International Journal of Automotive Technology, Vol. 14, No. 1, pp. 19−27 (2013)
DOI 10.1007/s12239−013−0003−5
Copyright © 2013 KSAE/ 069−03
pISSN 1229−9138/ eISSN 1976−3832
GASOLINE MULTIPLE PREMIXED COMPRESSION IGNITION (MPCI):
CONTROLLABLE, HIGH EFFICIENCY AND CLEAN COMBUSTION
MODE IN DIRECT INJECTION ENGINES
H. Q. YANG, S. J. SHUAI*, Z. WANG and J. X. WANG
State Key Laboratory of Automotive Safety and Energy, Tsinghua University, Beijing 100084, China
(Received 4 April 2012; Rrevised 12 July 2012; Accepted 19 September 2012)
ABSTRACT−A novel combustion concept namely “multiple premixed compression ignition” (MPCI) in gasoline direct
injection compression ignition (GDICI) regime is proposed. Its predominant feature is the first premixed and followed quasipremixed combustion processes in a sequence of “spray-combustion-spray-combustion”. The multiple-stage premixed
combustion decouples the pressure rise with pollutants formation process, which means the pressure rise rate and emissions
can be reduced simultaneously, while achieving a high thermal efficiency. The gasoline MPCI mode has been demonstrated
in a research engine with a compression ratio of 18.5. Gasoline with the research octane number (RON) of 94.4 was tested
under 1400 rpm, 0.6 MPa IMEP conditions, without EGR and intake boosting. A parameter study of common rail pressure and
intake temperature was implemented to investigate their effects on the performance of MPCI mode. Compared to the singlestage diffusion combustion in traditional diesel engines, the gasoline MPCI mode achieves lower emissions of soot, NO, CO,
as well as slightly higher indicated efficiency, with a penalty of higher THC emissions when the common rail pressure is larger
than 80 MPa in this study. With intake temperature sweeping, the gasoline MPCI mode also has the foregoing advantages
compared to the diesel under the same operating conditions.
KEY WORDS : Multiple premixed compression ignition (MPCI), High efficiency, Low pollutants, Common rail pressure,
Intake temperature
1. INTRODUCTION
compression ignition engine. The idea is to realize a separation between the end of injection (EOI) and start of
combustion (SOC) without using very high exhaust gas
recirculation (EGR), early injection and low compression
ratio. The engine is operated in gasoline partially premixed
compression ignition (PPCI) mode, where the stratification
of fuel concentration plays a key role, and high efficiency
and power density and low emissions have been achieved.
This concept has been researched subsequently by Hanson
et al. (2009), Shi and Reitz (2010), Dempsey and Reitz
(2011), Ra et al. (2011), Hildingsson (Hildingsson et al.,
2009; Hildingsson and Johansson, 2010), Manente et al.
(2009, 2010), Lewander et al. (2011), Weall and Collings
(2009), Zhang et al. (2011), Sellnau et al. (2011), and Dec
et al. (2011), on high octane number fuels in PPCI mode
with different EGR rates, different intake temperatures and
boosting pressures under different loads and speeds. The
results reported in the literatures have shown that the
gasoline PPCI mode has high potential to achieve lower
NOx and PM emissions while maintaining low indicated
specific fuel consumption (ISFC) compared to traditional
diesel engines. However, high CO and HC emissions at
high load still need to be controlled, and the high pressure
rise rate remains a big issue.
A novel combustion concept termed “multiple premixed
The thermal efficiency of traditional spark-ignition (SI)
gasoline engines is relatively low because of the low
compression ratio and high pumping losses and many
efforts have been done to improve the fuel economy of IC
engines. Homogeneous charge compression ignition
(HCCI) is well known for its high thermal efficiency and
low particulate matter (PM) and NOx emissions (Lee and
Huh, 2010). However, different from the traditional SI
gasoline engine whose heat release is controlled by flame
propagation and the conventional diesel engine whose heat
release is controlled by mixing process of fuel and air; the
HCCI engine is wholly governed by chemical kinetics. As
a result, the start of combustion (SOC) and heat release
process are hard to control. Furthermore, the biggest
challenge is perhaps to expand its operating range to high load
due to the unac--ceptable pressure rise rate and combustion
noise.
Kalghatgi (Kalghatgi et al., 2006, 2007, 2009, 2010)
have introduced a successful concept of HCCI using the
high octane fuels such as gasoline by direct injection into a
*Corresponding author. e-mail: sjshuai@tsinghua.edu.cn
19
20
H. Q. YANG, S. J. SHUAI, Z. WANG and J. X. WANG
compression ignition” (MPCI) for gasoline direct injection
engines working in the compression ignition (GDICI)
regime is proposed in this study. In order to build up a clear
difference between PPCI and MPCI mode, the spray events
and heat release profile of them are given in Figure 1 and
Figure 2. Obviously, the first premixed and followed quasipremixed combustion processes in a sequence of "spray—
combustion—spray—combustion" zaround the compression top dead center (TDC) is the predominant feature of
MPCI mode, rather than "spray—spray—combustion"
sequence in PPCI mode. The interlaced spray and
combustion events in MPCI mode can be repeated twice or
more, but it is better to fully separate each spray from
followed combustion without overlap so as to insure the
premixed combustion processes. The biggest advantage of
MPCI mode is that it decouples the pressure rise with the
pollutants formation process. In other words, the maximum
pressure rise rate (MPRR) can be tuned by optimizing the
injection times and dwell, as well as the split ratio and
injection timing; while the NOx and PM formation as well
as total hydrocarbon (THC) and CO can be regulated in
each premixed combustion process respectively. As a
result, this type of combustion mode overcomes the
difficulties in combustion control when a more homogeneous charge is formed to reduce NOx and PM emissions
in HCCI or PPCI mode at high engine load. The
combustion duration in MPCI mode can be longer because
the multiple spray and combustion events will control the
heat release rate. Although an overlong combustion
process is not preferable, the injection timing and dwell can
be optimized to obtain high thermal efficiency.
On the other hand, each couple of spray and premixed
combustion can be shorter due to multiple injections, and
the combustion region can be concentrated in the
combustion chamber. Consequently, the gap between the
combustion area and cylinder wall can be filled with air,
which means the THC and CO from incomplete
combustion and the heat loss from the chamber will be low.
The key point of the gasoline MPCI mode is to separate
each spray from the followed combustion to achieve a
premixed combustion one by one. In fact, the ignition delay
of the first spray will be longer than the followed ones.
Therefore the first combustion will be cleaner in an oxygen
rich environment with little NOx and soot formation. But
Figure 1. Schematic of spray events and heat release profile
in PPCI mode.
Figure 2. Schematic of spray events and heat release profile
in MPCI mode.
the followed gasoline will be injected into the burned
mixture of the previous combustion, which is hot and lack
of oxygen. However, thanks to the high octane number
(ON) and high volatility of gasoline, the followed combustions can be premixed or quasi-premixed. Furthermore, the
local temperature and equivalence ratio of followed
premixed combustion process will be higher and larger
than the first premixed combustion. As a result, the
expected path of the gasoline MPCI combustion in a φ-T
diagram (Kook et al., 2005) can be plotted in Figure 3.
Although the followed combustion takes place in a high
temperature and fuel rich atmosphere, the NOx and soot
emissions can be low after optimization of injection timing
and split ratio. It can be easily accepted that the first
injection is preferred to be more and earlier than the
followed ones in order to form more premixed-enough
charge, but the upper limit of MPRR is also a restraint of
the first premixed combustion just as that in HCCI and
PPCI mode.
Based on the analysis above, it is expected that the
gasoline MPCI has a potential to realize high efficiency
and low emissions simultaneously, especially at high load
engine conditions. However, because of the complexity of
the multi-stage spray and combustion characteristics of the
Figure 3. Expected path of the gasoline MPCI combustion
process in φ-T diagram.
GASOLINE MULTIPLE PREMIXED COMPRESSION IGNITION (MPCI)
gasoline MPCI mode, a number of parameters have to be
tuned in order to achieve low fuel consumption and
emissions. The authors have studied the influence of
injection timing and split ratios in the previous work (Yang
et al., 2012), and the effect of the common rail pressure and
intake temperature on the performance of gasoline MPCI
mode is going to be investigated in this paper. A single
cylinder compression ignition engine with the displacement of 0.7 liter and compression ratio of 18.5 retrofitted
from a four-cylinder light-duty diesel engine is used to
investigate the gasoline MPCI combustion mode. The
gasoline with the research octane number (RON) of 94.4 is
tested in this study, without EGR introducing and intake
boosting. Sweeping of common rail pressure and intake
temperature are carried out for gasoline in two-stage
premixed combustion mode. As comparative study, the
single-stage diffusion combustion mode of conventional
diesel engines under different common rail pressure, and
double injection strategy of diesel at various intake
temperature are also tested in this paper.
2. EXPERIMENTAL APPARATUS AND
METHOD
2.1. Test Engine
The test engine is a single cylinder compression ignition
engine with a displacement of 0.7 liter and compression
ratio of 18.5, mounted with a typical diesel high pressure
common rail injection system. It is retrofitted from a 4
cylinder 2-OHV (Over Head Valve) direct injection lightduty diesel engine and the first cylinder of the engine is
used for testing. The other three cylinders are deactivated
without fuel injection and the pistons of these cylinders are
holed to remove compression. The specifications of the
engine are listed in Table 1.
Table 1. Specifications of the test engine.
Engine model
SOFIM8140.43S3
Engine type
2-valve Compression Ignition
Bore [mm]
94.4
Stroke [mm]
100
Comprssion ratio
18.5
Cylinder number
4
Displacement [L]
2.8
Piston type
Piston bowl geometry
Inhector type
Articulated
ω-offset
Common Rail, Direct Injection
Max. power/Speed
[kW/r·min-1]
93 / 3600
Max. torque/Speed
[N·W/r·min-1]
290 / 1800
21
The engine is connected to an electric eddy current
dynamometer, and the original intake and exhaust manifolds
with a turbocharger are replaced by a naturally aspirated
intake and exhaust system for this fundamental combustion
mode study. So the intake pressure in this paper is
maintained in the ambient condition of around 0.1 MPa.
Moreover, an electric heater is fixed in the intake system to
regulate the inlet air temperature for the parametric study.
The engine electronic control unit (ECU) is changed to a
research module for flexible control of the fuel injections.
The experiments were carried out after the engine was
warmed up and reached a stable operating condition.
2.2. Test Fuels
The test gasoline fuel has a RON of 94.4 and a motor
octane number (MON) of 82.7, and is widely used in
Beijing market at present. It is well known that gasoline
has a lower lubricity than diesel. But in this study, no
lubricity additive is added into the gasoline fuel. The
purpose is to ensure that the fuel composition is just the
same as regular gasoline and reduce the impact of the
additive on combustion characteristics and emissions. In
fact, no obvious damage of the fuel injection system is
found after the experiments, which means it is possible to
use regular gasoline fuel in a typical common rail diesel
injection system without any lubricity additive in a
laboratory for a short time running. In order to give a
comparison with the gasoline MPCI combustion mode, the
0# diesel whose freezing points is 0oC also tests under the
same running conditions as the gasoline fuel in this study.
The 0# diesel has a cetane number (CN) of 52.6, and it is
also commonly used in Beijing now. The main properties
of the two fuels are listed in Table 2.
2.3. Test Conditions and Facilities
All the tests are carried out for the engine speed of
1400 rpm, load of 0.6 MPa IMEP condition, without the
intake boosting and EGR introducing. From the previous
Table 2. Fuel properties for testing.
93# Gasoline
0# Diesel
RON
94.4
---
MON
82.7
---
---
52.6
Density [kg/m ]
755.4
839.3
LHV [kJ/kg]
43500
43900
A/F stoich
14.02
14.30
Aromatics [v %]
CN
3
35.5%
29.6%
o
59.7
214.8
o
107.3
266.1
160.2
333.6
T10 [ C]
T50 [ C]
o
T90 [ C]
22
H. Q. YANG, S. J. SHUAI, Z. WANG and J. X. WANG
study (Yang et al., 2012), it is found that when the injection
dwell is 50o crank angle (CA), the fuel consumption and
emissions can be balanced well. Consequently, for the
study of the two-stage MPCI mode in this paper, the first
injection is located at -45o CA after top dead center
(ATDC) and the second injection at 5o CA ATDC. A split
ratio of 50%, which means that the fuel mass in the first
injection is 50% of the total, is used both for gasoline
MPCI mode and diesel double injection strategy when the
intake temperature sweeping.
The key issue for the gasoline MPCI is that the fuel and
air must be mixed well and rapidly and then a fast
combustion should be organized during each couple of
spray and combustion process. So the effect of common
rail pressure on the performance of gasoline MPCI mode is
investigated in this paper, with an intake temperature of
100oC to ensure the stable compression ignition of
gasoline. For a comparative study, the 0# diesel fuel is also
tested under the same conditions but adopting a single
injection strategy as in conventional diesel engines, with an
injection timing of -12o CA ATDC. Then, a sweeping of
intake temperature is also carried out, while the common
rail pressure is constant at 80 MPa. Double injection
strategy for diesel is adopted in this section, and all the test
conditions are just the same as that in the gasoline MPCI
mode when the intake temperature sweeping.
The in-cylinder pressure is measured using a Kistler
6125C pressure transducer mounted in the cylinder head
and recorded using an AVL IndiModul 621. The crankshaft
encoder used in this study is from Pepperl+Fuchs, whose
angular resolution is 0.5o CA. Combustion analysis data is
calculated based on the average cylinder pressure of 50
consecutive cycles. The exhaust gas emissions are
measured using a Horiba MEXA-584L gas analyzer,
including CO, CO2, O2, NO, THC and exhaust lambda
values. Smoke measurements are taken in BSU (Bosch
Smoke Unit) using a FBY-1 type unheated sample smoke
meter which uses the filter paper sampling method. Fuel
consumption is measured using an ONO SOKKI DF-313
Table 3. Main experimental apparatus employed in this
study.
Equipment
Manufacturer
Model
Dynamometer
ZÖLLNER
B-220AC
In-cylinder
Pressure Sensor
KISTLER
6125C
Crankshaft Encoder
Combustion Analyzer
Gaseous Pollutants
Analyzer
Smoke Meter
Fuel Consumption
Meter
PEPPERL+FUCHS TRD-J 720-RT2
AVL
IndiModul 621
HORIBA
MEXA-584L
FOFEN
FBY-1
ONO SOKKI
DF-313
Figure 4. Schematic of the experimental apparatus.
digital fuel meter. The main experimental apparatus
employed are listed in Table 3, and the schematic of the
experimental apparatus is shown in Figure 4.
3. RESULTS AND DISCUSSION
3.1. Performance under Different Common Rail Pressure
The mean gas temperature in the cylinder is something like
the “cylinder bulk temperature” (Christensen et al., 1999),
which is calculated from the cylinder pressure trace. It is
assumed that the temperature and the gas composition is
the same in the whole bulk, and this is a quite good
assumption with HCCI. Although the temperature of
gasoline MPCI mode is not homogeneous in the whole
cylinder, the mean gas temperature can reveal the main
feature of each premixed combustion process, and it is
closely relative to the formation of NOx and soot.
Calculation of the heat release rate in this paper is
according to the first law of thermodynamics, and the
interval of real time calculation is 1o CA. The calculation is
based on a constant polytropic coefficient of 1.37 to
simulate the property of fuel and air mixture, as well as the
combustion products in this study. An ideal gas assumption
is made in the calculation of mean gas temperature. Then,
the mean gas temperature is derived from the equation of
state, where the gas mass is set as the intake air mass and
the fuel mass is ignored in this simplified calculation.
Figure 5 and Figure 6 give the mean gas temperature and
the heat release rate profile of the gasoline MPCI and diesel
single injection mode under different common rail
pressure, respectively. It is clear that the mean gas
temperature is affected by the SOC significantly, and the
GASOLINE MULTIPLE PREMIXED COMPRESSION IGNITION (MPCI)
Figure 5. Mean gas temperature and heat release rate of
gasoline MPCI mode under different common rail pressure,
and 1400 rpm, 0.6 MPa IMEP, 100oC intake temperature
conditions.
Contrarily, the ignition delay of diesel single injection
mode becomes shorter when the common rail pressure
rises. This is perhaps because of the different viscosity of
gasoline and diesel. As is known to all, under the same
temperature and pressure conditions, the stoichiometric
mixture of fuel and air has the biggest tendency to get autoignition. So because of the large viscosity of diesel fuel, the
liquid core of diesel spray is hard to break up and evaporate
to form a stoichiometric mixture. As a result, if the
common rail pressure becomes higher, it is much easier for
diesel spay to produce a stoichiometric mixture with air,
and this promotes the auto-ignition process of diesel. On
the contrary, because of the high volatility and low
viscosity of gasoline, the higher common rail pressure
makes the mixture leaner than stoichiometric charge,
which prolongs the ignition delay time of gasoline.
The indicated efficiency is evaluated from the fuel flow
and the indicated mean effective pressure during the
compression and expansion stroke only (Christensen et al.,
1999). Pumping work and engine friction is not included
here. Defined on one cycle only, this efficiency can be
expressed as the ratio between the work on the piston
during the compression and expansion stroke only, Wi and
the input fuel energy:
Wi
ηi = --------------= ηc ⋅ ηt, i
mf ⋅ qlhv
Figure 6. Mean gas temperature and heat release rate of
diesel single injection mode under different common rail
pressure, and 1400 rpm, 0.6 MPa IMEP, 100oC intake
temperature conditions.
gasoline MPCI mode has a lower peak temperature than
diesel single injection mode due to a multiple-stage heat
release process. Though the local peak temperature will be
higher than that of mean gas temperature, from Figure 5 we
can still find that it is possible to produce a higher
temperature in the second premixed combustion process
than that of the first one, especially at higher common rail
pressure. This phenomenon supports the expected pathway
of gasoline MPCI mode in Figure 3 to a certain extent,
which demonstrates that the gasoline MPCI combustion
mode is promising to reduce the NOx and soot emissions
simultaneously while maintaining relatively high fuel
efficiency.
Another remarkable trend in Figure 5 and Figure 6 is
that the SOC of the first premixed combustion of gasoline
gets later and later as common rail pressure increasing.
23
(1)
Where mf is fuel mass per cycle and qlhv is the lower
heating value per mass unit fuel; ηc is the combustion
efficiency, which is evaluated from the exhaust gas
composition and it is a measure of how complete the
combustion is. The indicated thermal efficiency, ηt,i, is
defined as the ratio between the work on the piston during
the compression and expansion stroke only, Wi, and the
heat release, Q:
W η
ηt, i = ------i = -----i
Q ηc
(2)
Figure 7 gives the ISFC and indicated efficiency of
gasoline MPCI and diesel single injection mode as the
common rail pressure sweeping. Obviously, the indicated
efficiency of gasoline MPCI mode is very sensitive to the
common rail pressure compared with diesel single injection
mode. It is mainly because that the common rail pressure
affects the combustion phasing of gasoline MPCI mode
more significantly than that of diesel, and this can be
observed in Fiugre 5 and Figure 6. At lower common rail
pressure, the first premixed combustion process of gasoline
happens too early, which leads to a marked negative
compression work and results in a poor fuel efficiency.
However, when the common rail pressure is larger than
90 MPa, the reasonable ignition timing of the first
premixed combustion and the rapid heat release process of
the second premixed combustion in gasoline MPCI mode
make its fuel efficiency higher than that of diesel. In addition,
the higher combustion efficiency of the multiple-stage
24
H. Q. YANG, S. J. SHUAI, Z. WANG and J. X. WANG
premixed combustion process in gasoline MPCI mode, and
the low heat transfer loss in the first premixed combustion
process resulting from the lower peak temperature of
gasoline MPCI mode, also lead to a higher indicated
efficiency than the single-stage diffusion combustion
process of traditional diesel engines.
The indicated specific NO and soot emission of gasoline
MPCI and diesel single injection mode as the common rail
pressure sweeping are shown in Figure 8. It is encouraging
to see that different from the trade-off relationship between
NO and soot in conventional diesel diffusion combustion
mode, the gasoline MPCI mode has very low NO and soot
emissions especially when the common rail pressure is
larger than 80 MPa. This phenomenon proves the possibility of the gasoline MPCI pathway in Figure 3, once again.
At 40 MPa common rail pressure, the soot emission of
gasoline MPCI mode is high. It is because at this common
rail pressure, the gasoline has the shortest ignition delay
time, which results in a poorer mixing process and a higher
peak temperature of the first premixed combustion. The
weak oxidation of soot due to a lower mean gas
temperature of the second premixed combustion at 40 MPa
common rail pressure is another reason of the high soot
emission. In spite of this, the high volatility and octane
rating of gasoline make it easy to produce premixed charge,
and the multiple-stage premixed combustion process
decouples the pressure rise and pollutants formation
process. With careful optimization of the injection parameters,
it is possible to achieve very low NO and soot emissions
simultaneously, while maintaining relative high fuel
efficiency.
Figure 9 gives the indicated specific THC and CO
emissions of gasoline MPCI and diesel single injection
mode as the common rail pressure sweeping. Under high
common rail pressure conditions, the lower combustion
temperature of the first lean and premixed combustion
process in gasoline MPCI mode prevents the NO and soot
formation, but the combustion temperature becomes too
low to oxidize the fuel completely. This low combustion
temperature results in higher emissions of unburned
hydrocarbons than that in diesel single injection mode. The
combustion temperature near the walls will be even lower,
due to heat losses. Combustion may be quenched or not
occur at all close to the walls (Fan et al., 2012; Jung et al.,
2012). However, the THC emission of gasoline MPCI
mode will lower than that of HCCI mode. This is because in
the gasoline HCCI mode the charge is nearly homogeneous
all around in the cylinder, and the homogeneous charge is
generally fuel lean to avoid an unacceptable high MPRR.
As a result, the local combustion temperature will be even
lower and more unburned hydrocarbons will be produced
especially near the cylinder wall.
Another notable phenomenon is that at 40 MPa common
rail pressure, extremely high CO is observed in gasoline
MPCI mode, just as the trend of soot emission in Figure 8.
This phenomenon implies that maybe a wall impingement
occurs at 40 MPa common rail pressure, because of the
Figure 8. Indicated specific NO and soot emissions of
gasoline MPCI and diesel single injection mode as the
common rail pressure sweeping, under 1400 rpm, 0.6 MPa
IMEP, 100oC intake temperature conditions.
Figure 9. Indicated specific THC and CO emissions of
gasoline MPCI and diesel single injection mode as the
common rail pressure sweeping, under 1400 rpm, 0.6 MPa
IMEP, 100oC intake temperature conditions.
Figure 7. ISFC and indicated efficiency of gasoline MPCI
and diesel single injection mode as the common rail
pressure sweeping, under 1400 rpm, 0.6 MPa IMEP, 100oC
intake temperature conditions.
GASOLINE MULTIPLE PREMIXED COMPRESSION IGNITION (MPCI)
longer spray penetration of gasoline in this condition. This
is different from the situation of diesel. Thanks to the high
viscosity and low volatility of diesel, higher common rail
pressure will lead to longer spray penetration. But for the
gasoline fuel, higher common rail pressure may make the
atomization process much quicker and more gasoline will
vaporize near the nozzle hole at higher common rail
pressure. This phenomenon may reduce the momentum of
the liquid gasoline and lead to a shorter spray penetration.
Fortunately, the THC and CO can be easily converted into
H2O and CO2 using diesel oxidation catalyst with low cost.
This measure ensures the gasoline MPCI concept to be a
high efficiency and controllable combustion mode, with
the potential to meet the increasingly stringent emission
standards in the future.
3.2. Performance under Different Intake Temperature
Similar to the above analysis, the combustion characteristics,
indicated efficiency, and emission performance of gasoline
MPCI and diesel double injection mode as the intake
temperature sweeping are discussed in this section. Figure
10 and Figure 11 give the mean gas temperature and the
heat release profile of these two combustion modes under
different intake temperature conditions. No doubt, no
matter the gasoline or diesel, their SOC become earlier as
the intake temperature rising. And the ignition timing of
gasoline is more sensitive to the intake temperature than
that of diesel because of the poorer ignitability of gasoline.
Another notable phenomenon in Figure 11 is that when
applying double injection strategy, the diesel combustion
process also achieves thorough separation between the
injection and heat release events. However, it is harder for
diesel to form premixed combustion process than gasoline,
owing to the bad volatility and big viscosity of diesel fuel.
What’s more, the higher cetane number results in a shorter
ignition delay, so there is less time for diesel to produce
premixed charge which leads to a larger concentration
Figure 10. Mean gas temperature and heat release rate of
gasoline MPCI mode under different intake temperature, and
1400 rpm, 0.6 MPa IMEP, 80 MPa common rail pressure
conditions.
25
Figure 11. Mean gas temperature and heat release rate of
diesel double injection mode under different intake
temperature, and 1400 rpm, 0.6 MPa IMEP, 80 MPa
common rail pressure conditions.
gradient of the mixture and local high temperature
benefiting the formation of NOx and soot.
In addition, the premature heat release of the first
injection of diesel makes a large negative compression
work and a bad fuel efficiency shown in Figure 12. Besides
the combustion phasing, the specific heat ratio is another
important factor which affects the indicated thermal
efficiency in IC engines. Without intake boosting, the
higher intake temperature means a lower inlet air density,
and a smaller excessive air coefficient. As is known to all,
the rich mixture has a smaller specific heat ratio and
indicated thermal efficiency. This is just the reason of the
reduction in fuel efficiency both for gasoline MPCI and
diesel double injection mode as the intake temperature
increasing in Figure 12.
The indicated specific emissions of NO, soot, and THC,
CO of the gasoline MPCI and diesel double injection mode
as intake temperature sweeping are shown in Figure 13 and
Figure 12. ISFC and indicated efficiency of gasoline MPCI
and diesel double injection mode as the intake temperature
sweeping, under 1400 rpm, 0.6 MPa IMEP, 80 MPa common
rail pressure conditions.
26
H. Q. YANG, S. J. SHUAI, Z. WANG and J. X. WANG
solution of this contradiction is to use low octane gasoline
fuels, which have high volatility like gasoline but a more
reasonable ignitability. By using low octane gasoline in the
MPCI combustion mode, intake air preheating is not
necessary; so the pressure rise rate and the emission
performance can be balanced better than that of high octane
gasoline, with a higher fuel efficiency at the same time.
4. CONCLUSION
Figure 13. Indicated specific NO and soot emissions of
gasoline MPCI and diesel double injection mode as the
intake temperature sweeping, under 1400 rpm, 0.6 MPa
IMEP, 80 MPa common rail pressure conditions.
Figure 14. Indicated specific THC and CO emissions of
gasoline MPCI and diesel double injection mode as the
intake temperature sweeping, under 1400 rpm, 0.6 MPa
IMEP, 80 MPa common rail pressure conditions.
Figure 14, respectively. Similar to the results in section 3.1,
the NO, soot and CO emissions of gasoline MPCI mode
are lower than that of the diesel double injection mode
especially at lower intake temperature, with a penalty of
slightly higher THC emission. The unusual high CO
emission of gasoline MPCI mode at 60oC intake
temperature is mainly caused by the late SOC of the first
premixed combustion of gasoline and consequently the
overlap of the second injection and the first combustion
process. It reveals the importance to separate the second
spray from the first premixed combustion process as much
as possible in gasoline MPCI mode. With the intake
temperature getting higher, the NO emission rises mainly
because of the higher combustion temperature, and the soot
and CO emissions increase mainly due to the lack of
oxygen. This trend proves that the intake temperature of
the gasoline MPCI mode should not be so high considering
the emission performance. However, if taking the
ignitability of gasoline and running stability into account,
the higher intake temperature is necessary. An ideal
A new combustion mode of multiple premixed compression ignitions (MPCI) for gasoline direct injection engines
is proposed in this study. It has been demonstrated in a
single cylinder engine with a high compression ratio of
18.5 and a displacement of 0.7 liter. The indicated
efficiency and emissions of the gasoline MPCI combustion
mode have been analyzed and compared with the
commercial 0# diesel single and double injection mode
under the same running conditions. A parameter study of
common rail pressure and intake temperature are carried
out to investigate their effects on the performance of
gasoline MPCI combustion mode.
When the common rail pressure is higher than 80 MPa,
the NO, soot and CO emissions, as well as the ISFC of
gasoline MPCI mode are lower than that of the diesel
single injection mode, with a penalty of slightly higher
THC emission. What’s more, the ignition delay of the first
premixed combustion in gasoline MPCI mode is longer
when the common rail pressure gets higher, and it is
contrary to the trend in diesel single injection mode. The
indicated efficiency of the gasoline MPCI mode is more
sensitive to the common rail pressure than that of the diesel
single injection mode, mainly because of the change on
combustion phasing.
It is possible to achieve a thorough separation between
the second injection and the first combustion process when
applying double injection strategy for diesel. However,
because of the poor volatility and premature ignition of
diesel, the indicated efficiency and emissions performance
is still worse than that of gasoline MPCI mode. As the
intake temperature getting higher, a deterioration in fuel
efficiency and emissions both of gasoline MPCI and diesel
double injection mode is observed. It is mainly caused by
the high combustion temperature and lower excessive air
coefficient in higher intake temperature condition.
In a word, the gasoline MPCI mode decouples the
pressure rise process with the pollutants formation process;
and with carefully optimization of the injection parameters,
it is possible to realize the directly controllable, efficient
and clean combustion mode in GDICI engines. Moreover,
thanks to the high volatility and feasible ignitability of low
octane gasoline fuel, its MPCI mode will be much easier to
realize and will have better fuel efficiency and emissions
performance than the high octane gasoline, due to the lack
of intake preheating. Hence, the MPCI combustion mode
of low octane gasoline fuel needs to be investigated in
GASOLINE MULTIPLE PREMIXED COMPRESSION IGNITION (MPCI)
further studies.
ACKNOWLEDGEMENT−The authors would like to acknowledge the National Natural Science Foundation of China (NSFC
51276097), the Key Project of the National Natural Science
Foundation of China (NSFC 51036004) and the Joint Research
Program between General Motors Corporation and Tsinghua
University for funding this research.
REFERENCES
Christensen, M., Hultqvist, A. and Johansson, B. (1999).
Demonstrating the multi fuel capability of a homogeneous charge compression ignition engine with variable
compression ratio. SAE Paper No. 1999-01-3679.
Dec, J. E., Yang, Y. and Dronniou, N. (2011). Boosted
HCCI - controlling pressure-rise rates for performance
improvements using partial fuel stratification with
conventional gasoline. SAE Paper No. 2011-01-0897.
Dempsey, A. B. and Reitz, R. D. (2011). Computational
optimization of a heavy-duty compression ignition
engine fueled with conventional gasoline. SAE Paper
No. 2011-01-0356.
Fan, Q., Bian, J., Lu, H., Li, L. and Deng, J. (2012). Effect
of the fuel injection strategy on first-cycle firing and
combustion characteristics during cold start in a TSDI
gasoline engine. Int. J. Automotive Technology 13, 4,
523−531.
Hanson, R., Splitter, D. and Reitz, R. D. (2009). Operating
a heavy-duty direct-injection compression-ignition
engine with gasoline for low emissions. SAE Paper No.
2009-01-1442.
Hildingsson, L., Kalghatgi, G. T., Tait, N., Johansson, B.
and Harrison, A. (2009). Fuel octane effects in the
partially premixed combustion regime in compression
ignition engines. SAE Paper No. 2009-01-2648.
Hildingsson, L. and Johansson, B. (2010). Some effects of
fuel autoignition quality and volatility in premixed
compression ignition engines. SAE Paper No. 2010-010607.
Jung, G. S., Sung, Y. H., Choi, B. C., Lee, C. W. and Lim,
M. T. (2012). Major sources of hydrocarbon emissions
in a premixed charge compression ignition engine. Int. J.
Automotive Technology 13, 3, 347−353.
Kalghatgi, G. T., Risberg, P. and Ångström, H. E. (2006).
Advantages of fuels with high resistance to auto-ignition
in late-injection, low-temperature, compression ignition
combustion. SAE Paper No. 2006-01-3385.
Kalghatgi, G. T., Risberg, P. and Ångström, H. E. (2007).
Partially pre-mixed auto-ignition of gasoline to attain
low smoke and low NOx at high load in a compression
ignition engine and comparison with a diesel fuel. SAE
27
Paper No. 2007-01-0006.
Kalghatgi, G. T., Hildingsson, L. and Johansson, B. (2009).
Low NOx and low smoke operation of a diesel engine
using gasoline-like fuels. ASME ICES 2009-76034.
Kalghatgi, G. T., Hildingsson, L., Harrison, A. and
Johansson, B. (2010). Low- NOx, low-smoke operation
of a diesel engine using “premixed enough” compression
ignition – Effects of fuel autoignition quality, volatility
and aromatic content. THIESEL 2010.
Kook, S., Bae, C., Miles, P. C., Choi, D. and Pickett, L. M.
(2005). The influence of charge dilution and injection
timing on low-temperature diesel combustion and
emissions. SAE Paper No. 2005-01-3837.
Lee, Y. J. and Huh, K. Y. (2010). Simulation of HCCI
combustion with spatial inhomogeneities via a locally
deterministic approach. Int. J. Automotive Technology
11, 1, 19−26.
Lewander, C. M., Johansson, B. and Tunestal, P. (2011).
Extending the operating region of multi-cylinder partially
premixed combustion using high octane number fuel.
SAE Paper No. 2011-01-1394.
Manente, V., Johansson, B. and Tunestal, P. (2009). Effects
of different type of gasoline fuels on heavy duty partially
premixed combustion. SAE Paper No. 2009-01-2668.
Manente, V., Zander, C. G., Johansson, B. and Tunestal, P.
(2010). An advanced internal combustion engine
concept for low emissions and high efficiency from idle
to max load using gasoline partially premixed
combustion. SAE Paper No. 2010-01-2198.
Ra, Y., Loeper, P., Reitz, R. D. and Andrie, M. (2011). Study
of high speed gasoline direct injection compression
ignition (GDICI) engine operation in the LTC regime.
SAE Paper No. 2011-01-1182.
Sellnau, M., Sinnamon, J., Hoyer, K. and Husted, H.
(2011). Gasoline direct injection compression ignition
(GDCI) - Diesel-like efficiency with low CO2 emissions.
SAE Paper No. 2011-01-1386.
Shi, Y. and Reitz, R. D. (2010). Optimization of a heavyduty compression-ignition engine fueled with diesel and
gasoline-like fuels. Fuel, 89, 3416−3430.
Weall, A. and Collings, N. (2009). Gasoline fuelled
partially premixed compression ignition in a light duty
multi cylinder engine: a study of low load and low speed
operation. SAE Paper No. 2009-01-1791.
Yang, H. Q., Shuai, S. J., Wang, Z. and Wang, J. X. (2012).
High efficiency and low pollutants combustion:
Gasoline multiple premixed compression ignition
(MPCI). SAE Paper No. 2012-01-0382.
Zhang, F., Xu, H. M. and Zhang, J. (2011). Investigation into
light duty dieseline fuelled partially-premixed compression
ignition engine. SAE Paper No. 2011-01-1411.