficient


272

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

Figure 6. Equivalence ratio at the time of ignition at Pinj-90
MPa.
Figure 8. Comparison of calculated soot and measured
flame luminosity at Tamb-870 K.

Figure 7 Comparison of calculated soot and measured
flame luminosity at Tamb-920 K.
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

Figure 9. Comparison of calculated soot and measured
flame luminosity at Tamb-820 K.
combustion factor concept described earlier, the LTC
regimes are evaluated to determine the best scenario of
combustion in terms of soot and NOx emissions. Note that


EVALUATION OF LOW-TEMPERATURE DIESEL COMBUSTION REGIMES WITH n-HEPTANE FUEL

Figure 10. Combustion factor at Tamb-920 K.

273

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

Figure 11. Combustion factor at Tamb-870 K.
Figure 14. Pressure rise comparison for different simulation conditions.

Figure 12. Combustion factor 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 fuellean mixtures (Kamimoto and Bae, 1988; Dec, 1997;

Kimura
., 1999; Akihama
., 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
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et al

Figure 15. Apparent heat release rate for different simulation conditions.
combustion factor, the quantity of fuel present in the
burned gases and the values of the equivalence ratio at the
instances when the soot formation is at its maximum were
determined. For brevity, only the case with ambient temperature of 870 K is presented in Figure 13. As mentioned
earlier, the fuel in the burned gases is consumed by the
diffusion flame. This fact can be utilized to determine the


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
.,
2003; Kook
., 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
., 1994).
The results above can be summarized to support the
argument that the diffusion mode can have a great influence 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 important 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 . (2001), showed that soot formation can be suppressed at temperatures
below 1700 K. Theoretical analysis of reaction rates performed by Jacobs and Assanis (2007), suggested that soot
formation is insensitive to equivalence ratio at temperatures below 1500 K. At such low temperatures, 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
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ignition/combustion modes makes the ECFM3Z model
universally and specifically applicable for accurately simulating conventional as well as alternative diesel combustion
regimes.
6. CONCLUSION

In this paper, a description of a new approach to evaluate
LTC regimes using the ECFM3Z model was presented.
The computed results, which were compared with available
experimental data, showed that this model was able to
accurately predict autoignition and combustion for both
conventional and LTC regimes. The specific major findings
are summarized as follows:
(1) The double-delay autoignition model, as a part of the
ECFM3Z model, was shown to be a good predictor of
autoignition delay for the entire range of conditions in
this paper. Computed results were compared with experimental data and good agreement was observed.
(2) Spray combustion evolution at various levels of O2
concentration and ambient temperature matched very
well with the experimental data, for spatial as well as
temporal jet flame development.
(3) Ambient gas temperature and fuel injection pressure
had a minor effect on flame temperature. Flame temperature values for different conditions were almost the
same when ambient gas temperature increased by 50K.
Flame temperature increased a little at higher injection
pressures. However, oxygen concentration had a great
effect and flame temperature considerably decreased
with decreasing O2 concentration in the ambient gas.
(4) Consequently, NOx level greatly decreased, for it depended only on flame temperature. For all the conditions with 12% oxygen concentration, NOx level decreased to almost zero because the flame temperature
was about 1900K. With higher injection pressure, NOx
level slightly increased because of the slight increase of
temperature.
(5) Equivalence ratio at the time of ignition stabilized at
around 2, as the fuel jet penetrated further into the
combustion chamber. However, for the conditions O221%, 920 K and O2-21%, the 870 K equivalence ratio
was higher because at normal oxygen concentrations

and higher ambient gas temperatures, the ignition delay
period is very short and there is not sufficient time for
the injected fuel to mix with the ambient gas and to
dilute before the start of combustion.
(6) Computed soot data were compared with experimental
data of flame luminosity. Similar patterns and trends in
soot formation were observed for all conditions. Soot
level decreased with decreasing O2 concentration and
increasing fuel injection pressure. However, at 21%
oxygen and 920K ambient gas temperature, the soot
level was higher at higher fuel injection pressures. This
may be explained by the insufficient fuel-ambient gas
mixing time, as well as the higher flame temperature.


EVALUATION OF LOW-TEMPERATURE DIESEL COMBUSTION REGIMES WITH n-HEPTANE FUEL

(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 concentration, and therefore, the combustion factor leaned
towards the premixed mode. This means that with the
decrease of O2 concentration, mixing of fuel with ambient gas improved and the fuel was mainly consumed
in the premixed zone. Furthermore, the average equivalence ratio at the time steps with the highest soot
formation level increased with decreasing O2 concentration. Heat release peaks were higher for the cases
with lower O2 concentrations and ambient gas temperatures, 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 geometry of a diesel combustion chamber, high ambient
gas temperatures and injection pressures, various ambient 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.

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Copyright © 2009 KSAE
1229−9138/2009/046−02

International Journal of Automotive Technology, Vol. 10, No. 3, pp. 277−284 (2009)

DOI 10.1007/s12239−009−0032−2

EFFECTS OF INTAKE FLOW ON THE SPRAY STRUCTURE
OF A MULTI-HOLE INJECTOR IN A DISI ENGINE
S. KIM , J. M. NOURI , Y. YAN and C. ARCOUMANIS
1)*

2)

2)

2)

Department of Automotive Mechanical Engineering, Silla University, Busan 617-736, Korea
School of Engineering and Mathematical Sciences, The City University, Northampton Square, London, EC1V 0HB, UK

1)

2)

(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

ignition which, in turn, depends strongly on the spray
characteristics and, in particular, its cycle-to-cycle stability
which otherwise may even cause a misfire. Thus, to utilize
the full benefit of DISI technology, knowledge of the
temporal evolution of the spray structure, its tip penetration
and distribution of the droplet velocities and diameters as a
function of nozzle design, and injection and chamber
pressures is a prerequisite. It should be mentioned that the
swirl pressure atomiser (first generation) was found to be
unsuitable for the new concept of mixture preparation due
to demonstrated spray cone angle instability with back
pressure, leading to a complete collapse of the spray
structure when fuel was injected during the compression

stroke (Li et al., 2004; Nouri and Whitelaw, 2006).
Recently, a number of injector manufacturers have
designed new high-pressure multi-hole injectors and outwards opening piezo injectors, referred to as ‘secondgeneration’ systems, based on the expectation that they
produce stable fuel sprays with fine fuel droplets independent of the time of fuel injection (Wirth et al., 2004). Multihole injectors have been studied because of their potential
for achieving good fuel stratification, thus extending the
lean limit further (Preussner et al., 1998). They also offer
the highest possible flexibility in adapting the spray pattern
layout to a particular combustion chamber design. The
investigations of (Ortmann et al., 2001; Lippert et al.,
2004; Mitroglou et al., 2006, 2007) on multi-hole injectors
for gasoline engines confirmed the improved stability of
the spray at elevated chamber pressures relative to that of
swirl injectors. Also, enhanced air entrainment has been
observed as a result of an enlarged surface area produced
by separated spray jets, enhanced flexibility to direct the
sprays towards the proximity of the spark plug and improved

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
*Corresponding author. e-mail:
277


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
, 2005, 2006, 2007) regarding gasoline spray characteristics 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
., 2006; Nouri
., 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
., 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 configuration and experimental techniques for the present experiments 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.
et

al.

et

al

et

et

al

al

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

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
Bore × Stroke 83 × 92
In. Vavle
(mm)
timing
Compression
Ex. Valve
10.5
ratio
timing

Tumble/Swirl
6oBTDC/
50oABDC
50oBBDC/
6oATDC

measured were 1.38 for 1000 rpm in a steady flow rig test
(Karaiskos, 2005).
Optical access to the combustion chamber was provided

from the side (vertical images) via a fused silica cylinder
liner (Figure 1(b) and 1(c)). As shown in Figure 1(b), there
are also two quartz windows on both sides of the cylinder
head to provide access to the pent roof area. The piston
crown has a flat design so that an optical window can be
fitted to obtain horizontal images. The injector and spark
plug are oriented longitudinally, as shown in Figure 1(d);
the line of the spark plug and injector is in the middle of the
pent roof, between the intake and exhaust valves. All tests
were carried out without combustion, and with the engine
being motored. Identification of the engine cycle and crank
angle position was achieved by an optical pick up sensor
mounted on the exhaust camshaft, and a crankshaft encoder
(Muirhead Vactric), which produced 1440 pulses per
revolution, thus resulting in a resolution of 0.25ºCA.
Engine control was achieved by using an advanced timer
card (NI PCI-6602) with in-house software (Labview),
which controlled injection and ignition.
The prototype injector used in the present experiments
had been designed and manufactured by Bosch specifically
for DISI engines, and it is a high pressure six hole injector
with the holes symmetrically arranged on the periphery of
an imaginary circle, as shown in Figure 2. The detailed
specifications of the multi-hole injector, which operates
with injection pressures up to 20 MPa, are described in


EFFECTS OF INTAKE FLOW ON THE SPRAY STRUCTURE OF A MULTI-HOLE INJECTOR IN A DISI ENGINE 279
Table 3. Experimental conditions.
SCV

Injection
position Open/Close pressure 7 MPa /12 MPa
Coolant
Fuel
ISO Octane
temperature 40ºC/ 90ºC
Intake air
Operating Homogeneous/
temperature ~20ºC
mode
Stratified
Figure 2. Injector nozzle and spray view.
Table 2. Specifications of multi-hole injector.
Hole
6
Cone angle
42o
numbers
Hole
140 µm L/D ratio
2.14
diameter
Gasoline DISI
Manufacturer BOSCH Production
type
Proto type
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 highresolution forward illuminated images recorded with a nonintensified 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 conditions 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 in
Figure 2; images taken in the A-A plane view were used to
obtain the spray cone angle. For the investigation of incylinder spray characteristics, the injected sprays of the 6hole injector were visualized in the B-B plane view at
engine speeds of 1000 rpm. There is a 0.2 ms delay
between the injection trigger signal and first appearance of
a spray. Therefore, spray evolution images were captured
from 0.3 ms ASOI to 1.1 ms ASOI at 0.1 ms time intervals.

3. RESULT AND DISCUSSION
3.1. Early Injection for Homogeneous Stoichiometric Operation
Multi-hole injectors are known to have stable spray structures under various operating conditions. The overall spray
cone angle remains close to the nominal design value with
increasing chamber pressure; thus, early and late injection
during an engine’s cycle appear to have almost identical
spray shape, affecting only the spray’s penetration in the
combustion chamber. Homogeneous operation dictates
early injection of the fuel during the induction stroke. A
selection of early injection timing includes injection of fuel
at ATDC 60oCA, 90oCA, and 120oCA.
From the previous LDV measurement of in-cylinder
flow under ‘SCV open’ or tumble flow condition (Kariskos,
2005), it was realized that high velocities were generated
during the intake process, rising to a maximum between
ATDC 60oCA and 120oCA, and then decreasing in response
to the piston motion. During this period, the incoming high
velocity annular air-jet flows were directed axially towards
the down-going piston and radially towards the exhaust.
The results also showed that the generated swirl flow was
neither strong nor well defined with respect to cylinder
axis. The injected spray pattern during the intake stroke

with ‘SCV open’ can be strongly affected by the tumble
motion and its variation will result from the turbulence of
the swirl motion.
Evolution of the spray pattern at different injection
timings of ATDC 60oCA, 90oCA, and 120oCA, with the
SCV fully closed (maximum swirl), a fuel injection pressure
of 7 MPa, and a coolant temperature of 40oC is displayed in
Figure 3. As shown, there are two distinct features in the
spray structures; one is that the multiple spray plumes (jets)


280

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

Figure 3. Mie images during the intake stroke under swirl
flow, P = 7 MPa, and T = 40ºC.

Figure 4. Mie images during the intake stroke under tumble
flow, P = 7 MPa, and T = 40ºC.

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 recognizable, with downstream injected fuel droplets largely distributed in the cross-flow direction. This phenomenon represents 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 distinguished. The injected fuel spray plumes cannot avoid the
strong influence of the incoming air cross flow during
intake valve opening due to the injector position in the
cylinder head. Generally, the tumble flow does not deflect
the spray pattern as strongly as the swirl flow, and there is
no fuel droplet separation phenomenon; the latter indicates

no impingement on the liner. The larger spray deflection
and droplets separation with swirl flow, as seen in Figure 3,
clearly suggest the presence of centrifugal force acting on
the fuel droplets away from the center of the cylinder.
Overall comparison with the spray patterns under the
flow of Figure 3 indicates that the spatial distribution of the
injected fuel spray under tumble flow is apparently less
than that of swirl flow, especially over the elapsed time of
0.7 ms. In addition, the tilt of the overall spray in the
direction of the intake cross-flow is not as much as the
swirl. Therefore, for a well distributed, homogenized and
stoichiometric mixture, it is more important for swirl flow
to be generated in the cylinder than tumble flow.

inj.

coolant

inj.

coolant

3.2. Late Injection for Stratified Lean Operation Mode
The concept of stratification needs to be clarified according
to the engine design. At the time of ignition, an ignitable
mixture cloud should be around the vicinity of the spark
plug. This mixture cloud could be slightly rich in fuel locally,
while the remaining volume of the combustion chamber is
occupied by air. The size of the mixture cloud increases
with increasing engine load, and the load is controlled

quantitatively by the amount of fuel injection. The most
common technique to achieve mixture stratification is by
injecting the fuel during the compression stroke, and after
the closure of the inlet valve. In this study, three injection
timings during the compression stroke have been selected
ATDC 270oCA, 285oCA, and 300oCA, which were defined


EFFECTS OF INTAKE FLOW ON THE SPRAY STRUCTURE OF A MULTI-HOLE INJECTOR IN A DISI ENGINE 281

Figure 5. Mie images during the compression stroke under
swirl flow, P = 7 MPa, and T = 40ºC.
inj.

coolant

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 velocity 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

Figure 6. Mie images during the compression stroke under
tumble flow, P = 7 MPa, and T = 40ºC.
inj.

coolant

Figure 7. Mie image processing for vaporizing region.
straight penetration, unlike that of the swirl flow.
From the spray pattern of late injection during the compression stroke, it can be argued that the spray shape and

penetration were affected by the RMS component of incylinder flow and piston movement. In particular, the spray
penetration of the latest start of injection is strongly restricted
by the upward moving piston.
3.3. Temperature Effect on Spray Droplet Vaporization
Since the Mie scattering technique is based on scattered
light by liquid droplets only the remaining non-yet-vaporized
spray could be captured. More specifically, assuming that
the base spray image for characterizing evaporation would
be at the lowest available temperature, then the combination of images taken at the base and at a higher temperature would provide important qualitative information on
the relative percentage of liquid already vaporized, as was
suggested by (Mitroglou, 2005). The principle of this approach is shown schematically in Figure 7, and the outcome would represent the probability density function of
the liquid fuel droplets that are most likely to be evaporated.


282

S. KIM, J. M. NOURI, Y. YAN and C. ARCOUMANIS
The spray penetration and spray cone angle at different
injection timing, injection pressure, and coolant temperature
at 1000 rpm are plotted in Figure 9 and Figure 10. The
spray penetration of 60oCA SOI and T = 40 C is shown
in Figure 9(a). The penetration is affected by fuel injectionpressure so that at initial stage till 0.5 ms ASOI, the spray
penetration at 7 MPa is a little greater than that of 12 MPa,
mainly because the mechanical operational delay time of
the injector at 12 MPa is longer. However, the injected fuel
droplets had a substantial momentum, as a result of the
higher fuel pressure, and consequently, penetrated further
into the cylinder than those injected at a lower injection
pressure as the time ASOI increases.
From the elapsed time of 0.6 ms ASOI, the spray penetration at 12 MPa becomes greater than that at 7 MPa. The

penetration continues to increase until 0.8 ms ASOI, and
from 0.8 ms ASOI onward, the penetration stops at about
40 mm due to loss of droplet momentum. The spray
penetration of 300 CA SOI under swirl and tumble flow
are shown in Figure 9(b) and Figure 9(c), respectively. The
spray penetration during the compression stroke has a
similar trend to that of the intake stroke. The penetration is
also affected by fuel injection pressure. But additionally, it
is strongly affected by the chamber pressure (moving
piston), which causes a maximum penetration of 35 mm,
shorter than that of the intake stroke. After 0.9~1.0 ms
ASOI, the spray tip starts to impinge on the piston. The
injected fuel of high pressure reaches the piston earlier than
that of lower pressure. Therefore, it is necessary to carefully
consider the extent of fuel impingement according to the fuel
pressure. But, the temperature effect on the spray penetration
is small and not as noticeable as the fuel pressure. The
plane (A-A), where the overall spray angle was calculated,
is shown in Figure 2, and the angle was measured between
the extreme edges of the two outer jet sprays near the
injector tip, where the effects of the cross-flow was
minimum. Figure 10 shows the spray cone angle during
intake and compression stroke, and at different injection
pressures and coolant temperatures.
The results showed that the overall spray angle remained
constant and almost independent of injection pressure,
chamber pressure, and coolant temperature. There is also a
small and gradual reduction in the overall spray cone angle
coolant

Figure 8. Effect of the coolant temperature on fuel vaporization at ATDC 300°CA SOI under tumble flow.
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 needed 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 temperatures 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

Figure 9. Spray penetration during intake and compression strokes.

o

o


EFFECTS OF INTAKE FLOW ON THE SPRAY STRUCTURE OF A MULTI-HOLE INJECTOR IN A DISI ENGINE 283

Figure 10. Spray cone angle 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 compression 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 temperature 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.
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3


Copyright © 2009 KSAE

1229−9138/2009/046−03

International Journal of Automotive Technology, Vol. 10, No. 3, pp. 285−295 (2009)

DOI 10.1007/s12239−009−0033−1

INFLUENCE OF INJECTION PARAMETERS ON THE TRANSITION
FROM PCCI COMBUSTION TO DIFFUSION COMBUSTION
IN A SMALL-BORE HSDI DIESEL ENGINE
T. FANG , R. E. COVERDILL , C.-F. F. LEE and R. A. WHITE
1)*

2)

2)

2)

Department of Mechanical and Aerospace Engineering, North Carolina State University,
3182 Broughton Hall-Campus Box 7910, 2601 Stinson Drive, Raleigh, NC 27695, USA
Department of Mechanical Science and Engineering, University of Illinois at Urbana-Champaign,
1206 West Green Street, Urbana, IL 61801, USA
1)

2)

(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 DirectInjection (HSDI) diesel engine using multiple injection strategies. The heat release characteristics were analyzed using incylinder 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

quently, the NOx emissions are extremely low compared
with 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 PM
are also greatly reduced. Early studies of the HCCI combustion mode were carried out in two-stroke engines
(Onishi
., 1979; Noguchi
., 1979) and in fourstroke engines (Najt and Foster, 1983; Thring, 1989) by
using heavy Exhaust Gas Recirculation (EGR). It was
shown that, in the HCCI combustion mode, the ignition
process is controlled by low temperature (950 K) hydrocarbon oxidation kinetics, while the energy release process
is controlled by high temperature (above 1000K) hydrocarbon oxidation.
Multiple injection strategies have been reported for

simultaneous reduction of NOx and PM in both large bore
DI diesel engines (Nehmer
., 1994; Tow
., 1994;
Han
., 1996) and small-bore high-speed DI diesel
engines (Zhang, 1999; Tanaka
., 2002; Chen, 2000).
Several studies (Nehmer and Reitz, 1994; Tow
., 1994;
Han
., 1996) have shown that pulsed injections may

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) combustion is a promising technique that provides a unique
approach to simultaneously reduce NOx and PM emissions
while maintaining high thermal efficiency. For HCCI combustion, a premixed or ideally homogeneous air-fuel mixture auto-ignites due to compression; it is a bulk combustion, eliminating local high temperature regions. Conse*

Corresponding author.

et al

et al

et al

et

et al

al

et al

et al

e-mail:

et al

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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 injection in double or triple injection 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, investigations 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
(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 reduction of NOx and PM by using multiple injections 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
., 2002) were similar to the UNIBUS combustion
mode using double injections, as discussed in the following
sections.
Hashizume
. (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) concept (Takeda
., 1996). The first stage is premixed lean

combustion (PREDIC), and the second stage is diffusion
combustion under high temperature and low oxygen conditions. 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
. (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
.
(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 comet al.

et

al

et al

et al

et al

et al

pression stroke by Helmantel and Denbratt (2004). In order
to decrease the fuel wall impingement, a small-included
angle injector was used. The results showed a dramatic
reduction of NOx. Smoke emissions also showed a significant reduction, while HC and CO emissions substantially
increased.
Investigations by Hasegawa and Yanagihara (2003) employed two injections in the HCCI combustion mode,
referred as UNIform BUlky combustion System (UNIBUS).
The first injection was used as an early injection for early
fuel mixing and to advance the changing of fuel to lower
hydrocarbons, while the second injection was used as an
ignition trigger. Bulk combustion was observed in the
combustion chamber. Low NOx and smoke were possible
in both injections using this combustion concept. Another
two-stage diesel fuel injection HCCI combustion study was
done in a single cylinder small diesel engine (Kook and
Bae, 2004). A large fraction of fuel was injected early
during the compression stroke or even during the induction
stroke. A second injection with a small amount of fuel was
injected near the compression TDC to ignite all of the airfuel mixture. The experimental results showed that the
second injection could only be used as a combustion trigger
for low intake air temperature. The first injection timing
should be advanced earlier than 100 CAD BTDC to
achieve homogeneous and non-luminous combustion. NOx
was greatly reduced using this injection strategy. HC, CO,
and fuel consumption were higher than in conventional
diesel combustion.

Conceptually speaking, HCCI combustion is an ideal
operation mode for low emission diesel engines (Choi
., 2004). However, in a real diesel engine, it is quite difficult to homogeneously mix air and fuel using in-cylinder
direct injection strategies, even with a very early injection
during the suction stroke (Swami Nathan
., 2007). A
heterogenous premixed charge often occurs under these
injection strategies. Mixture heterogeneity often exists, even
for very early in-cylinder injection timings. In general, Premixed Charge Compression Ignition (PCCI) combustion is
a more accurate terminology for these conditions than
“HCCI”. PCCI only requires a premixed charge, and the
mixture is not required to be homogeneous. “PCCI” is a
broader concept than “HCCI”. Most of the above mentioned combustion modes are types of PCCI combustion.
In these previous studies, the combustion processes were
often visualized through an optical engine with modified
piston geometries. The replacement of the true piston shape
changes the flow field into which the fuel is injected. In this
work, the investigation uses an optical engine with a realistic
piston geometry. Among the current operating conditions, a
transition from PCCI combustion mode to conventional
diesel combustion mode was seen for the second main
injection. The influential factors such as injection pressure,
injection timing, and injection fuel quantities are studied
and the effects of the first injection parameters on the
combustion mode for the second injection are addressed.
et

al

et al

INFLUENCE OF INJECTION PARAMETERS ON THE TRANSITION FROM PCCI COMBUSTION

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 combustion 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
., 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 highspeed 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.
et al

3. ENGINE OPERATING CONDITIONS

The results presented in this paper are based on operating
conditions considered typical for this engine. Intake temperatures and pressures were increased to match the TDC
conditions of the metal engine with the same geometry and
operating conditions. The operating conditions are summarized 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
Table 1. Specifications of the single cylinder DIATA research
engine.
Bore
70 mm
Stroke
78 mm
Displacement/Cylinder
300 cc
Compression ratio
19.5:1
Swirl ratio
2.5
Valves/Cylinder
4
Intake valve diameter
24 mm
Ex. valve diameter
21 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.
First Pilot Main injecRail timing
tion timing
Case pressure
[CAD
number [bar] [CAD quantity
3
ATDC] [mm ] ATDC]
1
600
−40
0.8
0
2
600
−30
0.8
0
3
600
−40

1.3
0
4
600
−30
1.3
0
5
1000
−40
0.8
0
6
1000
−30
0.8
0
7
1000
−40
1.3
0
8
1000
−30
1.3
0

287

IMEP
[bar]
5.05
4.99
5.08
5.01
4.94
5.08
5.09
5.07

Table 3. Selected properties of the low-sulfur European
diesel fuel used during experimentation.
Specific gravity
0.8352
Cetane number
52.9
Sulfur, ppm
27.5
Mid boiling point, °C
260
load for all of the cases to be 5.0 bar IMEP. The injection
timing of the main injection was set at TDC for all of the
cases. The fuel used was a low-sulfur European Diesel fuel,
selected properties of which are shown in Table 3. Due to
the extensive optical access provided by the optical DIATA
engine, 3-D like combustion imaging was feasible (Fang
., 2005, 2006, 2007, 2008; Miles, 2000). Combustion
images were obtained using the high-speed video camera
by setting the operating frame rate at 12000 frames per

second with the resolution at 512×256 to capture the
images from the bottom and side. For all of the cases, the
exposure time was 2 ms.
et

al

4. RESULTS AND DISCUSSIONS
4.1. In-cylinder Pressure and Heat Release Analysis
The optical engine was warmed up by circulating heated
coolant and lubricating oil to simulate a warm engine
environment. The engine operated in skip fire mode in
order to reduce the heat load of the quartz piston, with one
injection cycle followed by 12 motoring cycles. Pressure
data were recorded and saved to the computer for post
processing.
Pressure traces for the eight cases are shown in Figures
1a and 2a. In the plots, 360 CAD corresponds to the
compression Top Dead Center (TDC). It is seen from the
figures that high injection pressure results in faster combustion, and thus more rapid pressure increase, due to
better fuel spray atomization and mixing. Combustion noise,
which is directly relevant to pressure rise rate, will be
higher for the higher injection pressure cases. Some knocklike combustion behaviors are seen for the high injection


288

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

Figure 1. In-cylinder pressures (a), heat release rates (b),

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.

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 observations 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, showing more rapid combustion and concentrated heat release
processes, while flatter and broader heat release rate
patterns are seen for the lower injection pressure cases. For
the second injection combustion, it is seen that some of the
high injection pressure cases are close to PCCI combustion. However, for the low injection pressure cases, diffusion combustion becomes more apparent, with increasing
first injection fuel quantity and decreasing retarding first
injection timing. The combustion mode transition from
diffusion combustion to PCCI combustion is observed in