Energy 64 (2014) 707e718
Contents lists available at ScienceDirect
Energy
journal homepage: www.elsevier.com/locate/energy
Experimental study on compound HCCI (homogenous charge
compression ignition) combustion fueled with gasoline and
diesel blends
Xingcai Lu*, Yong Qian, Zheng Yang, Dong Han, Jibin Ji, Xiaoxin Zhou, Zhen Huang
Key Lab. for Power Machinery and Engineering of M. O. E., Shanghai Jiao Tong University, 200240 Shanghai, PR China
a r t i c l e i n f o
a b s t r a c t
Article history:
Received 29 May 2013
Received in revised form
10 September 2013
Accepted 20 October 2013
Available online 21 November 2013
Because the energy crisis and environmental pollution are significant concern, a next-generation combustion mode for internal combustion engines that can simultaneously reduce exhaust emissions and
substantially improve thermal efficiency has attracted increasing attention. In the last two decades,
diesel-fueled HCCI (homogenous charge compression ignition) combustion has been widely researched.
It has been determined that diesel HCCI combustion has the potential to improve NOx and soot emissions
at low-to-medium loads; however, it suffers from higher HC (hydrocarbon)/CO (carbon monoxide)
emissions, narrow operating ranges, and uncontrollable ignition timing and combustion rates due to low
volatility and high ignitability. For this reason, 30%e50% (v/v, briefly G30, G40, and G50) gasoline/diesel
fuel blends are used in a novel technology of a compound HCCI combustion mode, in which port fuel
injection and in-cylinder direct injection are combined. The combustion and emission characteristics of
the compound HCCI combustion using blend fuels are investigated on a single-cylinder engine. The effect
of the gasoline volume in the blends, the premixed ratio, and the overall fuel supply rate on compound
HCCI combustion are initially evaluated, and the effects of the intake air boost on G30 compound HCCI
combustion is also investigated.
The experimental results indicate that the maximum heat release rate, maximum in-cylinder pressure,
and NOx emissions of G40 and G50 compound HCCI combustion significantly increase when compared
to that of G30 compound HCCI combustion. Moreover, it is determined that the intake air boost has great
potential to reduce the NOx and soot emissions of compound HCCI combustion simultaneously. CO and
HC emissions of optimized G30 compound HCCI combustion with/without boost are relatively higher
compared to that of traditional DICI (direct injection compression ignition) combustion. The NOx and
soot emissions of optimized G30 compound HCCI combustion with intake air boost are far lower than
those of the DICI diesel engine. Specifically, the NOx emissions can be maintained within 100 ppm, and
the soot emissions are below 10% at the full load ranges.
Ó 2013 Elsevier Ltd. All rights reserved.
Keywords:
Compound homogenous charge
compression ignition
Gasoline/diesel blend
Two-stage fuel supply
Intake air boost
Combustion
1. Introduction
Since Onishi et al. [1] and Noguchi et al. [2] proposed the HCCI
(homogenous charge compression ignition) concept in the early
1980’s, many researchers worldwide have conducted detailed investigations and acquired a deep understanding of this novel
combustion model over the last two decades [3]. HCCI combustion
is considered to be the most promising clean combustion method
with high efficiency that will be able to meet future emissions
* Corresponding author. No. 800, Dongchuan Road, School of Mechanical & Power
Engineering, Shanghai Jiaotong University, Shanghai 200140, PR China. Tel.: þ86 21
34206039; fax: þ86 21 34205949.
E-mail address: [email protected] (X. Lu).
0360-5442/$ e see front matter Ó 2013 Elsevier Ltd. All rights reserved.
http://dx.doi.org/10.1016/j.energy.2013.10.068
regulations [4]. However, the ignition timing and heat release
process of HCCI combustion is difficult to control. Therefore, a novel
compound HCCI combustion strategy is developed, which solves
the combustion control problem while retaining the advantages of
pure HCCI combustion. Currently, two types of compound HCCI
combustion are proposed: compound HCCI combustion based on
mode transitions [5] and compound HCCI combustion based on
mixture stratification [6].
For HCCI combustion fueled with gasoline-like fuels, which have
a lower boiling point and higher octane number, a homogenous
mixture may be obtained before autoignition regardless of port fuel
injection or in-cylinder direct injection; thus, the major obstacle for
its combustion control is to depress the combustion rate at larger
engine loads. Until now, the traditional strategy, such as internal
with/without external exhaust gas recirculation, could not match
708
X. Lu et al. / Energy 64 (2014) 707e718
the need for full load HCCI combustion, and therefore compound
HCCI combustion based on mode transitions may be the optimal
choice. Compound HCCI combustion based on mode transitions
means that the engine operates in a traditional SI (spark ignition)
mode at cold start, idle and high load conditions and switches to
HCCI mode at low and medium loads [7,8]. Tian et al. obtained
steady switching processes from the HCCI to SI (spark ignition)
mode within 10 engine cycles, but the smooth transition process
from the SI to HCCI mode is still difficult to obtain [9]. In addition,
Zhang et al. confirmed that switching from HCCI to SI operation is
less problematic than switching from SI to HCCI [10].
For HCCI combustion using diesel fuel, which has a high cetane
number and high viscosity, the major problem is the preparation of
a homogenous mixture during the short ignition delay [11]. To
resolve this problem, a compound HCCI combustion based on
mixture stratification, including a multi-pulse injection scheme
[12], internal with/without external EGR (exhaust gas recirculation)
[13], low temperature combustion with higher EGR rate [14], and a
PFI (port fuel injection) of gasoline fuel combined with in-cylinder
direct injection of diesel or biodiesel fuel [15], is used to control the
in-cylinder mixture distribution and combustion process. Occasionally, this combustion mode is also called as SCCI (stratified
charge compression ignition) [16]. In fact, stratification includes the
inhomogeneity distribution of temperature, mixture concentration,
and fuel composition during the overall combustion history [17].
Kumano et al. [18] confirmed that charge inhomogeneity has the
potential to prolong the combustion duration. Richter et al. [19] also
demonstrated that charge inhomogeneity has a modest effect on
the combustion process. Yu et al. [20] determined that temperature
inhomogeneity has a clear effect on ignition timing, and Dec et al.
[21] concluded that thermal stratification over the bulk gases is an
effective method for controlling the maximum pressure rising rate.
According to the abovementioned studies, it is widely accepted that
SCCI combustion is an effective method for the control of combustion phasing and can expand the operating range to high loads
[22]. However, there are still various challenges for SCCI combustion. One important issue among these challenges is the preparation of stratified mixtures because the formation of stratified
mixture can be influenced by many factors, such as turbulence and
heat transfer.
Lu et al. developed a novel combustion strategy by combining
the concepts of compound HCCI combustion and fuel design and
injection management [23]. This combustion technology focuses on
the adjustment of fuel physicalechemical properties, the development of fuel injection strategies, and incorporation of the advantages of traditional DICI (direct injection compression ignition),
HCCI and SCCI combustion strategies. In this combustion strategy,
the fuel composition can be modulated in real time, and the combustion phasing and heat release rate can be controlled [24]. Ma
et al. indicated that this compound HCCI combustion mode fueled
with PRF (primary reference fuels) can achieve the full load while
maintaining low emissions and high efficiency [25]. However, primary reference fuels cannot be directly used in commercial engines. Since the invention of the engine, gasoline and diesel fuels,
which are products refined from crude petroleum oil, are two basic
fuels that have been widely supplied as the main fuels for internal
combustion engines. Gasoline has high volatility and low ignitability while diesel fuel has high ignitability and low volatility,
which is better suited for self-ignition. For these reasons, a gasoline/diesel fuel blend, known as dieseline, was proposed to be used
in traditional and advanced combustion modes to improve thermal
efficiency and emissions.
Zhong et al. [26] conducted a study combining the properties of
gasoline and diesel using a single-cylinder HCCI engine equipped
with port fuel injection and either NVO (negative valve overlap) or
intake charge heating. Weall et al. [27] researched the PPCI
(partially premixed compression ignition) combustion fueled with
a 50% gasoline/diesel blend. Han et al. [28,29] experimentally
demonstrated the potential of using blends of diesel and gasoline to
simultaneously reduce nitrogen oxides and soot emissions in the
premixed LTC (low-temperature combustion) modes. Won et al.
[30] reported the experimental results of a single-cylinder diesel
engine running with PPCI using gasoline/diesel blends. The results
demonstrated that the engine could run on such blends with
extremely low smoke and low nitrogen oxide emissions at speeds
of up to 4000 r/min, IMEP (indicated mean effective pressures) of
up to 10 bar with an injection pressure of only 400 bar. Zhang et al.
[31] confirmed that the ease of ignition and difficulty of the
vaporization of diesel fuel made it imperfect for PPCI combustion;
however, the gasoline/diesel blend fuels have the potential to
simultaneously reduce emissions by more than 95% with dieseline
fueled PPCI combustion. However, the penalty is a slightly
increased noise level and a lower indicated efficiency, which is
decreased from 40% to 38.5% [32]. Turner et al. [33] found that the
gasoline/diesel blend provides a few unexpected benefits to the
expansion of the operating window and the reduction of hydrocarbon emissions in HCCI engines; these benefits include an
extended low misfire limit, increased engine stability, reduced peak
cylinder pressures and reduced emissions within the entire HCCI
operating window.
Park et al. [34] investigated the effects of gasolineediesel blends
on the fuel properties, droplet atomization, combustion performance, and exhaust emission in a four-cylinder diesel engine. The
results determined that the blending of gasoline caused a decrease
in droplet size by increasing the small droplets and decreasing the
large droplets as the surface tension decreased with the additional
gasoline fuel, thereby inducing an increase in droplet instability.
However, the gasoline blending resulted in an extension of the
ignition delay and the formation of a more homogeneous mixture.
These combustion characteristics caused the simultaneous reduction of ISNOx and ISsoot. However, the ISHC (indicated specific
hydrocarbon) and ISCO (indicated specific carbon monoxide)
emissions were slightly increased. Valentino et al. [35] evaluated
the effects of gasoline/diesel fuel blends on performance and
engine-out emissions by attempting to deliver the entire amount of
fuel before the ignition. By taking advantage of the higher resistance of G40 to auto-ignition, it was possible to extend the range in
which a partial premixed combustion was achieved. The longer
ignition delay and improved mixing before combustion created
more advanced injection timings, which possibly reduced smoke
and nitrogen oxide emissions. The joint effect of higher resistance
to auto-ignition and a higher volatility of the gasoline improved the
emissions of the blends compared to the neat diesel fuel, with a low
penalty on fuel consumption.
For the gasoline/diesel blend fuel, several studies have focused
on traditional direct injection compression ignition combustion,
homogenous charge compression ignition combustion, and low
temperature combustion modes; investigations have rarely been
conducted on compound HCCI combustion. Due to the acceptable
mixture formation characteristics and moderate ignition characteristics of gasoline/diesel blends, a two-stage fuel supplying
strategy, a port fuel injection combined with in-cylinder direct injection, was used in this study. The mixture formation timescale
and ignition delay can be controlled both macroscopically and
locally in the combustion chamber using the mixture and component stratifications; therefore, high efficiency and low emissions
are achieved in wide operating ranges. To this target, different
blends of gasoline and diesel (G30, G40, and G50) are used in
compound HCCI combustion, which are realized through fuel port
injection combined with in-cylinder direct injection. The effects of
X. Lu et al. / Energy 64 (2014) 707e718
Table 1
Engine specification.
Bore Stroke
Displacement
Combustion chamber
Compression ratio
Needle Open Pressure
Intake valve open
Intake valve close
Exhaust valve open
Exhaust valve close
a
709
Table 2
Fuel properties.
98 mm 105 mm
0.792 L
u type
18.5
19 MPa
344 CA ATDCa
128 CA BTDC
114 CA ATDC
348 CA BTDC
Density (g/ml @ 298 K)
Lower heat value (MJ/kg)
Cetane number
Research octane number
T90 C
Viscosity (mm2/s @40 C)
Latent heat of vaporization (kJ/kg)
Sulfur (mass %)
Diesel
Gasoline
0.834
42.9
56
e
350
3.410
260
<0.005
0.746
44.1
e
97
180
0.585
340
<0.005
TDC in this study is referred to be the combustion Top Dead Center.
the gasoline volume, the premixed ratio, and the total fuel quantity
on the combustion and emissions characteristics of compound
HCCI combustion fueled by diesel and gasoline blends are investigated. Moreover, the potential of the intake air boost to broaden the
operational range of compound HCCI combustion is studied. Last,
the authors compared the emission levels of the G30 traditional
DICI combustion mode and compound HCCI with/without an
intake boost for an original diesel engine. It is possible that an
optimal strategy to solve the NOx-soot trade-off relationship for
dieseline compound HCCI combustion under full load operating
ranges can be determined.
2. Experimental apparatus and procedure
The experiment was conducted on a single-cylinder, direct-injection and four-stroke naturally aspirated diesel engine. The main
engine specifications are listed in Table 1. Fig. 1 shows the schematic
of the experimental setup. A port fuel injection system was
employed to supply a part of fuel into the intake manifold to form
the homogeneous charge. The injector of this system was mounted
approximately 0.45 m upstream from the intake valve, with an injection pressure of 11 MPa and port fuel injection timing of 340 CA
BTDC. Another injector with a cone angle of 154 was used to inject
fuel directly into the cylinder. The DI (direct injection) timing can be
held constant by maintaining the fuel supply advance angle at 9 CA
BTDC. To adjust the intake air pressure, an intake air boost system
was designed for the engine. Therefore, the test engine can operate
in both the naturally aspirated and intake boost modes.
Blends of gasoline and diesel were used in this study. The fuel
properties of commercial gasoline and diesel are listed in Table 2.
The in-cylinder gas pressure was measured by a pressure
transducer (Kistler model 6125B). The charge output from this
transducer was converted to an amplified voltage using an amplifier (Kistler model 5015A) and recorded at a 0.25 CA resolution. The
in-cylinder gas pressure is averaged from 50 consecutive cycles for
each operating point, and the heat release rate, bulk gas temperature, and indicated mean effective pressure are calculated from a
zero-dimension combustion model based on cylinder pressure
[36]. The exhaust gas composition of the CO, UHC (unburned hydrocarbon), and NOx emissions were measured by a gas analyzer
(AVL Digas 4000). The smoke opacity was measured by a smoke
meter (AVL 439 Opacimeter). The measured parameters and their
accuracy are summarized in Table 3.
For all data presented, 0 CA is defined as the TDC (top dead
center) at the compression stroke. To ensure the repeatability and
comparability of the measurements for the operating conditions,
the temperatures of the intake air, lubricant oil and coolant water
were held constant during the experiments. The engine speed was
maintained at 1800 rpm.
3. Definition of the combustion parameters
To investigate the combustion and emissions characteristics of
compound HCCI combustion fueled with gasoline/diesel blends,
several basic combustion parameters are defined as follows: the
ignition timing of compound HCCI combustion is defined as the
crank angle position at which 10% of the fuel is burned, which is
Fig. 1. Schematic of experimental setup.
710
X. Lu et al. / Energy 64 (2014) 707e718
Table 3
Accuracies of the measurements.
Measured parameters
Unit
Measurement range
Accuracy
Engine speed
Engine torque
Cylinder pressure
Intake pressure
Intake air temperature
Exhaust gas temperature
Coolant temperature
Lubricant oil temperature
CO emissions
HC emissions
NOx emissions
Soot emissions
Fuel consumption
r/min
Nm
MPa
kPa
C
C
C
C
%
ppm
ppm
%
kg/h
0e10,000
0e200
0e25
0e500
0e200
0e800
0e150
0e150
0e4.0
0e10,000
0e4000
0e100
0e20
1
0.1
0.0005
0.5
0.1
1
0.1
0.1
0.01
1
1
1
0.01
also referred to as CA10 and the CA50, another key parameter used
to assess the combustion phasing, is defined as the crank angle
position at which 50% of the fuel is burned. The burn duration of the
combustion is defined as the crank angle interval between 10% and
90% of the mass fraction burned [37].
The gasoline and diesel blends used in this test are denoted as
G30, G40, and G50, indicating that there is 30%, 40%, and 50%
gasoline by volume in the blended fuel, respectively. The total fuel
consumption per cycle, which is the sum of the premixed and
directly injected fuel amounts, and the intake air pressure are
denoted as b and Pin, respectively. In this paper, the premixed ratio
rp is defined as the ratio of the energy of the premixed fuel QPI to the
total energy Q. The premixed ratio can be calculated using the
following formula:
mPI Hu;PI
Q
Rp ¼ PI ¼
Q
mPI Hu;PI þ mDI Hu;DI
where mPI and mD represent the mass consumption rate of the
premixed and directly injected fuels, respectively. Hu;PI and Hu;DI
are lower heating values of the premixed and directly injected fuels,
respectively.
Fig. 3. Effect of premixed ratio on the maximum pressure and maximum massaveraged temperature of dieseline compound HCCI combustion.
4. Experimental results and discussion
4.1. Effects of the premixed ratio on the combustion and emissions
characteristics of dieseline compound HCCI combustion
In this section, the total fuel consumption per cycle is maintained at 29.63 mg/cycle in the naturally aspirated operation, and
the effects of the premixed ratio on the combustion and emissions
characteristics of compound HCCI combustion using different
blends are investigated.
Fig. 2 shows the HRR (heat release rate) and in-cylinder gas
pressure of dieseline compound HCCI combustion. Fig. 2a shows
the effects of the premixed ratio on G30 compound HCCI
Fig. 2. Effect of the premixed ratio and gasoline volume on the in-cylinder gas pressure and the heat release rate of dieseline compound HCCI combustion.
X. Lu et al. / Energy 64 (2014) 707e718
Fig. 4. Effect of the premixed ratio on the maximum pressure rising rate of dieseline
compound HCCI combustion.
combustion characteristics, and Fig. 2b illustrates the effects of the
gasoline volume in blends while maintaining the same premixed
ratio on the combustion characteristics. In Fig. 2a, similar to our
previous research [38,39], the heat release rate of compound HCCI
combustion is shown to exhibit a three-stage combustion process
containing a LTR (low temperature reaction), a HTR (high temperature reaction) of the premixed fuel and diffusive combustion of the
directly injected fuel. Additionally, it can be observed that with an
increase in Rp, the peak values of the heat release rate of the LTR and
711
HTR increase while the peak value of the heat release rate of the
diffusive combustion decreases.
In Fig. 2b, with an increase in the gasoline volume in the blended
mixtures, the ignition timing of the low temperature combustion is
gradually delayed and the peak value of the LTR is reduced. This
trend is due to the diesel fuel, which possibly reacts and releases
heat during the low temperature ranges decreases as the gasoline
volume increases. Moreover, it can be found from Fig. 2b that the
peak points of the high temperature reaction and maximum gas
pressures of G40 and G50 are larger compared to those of G30. This
result is because both G40 and G50 have the excellent ability to
form a homogenous mixture before the high temperature combustion occurs. Lastly, it is observed that the ignition of the high
temperature combustion for G40 is comparable to that of G30 but is
larger compared to that of G50.
Fig. 3 presents the effect of the premixed ratio on the maximum
pressure and maximum mass-averaged temperature of dieseline
compound HCCI combustion. From Fig. 3 to Fig. 5, six premixed
ratios (0, 0.23, 0.375, 0.5, 0.625, and 0.75) were selected and
compared for three blends. In Fig. 3, it can be determined that with
an increase in Rp, the maximum pressure and maximum massaveraged temperature initially decrease and then increase. In
addition, it can be observed from the figures that the maximum incylinder pressure and in-cylinder mean gas temperature of G40
compound HCCI combustion clearly increase and are significantly
higher compared to those of the other two blends’ combustion
events. Furthermore, the maximum in-cylinder mean gas temperature of traditional DICI combustion events (Rp ¼ 0) are higher than
those for compound HCCI combustion regardless of the premixed
ratios.
Fig. 5. Effect of the premixed ratio on emission characteristics of dieseline compound HCCI combustion.
712
X. Lu et al. / Energy 64 (2014) 707e718
Fig. 8. Effects of total fuel consumption per cycle on the maximum pressure and
maximum mass-averaged temperature of dieseline compound HCCI combustion.
Fig. 6. Effect of total fuel consumption per cycle on the heat release rate of G30
compound HCCI combustion.
Fig. 4 displays the effect of the premixed ratio on the maximum
pressure rising rate of dieseline compound HCCI combustion. It is
observed that the macro tendency of the maximum pressure rising
rate versus the premixed ratio is similar to those of the maximum
pressure and the maximum mass-averaged temperature. However,
for larger premixed ratios, for example, larger than 0.4, the
maximum pressure rising rate of G40 compound HCCI combustion
is significantly higher compared to that of G30 and G50 combustion. This trend may be attributable to the higher combustion rate
of G40 but with a similar ignition timing when compared to G30.
For G50 compound HCCI combustion, the ignition occurred after
Fig. 7. Effects of total fuel consumption per cycle on the combustion phasing and burn
duration of dieseline compound HCCI combustion.
the top dead center, and the main combustion process was
completed during the expansion stroke. Therefore, the maximum
gas pressure, maximum mass-averaged temperature, and
maximum pressure rising rate of G40 compound HCCI combustion
are highest when compared to those of G30 and G50 compound
HCCI combustion.
In fact, with an increased premixed ratio, the proportion of the
diffusive combustion gradually decreases while the proportion of
HCCI combustion increases. Therefore, the combustion characteristics of dieseline compound HCCI combustion gradually transitions
from DICI-based combustion to HCCI-based combustion. That is, at
a low premixed ratio, compound HCCI combustion is dominated by
diffusive combustion, and several combustion parameters, such as
the maximum pressure, maximum mass-averaged temperature
and maximum pressure rising rate, are determined by the fuel/air
equivalence ratio of the directly injected fuel. In this condition, with
an increased premixed ratio, the fuel/air equivalence ratio of the
directly injected fuel decreases, and the maximum pressure,
maximum mass-averaged temperature and maximum pressure
rising rate gradually decrease. However, as the premixed ratio increases, HCCI combustion begins to hold a dominant position over
compound HCCI combustion, and the combustion parameters are
determined by the fuel/air equivalence ratio of the premixed fuel. In
this condition, the increased fuel/air equivalence ratio of the premixed fuel leads to an increased maximum pressure, maximum
mass-averaged temperature and maximum pressure rising rate.
Fig. 9. Effect of total fuel consumption per cycle on the maximum pressure rising rate
of dieseline compound HCCI combustion.
X. Lu et al. / Energy 64 (2014) 707e718
713
Fig. 10. Effects of total fuel consumption per cycle on the emission characteristics of dieseline compound HCCI combustion.
Fig. 5 shows the effects of the premixed ratio on the emissions
characteristics of compound HCCI combustion. It can be observed
that with an increased premixed ratio, the CO, HC and soot emissions initially increase and then decrease, while the NOx emissions
decrease monotonically for each tested fuel. These phenomena can
also be attributed to changes in the combustion characteristics. In
generally, the overall combustion event of compound HCCI combustion consists of premixed combustion of port injected fuels and
diffusion combustion of directly injected fuels, as shown in Fig. 1.
Specifically, at low premixed ratios, diffusive combustion plays a
vital role; thus, CO and HC emissions are at a low level, which is a
feature of the DICI combustion mode. When the premixed ratio
increases, premixed fuel/air mixture concentrations before autoignition increase correspondingly, then the overall combustion
event is dominated by HCCI combustion. Therefore, the CO and HC
emissions start to increase while the NOx emissions decrease. For
soot emissions, low-temperature and high-temperature reactions
of the premixed fuel consumed amount of oxygen before the
directly fuel injection. Thus, it reduces the in-cylinder oxygen
concentration and elevates the in-cylinder mixture temperature,
similar to the effects of the internal EGR, which contributes to soot
formation. Therefore, soot emissions increase when the premixed
ratio initially increases. A further increased premixed ratio allows
the premixed fuel to burn completely, and the CO and HC emissions
of the HCCI combustion can be improved. The decreasing tendency
of soot emissions at this time is most likely due to the higher
proportion of HCCI combustion.
When considering the effects of gasoline volume on emissions,
it can be observed that the CO emissions of G40 with a premixed
ratio larger than 0.6 are clearly lower than those of G30 and G50
(Fig. 5a). The CO emission level is mainly due to the incomplete
oxidization. Thus, a higher combustion temperature (Fig. 3) will
lead to lower CO emissions levels. In Fig. 5b, the HC emissions of
G50 with a premixed ratio larger than 0.5 are significantly higher
compared to other blend fuels. This trend may be due to the high
volatility but low ignitability of gasoline in G50. Regarding the NOx
emissions levels, it is observed that the gasoline volume has a
negligible effect on the different premixed ratios, as shown in
Fig. 5c. In Fig. 5d, the smoke opacity of G40 is slightly higher
compared to that of G30 and G50 under the same conditions. The
following explanations may help to understand this trend. The
ignition timing of G40 is similar to G30 but its combustion (the
maximum heat release rate) is larger compared to that of G30 and
G50. Then, the shortly ignition delay allows the local rich fuel/air
mixture to burn at a high speed. Thus, high smoke emissions are
formed and released.
4.2. Effects of total fuel consumption per cycle on the combustion
and emission characteristics of dieseline compound HCCI
combustion
In this section, the premixed ratio is fixed at 0.5 for the naturally
aspirated operation, and the effects of the total fuel consumption
per cycle on the combustion and emissions characteristics of
compound HCCI combustion using different gasoline/diesel blends
are studied. From Fig. 6 to Fig. 10, four total fuel consumptions
(14.81 mg/cycle, 22.22 mg/cycle, 29.63 mg/cycle, and 37.04 mg/
cycle) were selected and compared for three blends at a premixed
ratio of 0.5.
Fig. 6 shows the effect of the total fuel consumption per cycle on
the heat release rate of G30 compound HCCI combustion. It can be
found that with an increase in the total fuel consumption per cycle,
714
X. Lu et al. / Energy 64 (2014) 707e718
Fig. 11. Effects of intake air pressure on the combustion parameters of G30 compound HCCI combustion.
the peak values of the heat release rate of the LTR and HTR increase,
and the ignitions of the LTR and HTR are advanced. For a fixed
premixed ratio, the mixture concentration before the direct injection increases as the total fuel quantity increases; this trend improves the low temperature and high temperature reactions of the
premixed homogenous mixtures. As shown in this figure, it can be
observed that the initial heat release phasing of the directly
injected fuels also advances as the total fuel increases. This result
may be attributed to two facts, a strongly thermal atmosphere and
a larger amount of directly injected fuel as the amount of total fuel
increases.
Fig. 7 illustrates the effect of the total fuel consumption per cycle
on the combustion phasing and burn duration of dieseline compound HCCI combustion. It is evident that with an increase in the
fuel flow rate, CA10 is significantly advanced. However, CA50
changes slightly compared to CA10. The reason for this phenomenon is that the increase in the total fuel consumption per cycle
extends the burn duration of compound HCCI combustion, which
retards the location of CA50.
Fig. 8 displays the effects of the total fuel consumption per
cycle on the maximum pressure and maximum mass-averaged
temperature of dieseline compound HCCI combustion. The
increased total fuel consumption per cycle brings additional energy into the cylinder, thus elevating the maximum pressure and
mass-averaged temperature. Another point that can be observed
is that the maximum in-cylinder pressure and averaged gas
temperature of G30 and G40 are slightly larger compared to that
of G50 in the 22.22 mg/cycle and 29.63 mg/cycle. However, at a
total fuel rate of 37.04 mg/cycle, the maximum temperature and
gas temperature of G30 are comparable to that of G50 but these
parameters are clearly lower than that of G40. This result is
because the ignition timing of G40 is comparable to that of G30,
and both occur near the top dead center; the ignition timing of
G50 is observed after the top dead center. Additionally, the
combustion rate of G40 is slightly larger than that of G30 and
G50.
Fig. 9 displays the effect of the total fuel consumption per cycle
on the maximum pressure rising rate of dieseline compound HCCI
combustion. The maximum pressure rising rate is maintained at a
relatively low level when the total fuel consumption per cycle is
below 30 mg/cycle. However, once the total fuel consumption per
cycle exceeds 30 mg/cycle, the maximum pressure rising rate increases remarkably. At a high total fuel consumption condition,
premixed fuel/air mixture concentrations increase clearly, and the
ignition of the premixed mixture occurs during the compression
stroke. As result, the maximum pressure rising rate increases
significantly.
Fig. 10 demonstrates the effects of the total fuel consumption
per cycle on the emissions characteristics of dieseline compound HCCI combustion. It can be observed that with an increase in the total fuel consumption condition, the CO
emissions initially increase and then decrease, and the HC
emissions decrease monotonically while the NOx and soot
emissions increase gradually. The increasing trend of CO emissions when the total fuel consumption condition is below
25 mg/cycle is likely due to the incomplete HCCI combustion of
the premixed G30 so that the conversion of CO-to-CO2 reaction
does not occur [40].
X. Lu et al. / Energy 64 (2014) 707e718
715
Fig. 12. Effects of intake air pressure on combustion phasing and the burn duration of G30 compound HCCI combustion.
4.3. Effects of intake air pressure on the combustion and emission
characteristics of G30 compound HCCI combustion
This section discusses the effects of the intake air pressure on
the combustion and emissions characteristics of G30 compound
HCCI combustion, where the total fuel consumption per cycle is
maintained at 37.04 mg/cycle. Six different values of intake air
pressure are chosen in this study, which are 100 kPa, 110 kPa,
120 kPa, 130 kPa, 140 kPa and 150 kPa.
Fig. 11 indicates the effects of the intake air pressure on the
combustion parameters of G30 compound HCCI combustion, such
as the maximum pressure rising rate, the maximum mass-averaged
temperature and the gross value of the indicated mean effective
pressure (IMEP-gross). The intake air boost can significantly reduce
the maximum pressure rising rate when the premixed ratio is
above 0.4, as shown in Fig. 11(a). This result occurs because with an
increase in the intake air pressure, the fuel/air mixture in the cylinder is leaner, and the HCCI combustion rate gradually decreases.
With an increase in the intake air pressure, the IMEP-gross initially
decreases and then increases while the maximum mass-averaged
temperature initially decreases rapidly and then gradually, as
shown in Fig. 11(b) and (c).
As mentioned above, the entire combustion can be divided into
the HCCI combustion of premixed fuel and the diffusive combustion of the directly injected fuel. However, the effects of the intake
air boost on these two parts of combustion are different. For the
HCCI combustion of premixed fuel, the intake air boost reduces the
concentration of the homogenous fuel/air charge. Therefore,
incomplete HCCI combustion may occur and the IMEP-gross can
likely be decreased. In contrast, the intake air boost can elevate the
O2 concentration, which may promote the ignition and combustion
of the directly injected fuel. Hence, the ignition timing of the
diffusive combustion can be advanced, and the IMEP-gross can
likely be improved. The decrease in the maximum mass-averaged
temperature is attributed to a higher volumetric heat capacity
when the intake air boost is used. However, the distinct variation of
the descending rate may be influenced by the different effects of
the boost.
Fig. 12 presents the effects of the intake air pressure on the
combustion phase and burn duration of G30 compound HCCI
combustion. With increased intake air pressure, CA10 and CA50 are
retarded and then advanced, as observed in Fig. 12(a) and (b). The
burn duration is progressively shortened when the boost intensity
grows, as shown in Fig. 12(c).
Fig. 13 shows the effects of the intake air pressure on the
emissions characteristics of G30 compound HCCI combustion.
Compared to natural aspiration, the CO emissions are higher with
the use of the intake air boost when the premixed ratio is above 0.5,
as shown in Fig. 13(a). However, once the intake air boost is used,
the CO emissions remain almost unchanged in the case of the fixed
premixed ratio. It appears that the boost intensity has minimal
effects on the CO emissions. The HC emissions are also maintained
at the same level with varied intake pressures, as shown in
Fig. 13(b). Regardless of the premixed ratio, the NOx emissions
clearly decrease at first and then increase slightly while the soot
emissions decrease monotonically with an increase in the intake air
pressure, as shown in Fig. 13(c) and (d). Therefore, the intake air
boost shows great potential to simultaneously reduce the NOx and
soot emissions of compound HCCI combustion while maintaining
the low level of CO and HC emissions.
4.4. Optimization of G30 compound HCCI combustion
This section demonstrates the results of the G30 compound
HCCI combustion when the operating parameters are optimized.
Specifically, for G30 compound HCCI combustion without intake air
716
X. Lu et al. / Energy 64 (2014) 707e718
Fig. 13. Effects of intake air pressure on the emission characteristics of G30 compound HCCI combustion.
boost, the optimal premixed ratio for the lowest NOx emissions is
selected. While, for the G30 compound HCCI combustion with
intake boost, the optimal premixed ratio at G30 without boost is
selected at each fuel supplying rate, and the intake charge pressure
is changed so as to obtain the lowest NOx emissions.
Fig. 14 presents the various emissions of G30 compound HCCI
combustion without the intake air boost under the optimal Rp, as
well as the emissions of G30 compound HCCI combustion with
boost under the optimal Pin. Additionally, Fig. 14 provides the results of G30 DICI combustion and diesel DICI combustion of the
prototype engine for comparison. It is evident that both G30 DICI
combustion and G30 compound HCCI combustion with/without
boost can achieve the full load of the prototype engine, as shown in
Fig. 14. The level of CO and HC emissions for G30 and diesel DICI
combustion is quite low at the full load ranges. However, the CO
and HC emissions of the optimized G30 compound HCCI combustion with/without boost are relatively higher compared to that of
the prototype engine in Fig. 14(a) and (b). Furthermore, the CO
emissions of the optimized G30 compound HCCI combustion with/
without boost initially increase and then decrease with an increase
in the engine load. Furthermore, the HC emissions of G30 compound HCCI combustion without boost decrease rapidly and then
remain approximately constant with an increase in the engine load,
and the HC emissions of G30 compound HCCI combustion with
boost decrease slightly.
Compared to the prototype engine, the G30 DICI combustion can
greatly reduce soot emissions, but the level of NOx emissions
remains nearly the same, as shown in Fig. 14(c) and (d). However,
the optimized G30 compound HCCI combustion without boost can
significantly reduce NOx emissions, but the level of soot emissions
is similar to that of the prototype engine. That is, neither of the
combustion methods can achieve the goal of reducing NOx and soot
emissions simultaneously. For the optimized G30 compound HCCI
combustion with intake air boost, NOx and soot emissions are far
lower than for the prototype engine. Specifically, the NOx emissions
can be maintained within 100 ppm, and the soot emissions below
10% at the full load ranges.
5. Conclusions
This study is dedicated to the fundamental research of the
combustion and emissions characteristics of compound HCCI
combustion fueled with gasoline and diesel blended fuel. Moreover,
this study attempts to investigate the potential of compound HCCI
combustion to achieve clean combustion at full load ranges with
the application of the intake air boost. Some conclusions can be
drawn as follows:
(1) Dieseline compound HCCI combustion exhibits a three-stage
combustion process. With an increase in the premixed ratio,
the peak values of the heat release rate of the LTR and HTR
increase while the peak value of the heat release rate of
diffusive combustion decreases. The maximum pressure,
mass-averaged temperature and pressure rising rate initially
X. Lu et al. / Energy 64 (2014) 707e718
717
Fig. 14. Comparison of emissions of G30 compound HCCI combustion with/without intake air boost with diesel/G30 DICI combustion.
decrease and then increase. The CO, HC and soot emissions
increase at first and then decrease while the NOx emissions
decrease monotonically.
(2) With an increase in the total fuel consumption per cycle, the
peak values of the heat release rate of LTR and HTR increase,
and the initial timing of LTR and HTR advances; CA10 and
CA50 advance, and the burn duration prolongs; the
maximum pressure and mass-averaged temperature gradually rise; the CO emissions increase at first and then
decrease; and the HC emissions decrease monotonically
while the NOx and soot emissions increase gradually.
(3) The effects of the intake air boost on G30 compound HCCI
combustion can also be separated into two parts: in the low
boost intensity condition, the dilution and inhibition effects
of the boost on the HCCI combustion occupies the dominant
position of the entire compound combustion; in the higher
boost intensity condition, the promotional effect of the boost
on diffusive combustion plays a primary role. Therefore, the
intake air boost can reduce the maximum pressure rising
rate and the maximum mass-averaged temperature. Moreover, with an increase in the intake air pressure, CA10 and
CA50 are initially retarded and then advanced while the burn
duration is progressively shortened. The intake air boost has
great potential to simultaneously reduce the NOx and soot
emissions of compound HCCI combustion while maintaining
the original CO and HC emission levels.
(4) Both G30 DICI combustion and G30 compound HCCI combustion with/without boost can achieve the full load of the
prototype engine. The CO and HC emissions of the optimized
G30 compound HCCI combustion with/without boost are
relatively higher compared to that of the prototype engine.
Compared to the prototype engine, G30 DICI combustion can
greatly reduce soot emissions but the NOx emission level is
quite similar. However, optimized G30 compound HCCI
combustion without boost can significantly reduce NOx
emissions but the level of soot emissions is similar to that of
the prototype engine. Neither of the combustion methods
could achieve the goal of reducing the NOx and soot emissions simultaneously. For the optimized G30 HCCI combustion with intake air boost, the NOx and soot emissions are far
lower than for the prototype engine. Specifically, the NOx
emissions could be maintained within 100 ppm, and the soot
emissions are below 10% at the full load ranges.
Ultra-low NOx and smoke emissions are obtained of dieseline
compound HCCI combustion using intake charge boost. However,
there have many works to be further conducted in the future. For
example, the combustion and emission characteristics of dieseline
advanced combustion process with boosting and exhaust gas
recirculation, the optimized gasoline volume in blends for different
advanced combustion modes, numerical simulation to reveal the
in-cylinder fuel/air mixture formation and distribution, ignition
mechanism, and emissions formation and evolution process, and
exploring the optimal control strategies for dieseline combustion
under high engine load (IMEP larger than 2 MPa).
Acknowledgments
This work was supported by the National Natural Science
Foundation of China (Grant No. 51176116) and National Basic
Research Program of China (Grant No. 2013CB228405).
718
X. Lu et al. / Energy 64 (2014) 707e718
References
[1] Onishi S, Jo SH, Shoda K, Jo DP, Kato S. Active thermo-atmosphere combustion
(ATAC)ea new combustion process for internal combustion engines. SAE Paper 790501, 1979.
[2] Noguchi M, Tanaka Y, Takeuchi Y. A study on gasoline engine combustion by
observation of intermediate reactive products during combustion. SAE Paper
790840, 1979.
[3] Zhao Fuquan, Asmus Thomas N, Assanis Dennis N, Dec John E, Eng James A,
Najt Paul M. Homogeneous charge compression ignition (HCCI) engines: key
research and development issues. Warrendale, PA: Society of Automotive
Engineers; 2003.
[4] Lu XC, Han D, Huang Z. Fuel design and management for the control of
advanced compression-ignition combustion modes. Prog Energ Combust Sci
2011;37:741e83.
[5] Santoso H, Matthews J, Cheng W. Managing SI/HCCI dual-Mode engine
operation. SAE Paper 2005-01-0162, 2005.
[6] Lu XC, Ji LB, Ma JJ, Zhou XX, Huang Z. Combustion characteristics and
influential factors of isooctane active-thermal atmosphere combustion
assisted by two-stage reaction of n-heptane. Combust Flame 2011;158:
203e16.
[7] Milovanovic N, Blundell D, Gedge S, Turner J. SI-HCCI-SI mode transition at
different engine operating Conditions. SAE Paper 2005-01-0156, 2005.
[8] Koopmans L, Ström H, Lundgren S, Backlund O. Demonstrating a SI-HCCI-SI
mode change on a volvo 5-cylinder electronic valve control engine. SAE Paper 2003-01-0753, 2003.
[9] Tian G, Wang Z, Ge Q, Wang J, Shuai S. Mode switch of SI-HCCI combustion on
a GDI Engine. SAE Paper 2007-01-0195, 2007.
[10] Zhang Y, Xie H, Zhao H. Investigation of SI-HCCI hybrid combustion and
control strategies for combustion mode switching in a four-stroke gasoline
engine. Combust Sci Technol 2009;181:782e99.
[11] Ganesh D, Nagarajan G. Homogeneous charge compression ignition (HCCI)
combustion of diesel fuel with external mixture formation. Energy 2010;35:
148e57.
[12] Pottker S, Eckert P, Delebinski T, Baumgarten C, Oehlert K, Merker GP, et al.
Investigations of HCCI combustion using multi-stage direct-injection with
synthetic fuels. SAE Paper 2004-01-2949, 2004.
[13] Shi L, Cui Y, Deng KY, Peng HY, Chen YY. Study of low emission homogeneous charge compression ignition (HCCI) engine using combined internal and external exhaust gas recirculation (EGR). Energy 2006;31:
2665e76.
[14] Feng HQ, Zheng ZQ, yao MF, Cheng G, Wang MY, Wang X. Effects of exhaust
gas recirculation on low temperature combustion using wide distillation
range diesel. Energy 2013;51:291e6.
[15] Soloiu V, Duggan M, Harp S, Vlcek B, Williams D. PFI (port fuel injection) of nbutanol and direct injection of biodiesel to attain LTC (low-temperature
combustion) for low-emissions idling in a compression engine. Energy
2013;52:143e54.
[16] Han D, Lu XC, Ma JJ, Ji LB, Huang C, Huang Z. Influence of fuel supply timing
and mixture preparation on the characteristics of stratified charge compression ignition combustion with n-heptane fuel. Combust Sci Technol 2009;181:
1327e44.
[17] Dec JE, Sjöberg M. Isolating the effects of fuel chemistry on combustion
phasing in an HCCI engine and the potential of fuel stratification for ignition
control. SAE Paper 2004-01-0557, 2004.
[18] Kumano K, Iida N. Analysis of the effect of charge inhomogeneity on HCCI
combustion by chemiluminescence measurement. SAE Paper 2004-01-0192,
2004.
[19] Richter M, Engström J, Franke A, Aldén M, Hultqvist A, Johansson B. The influence of charge inhomogeneity on the HCCI combustion process. SAE Paper
2000-01-2868, 2000.
[20] Yu R, Bai XS. Effect of turbulence and initial temperature inhomogeneity on
homogeneous charge compression ignition combustion. SAE Paper 2006-013318, 2006.
[21] Dec JE, Hwang W, Sjöberg M. An investigation of thermal stratification in HCCI
engines using chemiluminescence imaging. SAE Paper 2006-01-1518, 2005.
[22] Viggiano Annarita, Magi Vinicio. An investigation on the performance of
partially stratified charge CI ethanol engines. SAE Paper 2011-01-0837, 2011.
[23] Lu XC, Shen YT, Zhang YB, Zhou XX, Ji LB, Yang Z, et al. Controlled three-stage
heat release of stratified charge compression ignition (SCCI) combustion with
a two-stage primary reference fuel supply. Fuel 2011;90:2026e38.
[24] Lu XC, Zhou XX, Ji LB, Yang Z, Han D, Huang C, et al. Experimental studies on
the dual-fuel sequential combustion and emission simulation. Energy
2013;51:358e73.
[25] Ma JJ, Lu XC, Ji LB, Huang Z. Evaluation of SCCI potentials in comparison to
HCCI and conventional DICI combustion using n-heptane. Energ Fuel 2008;22:
954e60.
[26] Zhong S, Wyszynski M, Megaritis A, Yap D, Xu H. Experimental investigation
into HCCI combustion using gasoline and diesel blended fuels. SAE Paper
2005-01-3733, 2005.
[27] Weall A, Collings N. Investigation into partially premixed combustion in a
light-duty multi-cylinder diesel engine fuelled with a mixture of gasoline and
diesel. SAE Paper 2007-01-4058, 2007.
[28] Han D, Ickes AM, Bohac SV, Huang Z, Assanis DN. Premixed low-temperature
combustion of blends of diesel and gasoline in a high speed compression
ignition engine. Proc Combust Inst 2011;33:3039e46.
[29] Han D, Ickes AM, Assanis DN, Huang Z, Bohac SV. Attainment and load
extension of high-efficiency premixed low-temperature combustion with
dieseline in a compression ignition engine. Energ Fuel 2010;24:3517e25.
[30] Won HW, Pitsch H, Tait N, Kalghatgi G. Some effects of gasoline and diesel
mixtures on partially premixed combustion and comparison with the practical fuels gasoline and diesel in a compression ignition engine. Proc Inst Mech
Eng Part D J Automobile Eng 2012;226:1259e70.
[31] Zhang F, Xu H, Zeraati-Rezaei S, Kalghatgi G, Shuai S. Combustion and emission characteristics of a PPCI engine fuelled with dieseline. SAE Paper 201201-1138, 2012.
[32] Zhang F, Xu H, Zhang J, Tian G, Kalghatgi G. Investigation into light duty
dieseline fuelled partially-premixed compression ignition engine. SAE Paper
2011-01-1411, 2011.
[33] Turner D, Tian G, Xu H, Wyszynski M, Theodoridis E. An experimental study of
dieseline combustion in a direct injection engine. SAE Paper 2009-01-1101,
2009.
[34] Park SH, Youn IM, Lim Y, Lee CS. Influence of the mixture of gasoline and
diesel fuels on droplet atomization, combustion, and exhaust emission characteristics in a compression ignition engine. Fuel Proces Technol 2013;106:
392e401.
[35] Valentino G, Corcione FE, Iannuzzi SE. Effects of gasolineediesel and nbutanolediesel blends on performance and emissions of an automotive
direct-injection diesel engine. Int J Engine Res 2012;13:199e215.
[36] Heywood JB. Internal combustion engine fundamentals. New York: McGrawHill Book Company; 1988.
[37] Zhao H. HCCI and CAI engines for the automotive industry. Cambridge, UK:
Woodhead Publishing Limited; 2007.
[38] Ji LB, Lu XC, Ma JJ, Huang C, Han D, Huang Z. Experimental study on influencing factors of iso-octane thermo-atmosphere combustion in a dual-fuel
stratified charge compression ignition (SCCI) engine. Energ Fuel 2009;23:
2405e12.
[39] Ma JJ, Lu XC, Ji LB, Huang Z. An experimental study of HCCI-DI combustion and
emissions in a diesel engine with dual fuel. Int J Therm Sci 2008;47:1235e42.
[40] Sjöberg M, Dec JE. Combined effects of fuel-type and engine speed on intake
temperature requirements and completeness of bulk gas reactions in an HCCI
engine. SAE Paper 2003-01-3173, 2003.