International journal of automotive technology, tập 10, số 1, 2009

International Journal of Automotive Technology, Vol. 10, No. 1, pp. 1−7 (2009)
DOI 10.1007/s12239−009−0001−9

Copyright © 2009 KSAE
1229−9138/2009/044−01

EFFECTS OF MIXTURE STRATIFICATION ON HCCI COMBUSTION OF
DME IN A RAPID COMPRESSION AND EXPANSION MACHINE
G. S. JUNG1), Y. H. SUNG1), B. C. CHOI2) and M. T. LIM2)*
1)

Graduate School of Mechanical Engineering, Chonnam National University, Gwangju 500-757, Korea
School of Mechanical Systems Engineering, Chonnam National University, Gwangju 500-757, Korea

2)

(Received 8 April 2008; Revised 10 June 2008)
ABSTRACT−Compression ignition of homogeneous charges in internal combustion (IC) engines is expected to offer high
efficiency of DI diesel engines without high levels of NOx and particulate emissions. This study is intended to find ways of
extending the rich limit of HCCI operation, one of the problems yet to be overcome. Exhaust emissions characteristics are also
explored through analyses of the combustion products. DME fuel, either mixed with air before induction or directly injected
into the combustion chamber of a rapid compression and expansion machine, is compressed to ignite under various conditions
of compression ratio, equivalence ratio, and injection timing. The characteristics of the resulting combustion and exhaust
emissions are discussed in terms of the rate of heat release computed from the measured pressure, and the concentrations of
THC, CO, and NOx are measured by FT-IR and CLD. The experimental data to date show that operation without knock is
possible with mixtures of higher equivalence ratio when DME is directly injected rather than when it is inducted in the form
of a perfectly homogeneous fuel-air mixture. Although fuel injected early in the compression stroke promotes homogeneity
of the DME-air mixture in the cylinder, it causes the mixture to ignite too early to secure good thermal efficiency and knockfree operation at high loads. Low temperature reactions occur at about 660K regardless of the fueling methods, fuel injection
timing and equivalence ratio. The main components of hydrocarbon emissions turned out to be unburned fuel (DME),
formaldehyde and methane.

KEY WORDS : DME, HCCI, RCEM, Exhaust emissions, Perfectly homogeneous fuel-air mixture, Direct fuel injection

1. INTRODUCTION

2005) have been suggested to resolve this difficulty by
relaxing almost simultaneous heat release of HCCI combustion. In actual engines, various causes for non-homogeneity exist including imperfect mixing of fuel, air, and
residual or EGR gas, differences in gas flow and boiling
points of fuel blends, and varying levels of heat transfer
along the cylinder walls.
The objective of this study is to investigate the effects of
stratification caused by uneven fuel-air mixing on the HCCI
combustion and consequent emissions. Different ways of
adding fuel to air are tried including premixing and direct
injection at various timing. A rapid compression and expansion machine (RCEM) is used in this study in order to
focus on the effects of fuel-air mixing caused stratification
with minimized contribution of other factors like residual
gas and gas motion.
DME (di-methyl ether) is the selected test fuel, since its
combustion-related properties are close to those of diesel
fuel, and it burns cleanly without generating soot. DME
also exhibits two-stage heat release, which is one of the
distinctive characteristics of HCCI combustion (Ogawa et
al., 2003; Teng et al., 2004; Yao et al., 2003).

New technologies that increase thermal efficiency and clean
exhaust gas of automobile engines are desperately being
sought to mitigate the problems of energy shortage & environmental pollution. Homogeneous charge compression
ignition (HCCI) is one of those representative technologies
being developed in the field of engine combustion. HCCI
engines are expected to have higher efficiency than SI

engines due to their high compression ratio and dispensability of throttle valves. Compared to diesel engines, they
emit less particulate matters and less NOx because only
lean premixed combustion without local fuel-rich zones is
present (Gray and Ryan, 1997; Thring, 1989; Chung et al.,
2008). Despite these advantages, commercial mass production of HCCI engines have not yet been realized. One of
the obstacles to the commercialization of HCCI engines is
objectionable knock occurring under heavy load conditions.
Knock results from an excessive rate of pressure rise in the
combustion chamber (Gray and Ryan, 1997). Methods of
utilizing mixture stratification (Inagaki et al., 2006; Kumano
and Iida, 2004; Sjöberg and Dec, 2006) and thermal stratification (Dec et al., 2006; Lim et al., 2006; Sjöberg et al.,

2. EXPERIMENTAL SYSTEM AND TEST SCOPE
As shown in Figure 1, the experimental system consists of

*Corresponding author. e-mail:
1


2

G. S. JUNG, Y. H. SUNG, B. C. CHOI and M. T. LIM

Table 1. Specifications of RCEM.

Figure 1. Schematic diagram of the experimental system.

Item

Specification

Combustion chamber
Bore/Stroke
Displacement volume
Connecting rod length
Crank Radius
Compression ratio
Top clearance height

Disc type
100/450 mm
3.534 dm3
900 mm
225 mm
8~23
17.5~61.8 mm

Table 2. Specifications of fuel injector.
Item

Specification

Valve

Lift

45~50 μm

Nozzle

Open pressure
Needle lift
Seat diameter
Nozzle holes

5~6 MPa
0.35 mm
Ø 2.25
6 × Ø 0.22

Table 3. Experimental conditions.
Premixed Direct injection
Figure 2. Definitions of LTR and HTR.
an RCEM and subsystems for fuel supply and injection, for
heating cylinder charge, and for acquisition of pressure and
temperature signals. Basic structure of the RCEM (specifications shown in Table 1) is similar to that of a reciprocating piston engine. The RCEM is a single-shot engine,
whose crankshaft is driven by the kinetic energy stored in a
large flywheel. Before the combustion test, residual gas is
removed from the combustion chamber by a vacuum pump,
and fresh air is charged. Design and specifications for this
RCEM are explained elsewhere in detail (Cho et al., 2004).
The initial temperature of the air or fuel-air mixture in the
combustion chamber of the RCEM is controlled by operating the electric heater wrapped around the cylinder head,
and also by circulating heated water through the jacket
around the cylinder.
The DME fuel is supplied to the combustion chamber
either premixed with air or by direct injection. Fuel and air
are mixed and stabilized for perfectly homogeneous charge
tests approximately a day before use. Direct injection of
liquid DME is accomplished at 50 MPa by controlling the

timing and duration of the injector open time. Table 2
provides the major specifications of the fuel injector in
detail. An in-house designed computer program performs
the injection control based on the crank angle signal from a
rotary encoder (Autonics, E40S-360-3-5). Since viscosity
of DME is insufficient for self-lubrication of the nozzle
(0.33 cSt at 293 K and 50 MPa), 750 ppm of viscosity
improver (ETHYL, hitec-4140, 17.2 cSt) is added to DME
(Longbao et al., 1999).

Fuel
Revolution speed
Compression ratio
Initial temperature
Fuel supply pressure
Equivalence ratio, ø
Injection timing

DME
100 rpm
12
303 K
−
0.1, 0.2,
0.25, 0.3
−

50 MPa
0.1, 0.2, 0.25,
0.3, 0.35, 0.4

BDC, 150°, 120°,
90°, 60°, 30° BTDC

Combustion tests were run with fuel introduced either
premixed or directly injected into the combustion chamber
under various conditions of equivalence ratio and injection
timing as described in Table 3.
Combustion characteristics are analyzed using the pressure in the combustion chamber taken by a piezoelectric
sensor (Kistler, 6016B) and a charge amplifier (Kistler,
5011A). Exhaust gas or combustion product is displaced
out of the cylinder and fed to the analyzers after each run of
the combustion tests. Engine exhaust gas analyzers are
inapplicable to RCEM tests because RCEM conducts only
one cycle at a time, producing a fixed volume of the exhaust gas (about 3.5 dm3 in the present case) which is too
small to keep a continuous sample flow to the analyzers.
Concentrations of unburned hydrocarbon (HC) and carbon
monoxide (CO) are therefore measured by an FT-IR
(MIDAC, I2004) that requires about 0.3 dm3 of sample gas,
and oxides of nitrogen (NO and NOx) are measured by a
CLD (Thermo Environmental Instruments, 42C-HL) that
requires about 25 cc/min of sample gas.


EFFECTS OF MIXTURE STRATIFICATION ON HCCI COMBUSTION OF DME IN A RAPID COMPRESSION

3

3. EXPERIMENTAL RESULTS AND
DISCUSSIONS
3.1. Characteristics of HCCI Combustion

LTR (low temperature reactions) and HTR (high temperature reactions) in HCCI combustion are identified in this
study as the first and the second phases of the oxidation
process on ROHR (rate of heat release) curves as depicted
in Figure 2. The particular curve in this example is the one
actually obtained from the pressure record of a combustion
test using a perfectly premixed DME-air mixture with equivalence ratio of 0.2. The LTR and HTR are found to be
associated with formation of formaldehyde and formation
of CO and CO2 respectively (Hamada et al., 2005).
Figure 3(a) shows a group of curves representing respectively the motored cylinder pressure, and fired cylinder pressure along with the associated ROHR and the gas temperature during a test run at an equivalence ratio of 0.1. This
figure gives an overall idea about what happens during the
entire period of compression and expansion strokes in the
combustion chamber during the test. Figure 3(b) shows the
similar sets of data for the combustion tests run at equivalence ratios of 0.1 and 0.3, but the range of data is
restricted only to the late phase of compression and early
phase of expansion. The blue curves in the latter figure
denote the results from the tests run with perfectly homogeneous fuel-air mixture supplied, while the red lines denote
those run with fuel directly injected at the start of compression (i.e. at bottom dead center). The combustion pressure in figure (a) gradually rises at first during the compression, just tracing the motored cylinder pressure before
the start of reactions at about 19°BTDC, and then it rapidly
increases thereafter. The in-cylinder gas completes combustion, and its temperature reaches 930 K at TDC. The
two peaks of LTR and HTR appear after ignition on ROHR
curve. Although auto-ignitions are observed to occur at
about 20o~15o BTDC, in all the cases shown in Figure 3,
the early burning rate in the cases of the premixed combustion is much larger, resulting in a shorter combustion
duration and higher peak pressure. On comparison of the
cases in Figure 3(b) where the equivalence ratio is 0.3, the
cylinder pressure during combustion rises fast enough to
cause obvious knock in the premix tests as characterized by
the sharp peak and the fluctuations in the pressure curve,
but knock is not observed in the case of direct injection test.
Figure 4 shows the maximum rate of combustion pressure measured in tests with various equivalence ratio and

fueling strategies. In the cases of premixed mixture the
maximum rate of pressure rise exceeds 4.3 MPa/ms (Lim et
al., 2006), which is taken as knock borderline in this study,
at equivalence ratios over 0.25. However, knock is not seen
to take place at equivalence ratios of up to 0.4 when fuel is
directly injected at various injection timing. This implies
that the rich operation limit of HCCI combustion can be
extended by direct injection of fuel. The effect may be
attributed to the stratification of the mixture and latent heat

Figure 3. Cylinder pressure, temperature and rate of heat
release as a function of crank angle for DME-air mixtures
of different equivalence ratios and mixing processes.
of vaporization. The maximum rate of pressure rise seems
to gradually increase as either the equivalence ratio increases,
or the fuel injection is retarded at equivalence ratio over
0.35.
Figure 5 represents the timing and duration of LTR and
HTR in the tests with various fuel injection timing and
equivalence ratios. The left, blue portions of the bars indicate time for LTR while the right, red portions indicate this
for HTR. Although the ignition delay is significantly shortened, the start of LTR and HTR is delayed less than 5° CA
as the fuel injection is retarded from BDC to 30° BTDC.
When compared with the premixed fuel-air mixtures, the
ignition starts later, and combustion lasts longer when
mixtures are formed through direct injection. The ignition
timing, taken as start of HTR, is slightly delayed when fuel


4

G. S. JUNG, Y. H. SUNG, B. C. CHOI and M. T. LIM

Figure 4. Maximum rate of pressure rise and knock as a
function of crank angle for DME-air mixtures of different
equivalence ratios, formed by premixing or direct injection.

Figure 6. Released heat and combustion efficiency as a
function of equivalence ratio and injection timing.

Figure 7. Combustion temperature as a function of equivalence ratio and injection timing.
Figure 5. Combustion timing as function of fuel injection
timing and equivalence ratio.
injection is substantially delayed. The delay is bigger as the
equivalence ratio increases, possibly due to the cooling
effect of the fuel vaporization and also due to the higher
degree of mixture stratification. Mixtures of greater equivalence ratio also have shorter combustion durations (especially for HTR).
Figure 6 shows net heat release and combustion efficiency in tests of mixtures with various equivalence ratios,
where fuel is either premixed or directly injected at different crank angles. Hatched columns in the figure indicate
total amount of released heat, while green curves represent
the corresponding combustion efficiency. The combustion
efficiency is rather low, between 40 and 70 percent, with
higher values associated with greater equivalence ratio.
Further observation reveals that the combustion efficiency
does not depend on the fuel injection timing at equivalence
ratio of 0.2 or 0.3, but it gradually decreases at equivalence
ratio of 0.1 as the injection is retarded from BDC toward

30° BTDC.
Figure 7 shows bulk combustion temperature of the incylinder mixtures during the combustion tests under various fueling strategies. The blue bottom portion, the red
upper portion, and the highest point of the bars, respectively, indicate the temperatures during LTR, during HTR,

and at the final stage of combustion.
The in-cylinder gas temperature before the ignition is
calculated from the equation (1) assuming an adiabatic
compression process because the fuel-air mixture in the
central part of the combustion chamber, where the ignition
is most likely to occur first, gets least influenced by the
heat loss to the walls (Iida et al., 2004; Sato et al., 2003).
P ( θi ) ⎞
T ( θ i )=T ( θ i – 1 ) ⋅ ⎛⎝ ------------------P ( θi – 1 ) ⎠

γ–1
----------γ

(1)

Since the adiabatic assumption included in the isentropic
process is not valid once ignition has occurred, the gas
temperature is calculated using the equation (2), the ideal
gas equation of state.


EFFECTS OF MIXTURE STRATIFICATION ON HCCI COMBUSTION OF DME IN A RAPID COMPRESSION

Figure 8. Imep as a function of fuel injection timing and
equivalence ratio.
P ( θi ) ⋅ V ( θ i ) ⋅ n ( θi – 1 ) ⋅ T ( θi – 1 )
T ( θ i )= ----------------------------------------------------------------------------P ( θ i – 1 ) ⋅ V ( θi – 1 ) ⋅ n ( θi )

(2)

T [K], P [Pa], V [m3], n [mol], γ, and θ [deg.] in the above
equations represent the temperature, pressure, volume,
specific heat ratio of the gas in the cylinder, and the engine
crank angle, respectively. A specific heat ratio of 1.3 was
obtained from the logarithmic plot of the measured pressure and volume pertaining to the engine operating conditions.
It is clear from the figure that LTR begins at about the
same temperature of 660 K ± 10 K regardless of either the
equivalence ratio or the fuel injection timing.
Figure 8 shows indicated mean effective pressure (imep)
obtained under various fueling conditions. As the fuel injection is retarded from BDC to 30° BTDC, imep steadily
increases a little (up to about 0.05 MPa) except at the leanest
condition. The general trend of increasing imep with delayed ignition, caused by retarded fuel injection, is attributed
to the decreasing compression work due to the delay in
pressure rise during the compression stroke.

Figure 9. Concentration of unburned hydrocarbons in
combustion products of DME-air mixtures formed by
premixing or direct injection.

5

3.2. Pollutants Formed in HCCI Combustion
HCCI engines produce significantly more hydrocarbons than
conventional diesel engines, contrary to the early expectations. Figure 9 shows concentrations of total hydrocarbons
and the three major species measured in combustion products of the tests under various fueling strategies. As can be
seen in the figure, hydrocarbons consist mostly of the
unburned DME fuel, formaldehyde and methane. Among
the three fueling scenarios for each of the same equivalence
ratio, the case of premixed fuel has the highest concentration of THC in Figure 9 followed by the case of earlier
fuel injection. The trend can be easily explained since fuel

injected in the center of the combustion chamber at an
earlier timing will have more chance of getting trapped in
the top ring crevice, which is considered to significantly
contribute to DME emission in HCCI operation. Among
the cases where fuel is supplied in the same manner, THC
seems to decrease as the equivalence ratio is raised from
0.1 to 0.2, and then slightly increases (premix fueling), or
stays about at the same level (fuel injected at BDC) for
further increases of equivalence ratio from 0.2 to 0.3. The
major decrease of THC in the first interval of equivalence
ratio results from a similar drop in DME. Since equivalence ratio 0.1 is close to the lean limit of auto ignition,
which will tend to help more DME near the walls or in the
crevices avoid combustion, the highest concentration of
DME and the low combustion efficiency is measured at
equivalence 0.1. When the equivalence ratio is increased to
0.2 or above, more DME will burn to reduce THC in the
exhaust gas, while more DME becomes trapped in the
crevices to increase THC. The latter effect will play a more
important role when the cylinder charge is more homogeneous, as explained above, and also become more pronounced at conditions of greater equivalence ratio. This may
be the factor to explain the significant, in the case of premixed fuel, or marginal, in the case of fuel injected at BDC,
increase of DME and also THC at the equivalence ratio of
0.3.
Summarizing the above discussions, one may cautiously
model HC formation in these cases so that the prevailing
mode in the very lean range of the equivalence ratio below

Figure 10. NOx and CO concentrations in combustion
products of DME-air mixtures formed by premixing or
direct injection.

6

G. S. JUNG, Y. H. SUNG, B. C. CHOI and M. T. LIM

0.1 is bulk quenching of the extra-lean mixture due to
incomplete ignition or cold walls in the neighborhood, while
the crevice volume effect is the dominant one in the richer
range.
Figure 10 shows NOx and CO concentrations in the
combustion products of mixtures formed through either
premixing or direct injection. Smaller amounts of CO are
generated at higher equivalence ratios regardless of the
fueling strategy, most likely due to higher combustion temperature associated with the greater equivalence ratios. Premixed mixtures look to generate a little more CO at every
equivalence ratio than do mixtures prepared by direct injection. This is considered quite natural because more fuel will
exist in the close vicinity of the cold walls if fuel is premixed before induction than if it is injected later near the
center of the cylinder. NOx concentration appears to be
roughly proportional to equivalence ratio, which is logical
because of the expected variation of the combustion
temperature. Nevertheless the absolute levels of NOx concentration are fairly low in comparison with those in the
conventional diesel combustion.

4. CONCLUSIONS
Characteristics of DME HCCI combustion and composition of the combustion product were experimentally investigated, while various quantities of DME fuel were introduced by different methods in the combustion chamber of an
RCEM. Major findings of the study are as follows.
(1) The knock-limited rich limit of HCCI operation is expanded by directly injecting fuel.
(2) A homogeneous DME-air mixture is associated with
shorter combustion duration and higher peak pressure
than a mixture formed by direct fuel injection.
(3) The timing of LTR is slightly delayed when fuel injection timing is retarded.

(4) Imep increases with delayed timing of fuel injection.
(5) The main components of THC are unburned fuel (DME),
formaldehyde and methane, and more fuel remains
unburned when fuel is supplied premixed.
(6) More THC and CO are formed for any equivalence
ratio tested when fuel is introduced premixed than when
it is directly injected at BDC.
(7) As fuel injection is retarded, ignition occurs later with
less THC and more CO formed.
(8) As equivalence ratio is raised, THC emission decreases
in the low range of equivalence ratio (0.1 to 0.2), but it
increases (fuel premixed) or levels off (fuel injected) in
the higher range of equivalence ratio (0.2 to 0.3). More
nitrogen oxides are generated as the equivalence ratio
increases, but their absolute levels are quite low.
ACKNOWLEDGEMENT−This research was conducted as a
part of “Development of Basic and Practical Technology for
HCCI Engine” under the financial sponsorship of the Ministry of
Knowledge Economy.

REFERENCES
Cho, S. H., Kim, K. S. and Lim, M. T. (2004). Development
of a rapid compression expansion machine and compression ignition combustion of homogeneous premixtures.
Korean Society of Automotive Engineers 12, 2, 83−90.
Chung, J. W., Kang, J. H., Kim, N. H., Kang, W. and Kim,
B. S. (2008). Effects of the fuel injection ratio on the
emission and compression performances of the partially
premixed charge compression ignition combustion engine
applied with the split injection method. Int. J. Automotive
Technology 9, 1, 1−8.

Dec, J. E., Hwang, W. and Sjoberg, M. (2006). An investigation of thermal stratification in HCCI engines using
chemiluminescence imaging. SAE Paper No. 2006-011518.
Gray, A. W. and Ryan, T. W. (1997). Homogeneous charge
compression ignition (HCCI) of diesel fuel. SAE Paper
No. 971676.
Hamada, K., Niijima, S., Yoshida, K., Shoji, H., Shimada,
K. and Shibano, K. (2005). The effects of the compression ratio, equivalence ratio, and intake air temperature
on ignition timing in an HCCI engine using DME fuel.
SAE Paper No. 2005-32-0002.
Iida, N., Yamasaki, Y., Sato, S., Kumano, K., Kojima, Y.
(2004). Study on auto-ignition and combustion mechanism
of HCCI engine. SAE Paper No. 2004-32-0095.
Inagaki, K., Fuyuto, T., Nishikawa, K., Nakakita, K. and
Sakata, I. (2006). Dual-fuel PCI combustion controlled
by in-cylinder stratification of ignitability. SAE Paper
No. 2006-01-0028.
Kumano, K. and Iida, N. (2004). Analysis of the effect of
charge inhomogeneity on HCCI combustion by chemiluminescence measurement. SAE Paper No. 2004-011902.
Lim, O. T., Nakano, H. and Iida, N. (2006). The research
about the effects of thermal stratification on n-heptane/
iso-octane-air mixture HCCI combustion using a rapid
compression machine. SAE Paper No. 2006-01-3319.
Longbao, Z., Hewu, W., Deming, J. and Zuohua, H. (1999).
Study of performance and combustion characteristics of
a DME-fueled light-duty direct-injection diesel engine.
SAE Paper No. 1999-02-3669.
Ogawa, H., Miyamoto, N. and Yagi, M. (2003). Chemicalkinetic analysis on PAH formation mechanisms of oxygenated fuels. SAE Paper No. 2003-01-3190.
Sato, S. and Iida, N. (2003). Analysis of DME homogeneous charge compression ignition combustion. JSAE Paper
No. 20030236.
Sjöberg, M. and Dec, J. E. (2006). Smoothing HCCI heatrelease rates using partial fuel stratification with twostage ignition fuels. SAE Paper No. 2006-01-0629.

Sjöberg, M., Dec, J. E. and Cernansky, N. P. (2005). Potential of thermal stratification and combustion retard for
reducing pressure-rise rates in HCCI engines, Based on
multi-zone modeling and experiments. SAE Paper No.


EFFECTS OF MIXTURE STRATIFICATION ON HCCI COMBUSTION OF DME IN A RAPID COMPRESSION

2005-01-0113.
Teng, H., McCandless, J. C. and Schneyer, J. B. (2004).
Thermodynamic properties of dimethyl ether - An alternative fuel for compression-ignition engines. SAE Paper
No. 2004-01-0093.

7

Thring, R. H. (1989). Homogeneous-charge compressionignition engine. SAE Paper No. 892068.
Yao, M., Zheng, Z., Xu, S. and Fu, M. (2003). Experimental
study on the combustion process of dimethyle ether
(DME). SAE Paper No. 2003-01-3194.


International Journal of Automotive Technology, Vol. 10, No. 1, pp. 9−16 (2009)
DOI 10.1007/s12239−009−0002−8

Copyright © 2009 KSAE
1229−9138/2009/044−02

EXPERIMENTAL INVESTIGATION OF CHARACTERISTICS OF
PRESSURE MODULATION IN A FUEL INJECTION SYSTEM
D. HUANG1,2)* and M.-C. LAI1)
1)

Mechanical Engineering Department, Wayne State University, Detroit, Michigan 48202, USA
2)
Mechanical Engineering Department, Shanghai University, Shanghai 200072, China
(Received 25 July 2007; Revised 24 April 2008)

ABSTRACT−A piezoelectric atomization device achieves fuel pressure modulation through vibration of a piezoelectric
pressure modulator. As a consequence, the fast alternating and slow moving streams collide with each other and further break
up the fuel drop. In this paper, an experimental investigation was carried out to study the fluid dynamic characteristics of the
spray atomization process of automotive port fuel injectors with a piezoelectric pressure modulator. The investigation mainly
focuses on: (a) the coupling characteristics between the piezoelectric stack and the hydraulic as well as the transfer
characteristics of pressure modulation from the piezoelectric modulator to the point above the orifice; (b) the time history of
the pressure dynamic response at the point above the orifice under a typical modulation frequency, which reflects the variation
of pressure modulation while the fuel injector is working; and (c) the time-variation characteristics related to mechanical
structure and fluid dynamics. The experimental results expose some important dynamic characteristics of pressure modulation,
which will be very significant and lead us to greatly improve the fuel injection system, optimize the control parameters and
implement spray atomization with a high quality performance in the near future.
KEYWORDS : Pressure modulation, Fuel injection system, Dynamic characteristic, Spray atomization

1. INTRODUCTION

330 Hz) produced by a pressure modulation machine.
Although different models (Ren and Mally, 1996) have
been developed to describe the fuel injection system, few
can accurately predict the transient response of a fuel injection system, especially its internal pressure modulation. In
order to fill the defect of the numerical computation, the
present paper is experimentally investigating the fluid
dynamic characteristics in a fuel injection system, which
contains a piezoelectric modulator. This study was motivated by the success of earlier work on fuel spray atomization
in the automotive port injector. After carefully observing

the coupling phenomenon of the piezoelectric-hydraulic
system and the dynamic response of the fuel injection
system with pressure modulation, the characteristics of pressure modulation in a fuel injection system are summarized. Along with this summary, this manuscript also describes the experimental setup, procedures and signal processing employed in the experiments. Experimental results
and future work are then discussed.

Spray atomization is very important to reduce the fuel/air
mixing time and minimize the attachment of liquid fuel to
port surfaces in automotive and turbo fan engines. Wellatomized gasoline spray has a high potential to reduce
hydrocarbon emissions and improve the engine cold starting properties. Dressler developed a pressure modulator
based on the piezoelectric principle (Dressler, 1993). The
pressure modulator is installed inside the fuel line and
generates the pressure modulation. It efficiently enhances
the atomization characteristics of gasoline spray injectors
(Zhao et al., 1996; 1995; 2002; Sipperley et al., 1998;
Schiller et al., 2006; Kim at al., 2004).
To optimize the performance of the pressure modulation
fuel system in spray atomization (Hu and Wu, 2001a;
2001b), presented mathematical modeling of an individual
injector and an entire fuel injector system by considering
one-dimensional, unsteady Bernoulli’s equation and loss
factors of kinetic energy Kf and Ko based on a discrete
segment of the injector (loss factors Kf and Ko are used to
account for the losses of kinetic energy as fluid enters the
injector through the filter at the top and discharges through
the orifice at the bottom, respectively) (Miller, 1990). They
predicted the dynamic response of an automotive fuel
system under low frequency pressure fluctuation (around

2. EXPERIMENTAL SETUP
Figure 1 shows the scheme of an experimental pressure

modulated port fuel injector system. The pressure modulator, as shown in Figure 1(a), is composed of a piezoelectric driver mounted inside a circular housing and bolted
to one end. The other end of the circular housing is closed
by a rigid endplate upon which the fuel injectors are

*Corresponding author. e-mail:
9


10

D. HUANG and M.-C. LAI

Figure 1. Schematic of experimental setup.
mounted by means of an adapter. The piezoelectric driver
consists of a basemount, a pair of piezoelectric disks, and a
piston. A hollow bolt clamps these items together into a
cylindrical assembly. The liquid from the fuel tank, pressurized by high pressure nitrogen gas, is introduced into
the piezoelectric driver through the basemount. It flows
through the hollow bolt and a passage in the piston and
then enters the fluid manifold surrounding the piston. The
fluid flows through a hole in the end plate and the injector
adapter, and then enters the gasoline port injector.
The piezoelectric elements receive the electrical signal
and convert it into a longitudinal motion. Such a process
produces a transient variation in the pressure of the fluid as
it passes through this gap, allowing for pressure modulation of the fuel system. Therefore, the device here is
referred to as a pressure modulator. It should be denoted
that even though there is significant pressure perturbation
inside the fuel line, the liquid flow rate remains nearly
constant. While the device can operate at many frequencies, the optimized operating point can be determined in

order to minimize energy loss and maximize pressure perturbation. The device has a small size and can be installed
easily between the port fuel injector and the fuel line. There
is no extra control difficulty associated with this system
since the fuel injection rate and injection timing are

controlled by the conventional fuel injector metering valve.
It should be noted that while ultrasonic atomizers are only
advantageous for atomization at low flow, the pressure
modulator technique can provide good atomization over a
wide range of fuel flow.
The working fluid used in the experiment is surrogate
fuel (Viscor 16B), which has the same viscosity and density as gasoline, but is more safe to use in tests of fluid
dynamics and atomization.
As shown in Figure 1(b), the system consists of a center
port injection (CPI) injector developed by (Zizelman et al.,
1992). The CPI injector has a single spray hole with a
diameter of 500 μm, and is connected to the pressure
modulator through an adaptor. The baseline case is the
spray with static pressure of 276 kPa (40 psi) regulated by
the pressure of the nitrogen gas. The pressure modulator
receives a sinusoidal signal from a wave generator operating between 3 and 30 kHz, and the signal is amplified by an
amplifier before it is used as a modulation voltage driving.
The final signal delivered to the crystal stacks is a 25~250
V sine wave. When the CPI injector is controlled by a pulse
wave, it alternatively opens and closes and the fuel passes
through the orifice as instructed. The pulse width is set at
16 ms for the pressure modulation test.
To measure the fluid dynamic pressure in the piezoelectric pressure modulator and injector, three KulteXT123C-100 pressure sensors are installed. The first one is
located at the point above the injector orifice and its
responding pressure is P1, which directly affects spray

atomization. The second one is near the entrance of the
injector chamber and its responding pressure is P2. The
third one is located in the middle of the pressure modulator
and its responding pressure is P3, which directly reflects
the source of pressure modulation. All sensors are connected to computer data acquisition.
A computer collects data from the measuring points and
processes the data for different statuses. With this setup,
one can simulate various working conditions with different
modulation frequencies, extract the dynamic characteristics
of pressure modulation in the fuel injection system and
observe the fuel spray atomization.

3. COUPLING CHARACTERISTICS
The measurement of the coupling characteristic between
the piezoelectric stack and the hydraulic system is carried
out with the port injector being closed. The single frequency sine wave is applied to the piezoelectric stack, and as
a result, the pressure modulation is generated in the fuel
injection system. By recording the pressure P1 at a point
above the orifice of the injector and its modulation frequency, and using the signal filtering process to extract the
pressure modulation component, the responding amplitude
is given for different driving voltages. Repeating this procedure for a series of modulation frequencies, the measurement result for the coupling characteristics is obtained, as


EXPERIMENTAL INVESTIGATION OF CHARACTERISTICS OF PRESSURE MODULATION

Figure 2. Measurement result of coupling characteristic
between the piezoelectric stack and the hydraulic, where
static fuel pressure is 276 kPa (40 psi) on the CPI injector
and the sine wave driving voltage is 25~250 V.
shown in Figure 2, where the driving voltage varies from

25 to 250 V and the frequency range is from 1 to 17 kHz.
The experimental data analysis shows a strong nonlinear
phenomenon in the piezoelectric stack and hydraulic coupling, which depends on the driving voltage over the piezoelectric stack and modulation frequency. At several frequency points, the magnitude of pressure modulation is not
sensitive to driving voltage. When the driving voltage is
greater than 50 V, the increment of pressure modulation
almost halts. In this application, using this nonlinear
characteristic, the biggest magnitude of pressure modulation can be attained at some special frequency points with
smaller driving voltages. For example, around frequencies
of 8 kHz, 9 kHz and11.5 kHz, satisfying satisfactory level
of pressure modulation can be obtained through a driving
voltage of 50 V. The optimized control parameters can save
the driving power in the fuel injection system. Although
different characteristics will occur in different structures of
the piezoelectric stack, through experimental investigation,
we can determine the optimal driving voltage for pressure
modulation.
In fact, saturation status of the coupling characteristic of
pressure modulation is almost attained for the given piezoelectric stack while the driving voltage is close to 250 V in
the range of working frequencies from 6 kHz to 17 kHz.
There is no benefit to using a higher voltage as a driving
power. Therefore, a voltage of 250 V is employed in examining the coupling characteristic of pressure modulation
in the fuel injection system study to obtain the best output
property over the given frequency range.
The coupling characteristics of pressure modulation are
very sensitive to minor changes in the mechanical structure
of the fuel injection system. While the port injector is open,
the coupling characteristics vary with flowing fuel. A detailed representation of the change in coupling characteristics
during the injector opening is presented in Figure 3, where

11

Figure 3. Comparison of coupling characteristics with
closed and opened CPI injector, where static fluid pressure
is 276 kPa (40 psi) and the driving voltage is 250 V.
the driving voltage is 250 V. Under the given voltage
condition, the coupling characteristics at P1 are greatly
altered below a frequency of 7 kHz while the port injector
is opened, and keeps the same in the higher frequency
range. Notably, the curve of the coupling characteristic tends
to be smooth over the whole working frequency range in
this situation. However, the maximum output magnitude of
pressure modulation is worth considering, and it can be
determined from the given coupling characteristics. At a
frequency of 5.4 kHz, the maximum output value is about
22 kPa (rms value). For the same reason, a similar phenomenon can be observed at measuring point P3.
The coupling characteristics of pressure modulation are
related to the static pressure in the fuel injection system.
The magnitude of pressure modulation at P1 is proportional
to the static pressure when the port injector is closed, and

Figure 4. Relationship between the magnitude of pressure
modulation and static pressure, where the driving voltage is
150 V over the piezoelectric stack while the CPI injector is
closed.


12

D. HUANG and M.-C. LAI

Figure 5. Transfer characteristics of pressure modulation
while the CPI Injector is (a) closed and (b) opened, where
the static fluid pressure is 276 kPa (40 psi) and the driving
voltage is 250 V.
this is confirmed by the detailed results shown in Figure 4.
It also can be seen that different modulation frequencies
have different magnitudes under a constant driving voltage
of 150 V.

4. TRANSFER CHARACTERISTICS
The term transfer characteristics, in particular, indicate the
transmission properties of pressure modulation from the
source of the pressure modulation to the point above the
injector orifice. It also expresses the pressure modulation
difference between P1 and P3. The transfer characteristics
mainly depend on the route through which the fuel flows,
working status (whether the port injector is closed or open),
and the quantity of fuel sprayed per second. While the
pressure signals P1 and P3 are sampled at the same time,
the transfer characteristics of the pressure modulation can
be obtained by signal processing, as shown in Figure 5,
where the driving voltage is 250 V. Figure 5(a) is the port
injector closed case, and Figure 5(b) is the port injector
opened case.
Though the route between the measuring points P1 and
P3 is very short, wherever the injector is closed or open,
due to the complicated inner space structure of the CPI
injector, the transfer characteristics of the pressure modulation cannot remain constant in the fuel fluid and take a
multiple harmonic form.

Figure 6. Time history record of pressure signal P1 while
the CPI injector is working, where the static fluid pressure
is 276 kPa (40 psi) and the driving voltage is 155 V.
injector is working, the pressure signal P1 is recorded.
Some typical records, such as the modulation frequencies
0, 2.2, 4.0, 8.0 kHz reflecting the variation of pressure
modulation, are plotted in Figure 6, where the transient
dynamic response and a 16 ms width pulse control signal
can be seen clearly. To make a comparison with another
result, the time history records at the measuring point P3
are shown in Figure 7.
No modulation case (Modulation Frequency is 0 Hz):
While the injector is open, the pressure drops immediately
because of rarefaction in the fluid. When the injector is
suddenly closed, compression occurs in the fluid system,
which causes a strong pressure surge known as the “water
hammer” effect on the fuel system. This dynamic response
oscillates, decaying exponentially to the ambient level due
to fluid viscosity. It should be mentioned that while the
injector is open, the pressure signals P1 and P3 also exhibit
a dynamic response, but the observing time is too short for
us to distinguish them clearly.
Modulation case: To estimate the variation in the

5. RESPONSE CHARACTERISTICS
Response characteristics mainly involve two aspects of the
fuel injection system and pressure modulation. One part is
the dynamic response that relies on the structure of the fuel
injection system, the physical property of fuel fluid and the
static pressure. The other is the variation of pressure

modulation, which we focus on in this paper. As the port

Figure 7. Time history record of pressure signal P3 while
the CPI injector is working, where the static fluid pressure
is 276 kPa (40 psi) and the driving voltage is 155 V.


EXPERIMENTAL INVESTIGATION OF CHARACTERISTICS OF PRESSURE MODULATION

Figure 8. Variation of pressure modulation at P1 while
the CPI injector is working (Unit: kPa), where the static
fluid pressure is 276 kPa (40 psi) and the driving voltage is
155 V.
pressure modulation, the pressure signal is processed using
a band pass filter. The harmonic component of modulation
is plotted in Figure 8.
As the modulation frequency is maintained at 2.2 kHz,
the magnitude of pressure modulation becomes smaller
when the injector is open versus when the injector is
closed. In fact, its amplitude is reduced to about 65%. The
coupling characteristic occurs in a soft manner for the
output property. While the port injector is open, the static
pressure mean decreases to 150 kPa, and the ratio of the
peak to peak values of the pressure modulation to the mean
static pressure Pm/so is around 10%. As such, in the case
when pressure frequency is 4 kHz, the magnitude of
pressure modulation becomes bigger when the injector is
opened versus closed. The amplitude is 6.5 times as much
as it is the closed case. The coupling characteristic occurs
in a harder manner for the output property. In this situation,

the ratio Pm/so is around 23%. However, whether the injector
is opened or closed, the modulation magnitude is the same
in the modulation frequency range from 7 to 13 kHz for the
output property. For example, at a frequency of 8.0 kHz,
the modulation magnitude is 17.5 kPa (rms value), and the
ratio Pm/so is around 30%. Its output property remains constant. However, at the moment the injector opens or closes,
there is an instantaneous loss in the amplitude of pressure
modulation, and tow gaps are formed in the response.

6. SPECTRAL ANALYSIS
FFT is used to analyse the harmonic components in the
pressure signal. As shown in Figure 9, the spectral analysis
of pressure signal P1 is carried out. The time history
records come from the case of 0.0 kHz and 2.2 kHz in
Figure 6, and the data analysis commences at the moment
the injector closes. Without modulation, the dominant
component is 45 Hz and is the natural frequency compo-

13

Figure 9. Comparison of spectral analysis on pressure
signal P1 while CPI injector is closed where the static fluid
pressure is 276 kPa (40 psi).
nent reflecting the response of the inner structure. In the
modulation case, the modulation frequency can be found as
an additional component of 2 kHz in the spectral picture.
Through spectral analysis, we know that there are two main
harmonic components in the fuel injection system when the
port injector is working. However, for the components with
a very short time response, especially instantaneous characteristics, there is no solution shown in the spectral analysis.

STFT (Short Time Fourier Transform), as shown in
Figure 10, is employed to observe instantaneous magnitude
and frequency while the injector is working. To see it clearly, its harmonic components with low frequency (natural
frequency in the fuel injection system) are removed from
the 3D picture. With the help of the 3D picture, we can
intuitively learn that the pressure modulation varies in
detail with time as well as the soft and hard output properties. At the moment the injector is opened or closed, the
magnitude of pressure modulation has instantaneous loss.
The phenomenon of instantaneous loss can easily be
observed from Figure 10(c), where there are two gaps in
the component of pressure modulation. Besides this, some
instantaneous harmonic components, both the frequency
and the amplitude varying with time significantly, such as 3
kHz and 8.5 kHz, are also revealed in the 3D picture, as
shown in Figure 10(b), which are given out by a step
excitation while the injector is opened and closed, and is
probably related to the inside structure of the port injector.
Indeed, they are interesting components, and we still need
to study their dynamic reaction.
CWT (Continuous Wavelet Transform) is used to describe the power spectral density in the time-frequency
domain. It has an advantage in depicting time-variation
characteristics for non-stationary signal. Minor disturbing
of the modulation frequency during the transition can be
observed from the CWT spectrogram shown in Figure


14

D. HUANG and M.-C. LAI

Figure 11. Contour of CWT spectrogram for pressure
modulation while the CPI injector is working, where the
modulation frequency is 4.0 kHz, the static pressure is 276
kPa and the driving voltage is 155 V.

7. CONCLUSIONS

Figure 10. STFT analysis of pressure modulation P1 while
the CPI injector is working, where the static fluid pressure
is 276 kPa (40 psi) and the driving voltage is 155 V.
11(a) in the contour form, where the pressure modulation
with a frequency of 4 kHz at P3 is analyzed. At the moment
the injector opens, the modulation frequency instantly decreases. In the inverse case, it instantly increases. This
phenomenon can be explained by the non-stationary fuel
flowing inside the pressure modulator. At that moment, the
flowing fluid may tend to the vortex case, which has an
influence on the coupling characteristics between the
piezoelectric stack and the hydraulic.

(1) The results of experimental investigation show that the
coupling characteristics between the piezoelectric stack
and the hydraulic possess nonlinear features in a fuel
injection system. The coupling characteristics will
attain saturation status in the higher frequency range
with a driving voltage beyond 250 V. It will be changed
with an open injector. Moreover, the curve of coupling
characteristics tends to be smooth over whole working
frequency range. With the given piezoelectric modulator, the magnitude of pressure modulation will be
reduced within the frequency range of 2~3 kHz and
5~7 kHz while the port injector is open; but the

magnitude of pressure modulation will be enlarged
within the frequency range of 3~5 kHz; it will remain
the same within the frequency range of 7~13 kHz.
The transfer characteristics from the source of pressure
modulation to the nozzle are sensitive to the injection
condition. Different properties will appear while the
injector is opened or closed.
(2) The maximum output of pressure modulation can be
determined for the fuel spray atomization from the
coupling characteristics. Based on the coupling characteristics, we can select the optimal driving voltage to
obtain the satisfying pressure modulation. At special
frequencies, smaller driving voltages can result in
bigger output magnitudes of pressure modulation to
output, which will save driving power in engineering
applications.
(3) To analyze the instantaneous characteristics in the fuel


EXPERIMENTAL INVESTIGATION OF CHARACTERISTICS OF PRESSURE MODULATION

15

Figure 12. Pressure in measuring point P3 with a driving
frequency of 27.2 kHz, where the static pressure is 276
kPa, the driving voltage is 155 V and the injector is closed.
injection system, STFT and CWT are employed to
study the response at P1 and P3. These are good signal
analysing tools that are well-suited to observe timevariation characteristics in the pressure modulation, by
which not only instantaneous harmonic components of
3 kHz and 8.5 kHz at P1 are captured from STFT, but

also negative changes in gaps of modulation components are observed while the injector is opened or
closed. On the other hand, a hidden variation, such as
instantaneous variation in the modulation frequency at
P3, is also discovered from CWT. Due to complications in the mathematical model of the fuel injector,
so far these time-variation characteristics are difficult to be numerically simulated by the Computational
Fluid Dynamics method, and their fluid dynamic
properties are also difficult to fully explain. Therefore,
it is necessary for researchers to study fluid dynamics
in the fuel injector through experimentation and
modern signal processing.
(4) At modulation frequencies of 23.3 kHz or 27.2 kHz,
extremely strong, but unsteady pressure pulsation is
observed at the measuring point P3. An example is
shown in Figure 12, where the static pressure is 276
kPa and the sine wave driving voltage is 155 V. This
phenomenon may be explained as the resonance in the
mechanical structure of the piezoelectric pressure
modulator. The pressure modulation could not be
directly used in spray atomization.
(5) Experimental investigation should expand the frequency range of pressure modulation up to 50 kHz in
the future research, in which more dynamic characteristics could be captured in the fuel injection system.
(6) To clarify the pressure modulation effect on spray
characteristics, Figure 13 shows the comparison of
spray performance, including the histograms of droplet
size and velocity between the baseline and pressure
modulation CPI injector based on the PDA measurements. It is clear that the weak bimodal distribution
appearing in the baseline spray size histogram dis-

Figure 13. Comparison of spray performance for the CPI
injector at a fuel injection pressure of 276 kPa.

appears completely for the pressure modulation spray.
This observation is also confirmed in size-velocity
correlation. Almost no large droplets are observed from
the pressure modulation spray, and the droplet size is
greatly reduced in the pressure modulation mode. The
fuel spray is more uniform in space with the pressure
modulation technique compared to the baseline condition spray. Moreover, the percentage of high velocity
droplets shows the apparent increase in the velocity
histogram of the pressure modulation spray. The
dependence of the atomization characteristics on driving power and frequency of this pressure modulation
technique indicates its potential application in controlling spray performance of automotive port fuel injectors. More detailed test results regarding the atomization improvement with pressure modulation techniques have been presented in the literature (Kim et al.,
2004).

REFERENCES
Dressler, J. L. (1993). Liquid Droplet Generator. U.S.
Patent 5248087.
Hu, Q. and Wu, S. F. (2001). Modeling of dynamic ran
automotive response of fuel rail system, Part 2. Entire
system. J. Sound and Vibration, 245, 815−834.
Hu, Q. and Wu, S. F. (2001). Modeling of dynamic response
of an automotive fuel rail system, Part 1: Injector. J.
Sound and Vibration, 245, 801−814.
Kim, H., Im, K.-S. and Lai, M.-C. (2004). Pressure modulation on micro-machined port fuel injection performance.
Int. J. Automotive Technology 5, 1, 9−16.
Miller, D. S. (1990). Internal Flow System. 2nd edn..
BHRA Information Services. Cranfield. Bedford. UK.


16

D. HUANG and M.-C. LAI

Ren, W. and Nally, Jr. J. F. (1996). Computer modeling of
steady and transient flows within a gasoline fuel injector.
Proc. American Society of Mechanical Engineers Fluids
Engineering Division, 242, 141−147.
Schiller, N. H., Saunders, W. R., Chishty, W. A., Vandsburger,
U. and Baumann, W. T. (2006). Development of a piezoelectric-actuated fuel modulation system for active combustion control. J. Intelligent Material a System and
Structures, 17, 403−410.
Sipperley, C. M., Dwards, C. F. E., Wang, D. F. and Ganji,
A. R. (1998). Piezoelectrically driven simplex atomizers
at atmospheric pressure. ILASS-Americas, Sacramento,
CA 128−132.

Zhao, F.-Q., Harrington, D. L., Lai, M.-C. (2002). Automotive
Gasoline Direct-Injection Engines. Society of Automotive
Engineers. 317−337.
Zhao, F.-Q., Lai, M.-C. and Harrington, D. L. (1995). The
sprays characteristics of automotive port fuel injection-A
critical review. SAE Paper No. 950506.
Zhao, F.-Q., Lai, M.-C., Amer, A. A. and Dressler, J. L.
(1996). Atomization characteristics of pressure-modulation
automotive port injector sprays. Atomization and Sprays,
6, 461−483.
Zizelman, J., Seino, M. J. and Graves, M. C. (1992). Center
port fuel injection. SAE Paper No. 920295.


International Journal of Automotive Technology, Vol. 10, No. 1, pp. 17−25 (2009)
DOI 10.1007/s12239−009−0003−7

Copyright © 2009 KSAE
1229−9138/2009/044−03

COMBUSTION DEVELOPMENT OF A BI-FUEL ENGINE
O. S. ABIANEH1)*, M. MIRSALIM2) and F. OMMI1)
1)

Department of Mechanical Engineering, Tarbiat Modares University, Chamran High Way,
P.O. Box 14115-111, Tehran, Iran
2)
Department of Mechanical Engineering, AmirKabir University, 424 Hafez Ave, P.O. Box 15914, Tehran, Iran
(Received 8 August 2007; Revised 20 July 2008)
ABSTRACT−Environmental improvement and energy issues are increasingly becoming more important as worldwide
concerns. Natural gas is a good alternative fuel that can help to improve these issues because of its large quantity and clean
burning characteristics. This paper provides the experimental performance results of a Bi-Fuel engine that uses Compressed
Natural Gas as its Primary fuel and gasoline as its secondary fuel. This engine is a modification of the basic 1.4-liter gasoline
engine. Generally, on the unmodified base engine, torque and power for CNG fuel are considerably lower than gasoline fuel.
In this paper, the influence of fuels on wall temperature, performance and emissions are investigated.
KEY WORDS : Bi-fuel engine, CNG fuel, Gasoline fuel, Fuel consumption, Emission, Engine wall temperature

1. INTRODUCTION

per mass compared with gasoline fuel. However, for a
natural aspirated engine, the volumetric efficiency for a
gaseous fuel like CNG is more reduced than for a liquid
fuel (Figure 1). As a result, the output of the engine burning
gasoline is higher than if the CNG fuel was used.In order to
prevent abnormal wear in the engine, the hardness of the
valves and valve seats was increased. Additionally, in order

to reduce deformation, high-strength pistons and a modified water jacket design were employed in the cylinder
head.
In order to take advantage of the high octane rating, the
engine compression ratio was raised. However, for the
engine to work acceptably with gasoline the compression
ratio was then decreased to 10.8.
As shown in Table 2, the CNG base engine is a 4cylinder 1.4 liter-DOHC 16-valve engine.
Thermocouples are mounted 2 mm below the surface for
measuring surface temperature, as shown in Figure 2~3.

The IKCO Company has been developing an engine that
can run on both CNG and gasoline. Some companies have
not adopted CNG dedicated systems because these vehicles
have lower mileage and, thus, need more refueling. Environmental improvement and energy issues are becoming more
and more important as worldwide concerns. Natural gas is
a good alternative fuel to help improve these issues because
of its large quantity and clean burning characteristics. Since
the fuel system of CNG vehicles is completely closed, fuel
evaporative emissions are practically eliminated. The evaporative emission control system commonly used in gasoline
vehicles is unnecessary for CNG vehicles. In addition, inuse total reactive organic gas (ROG) emissions from fuel
storage and refueling of CNG vehicles are small (Gas
Research Institute, 1994). Therefore, CNG vehicles are highly beneficial for environmental protection relative to other
engines. In order to achieve a higher catalytic conversion
ratio, an A/F ratio sensor on the upstream of a closecoupled converter (to control to stoichiometric A/F ratio),
and a heated oxygen sensor on the downstream of an under
floor converter (to precisely compensate the A/F ratio)
were employed. The engine is also designed to pass the
Euro 5 emissions standard.A precisely controlled air/fuel
(A/F) ratio and a higher catalytic conversion using three
way catalysts are necessary to reduce exhaust gas emissions.

For precisely controlled A/F, a sequential multi-port gaseous
injection system (MPI) was chosen. Consequently, the
injectors and pressure regulator are newly developed.As
shown in Table 1, natural gas has a higher calorific value

2. IGNITION AND FLAME PROPAGATION
Methane has a lower unstretched and stretched laminar
burning velocity when the gas temperature is higher than
450 K (due to Markestein numbers) (Liss and Thrasher,
1991; Goodwin and Whiston, 1991; Duan, 1996; Bradley
et al., 1992). This proves that the total combustion duration is prolonged compared to diesel and gasoline fuel.
Turbulent burning velocity is approximately proportional
to uL0.6 (unstretched laminar burning velocity) and inversely
related to Le0.3 (Lewis number). As the compression
increases, the Lewis number decreases. Therefore, there is
less observed difference between the turbulent burning
velocities of natural gas and gasoline in the final stages of

*Corresponding author. e-mail:
17


18

O. S. ABIANEH, M. MIRSALIM and F. OMMI

Table 1. CNG and Gasoline specifications.
CNG

Gasoline

51.25

45.55

Density (kg/m ) at 25 C, 100 Kpa

0.739

749.1

Volumetric efficiency (Relative), (mixture base)

0.92

1

Poor power

Octane number (MON)

110*

94.7

Hard to knock

C4~C13

Lower pollution

lower CO2
lower reactivity

Calorific value per mass (Mj/kg) HHV
3

o

CH4: 88.9%
C2H6: 4.8%
C3H8: 1.21%
C4 and higher HC: 5.9%

Properties

Feature of CNG
Shorter mileage

Figure 1. Volumetric efficiency.
Figure 2. Cylinder block sensor position.
Table 2. Engine specifications.
Gasoline
Bore Stroke

7872

Displacement (cm)

1376

Compression ratio

10.8

IVO/IVC@1m lift

7°ATDC/197°ATDC
intake

Combustion chamber

CNG

←

pent roof

Max torque (N.m)

132

117

Max power (KW)

76

69

combustion. (Bradely et al., 1992; Jones and Evans, 1985).

However, some experiments suggest (Sharma et al., 1981)
that when one to five percent of methane is replaced with
ethane, there is a slight increase in burning velocity. In our
study, 4.8% of the test fuel is ethane (Table 1).
A comparison of the heat release for both fuels for
research engine is shown in Figure 4~7.
The heat release formula that is used in this paper is
displayed below: (Equation 1)
k
Q i = ------------ [ κ ⋅ p i ⋅ ( ν i + 1 – ν i – 1 ) + v i ⋅ ( p i + n – p i – n ) ]
κ–1
n

κ

: interval (1 deg of crank angle)
: polytropic coefficient

Figure 3. Cylinder head sensor position.

Figure 4. Crank angle difference between position of
maximum pressure and position of 5% Heat release (α max−
pressure − α 5%), full load, lambda=0.96 for CNG and 0.9 for
gasoline.
p

ν
k

: cylinder pressure

: volume
: constant


COMBUSTION DEVELOPMENT OF A BI-FUEL ENGINE

19

Figure 5. Crank angle difference between position of 5%
Heat release and ignition position (α5%−αignition), full load,
lambda=0.96 for CNG and 0.9 for gasoline.

Figure 8. Cooling water temperature difference (Toutlet,water−
Tinlet,water), at full load.

Figure 6. Crank angle difference between position of 50%
and 5% Heat release (α5%−α5%), full load, lambda=0.96 for
CNG and 0.9 for gasoline.

Figure 9. Exhaust gas temperature and temperature difference between CNG and gasoline exhaust gasses (TGasoline
exhaust gas−TCNG exhaust gas), 2000 rpm, lambda=1, 90ºC coolant.

Figure 7. Crank angle difference between position of 90%
and 5 % Heat release (α90%−α5%), full load, lambda=0.96
for CNG and 0.9 for gasoline.
As shown in Figure 5, the duration between 5% heat
release and ignition timing of the CNG is longer than the
duration for gasoline fuel in total speed range.
The duration between 50% heat release and 5% heat
release is nearly the same for both fuels, as shown in Figure

6. However, as shown in Figure 7 for the CNG fuel, the
duration between 90% and 5% heat release is more prolonged than for gasoline by approximately 10 to 20° of CA.
These results prove that the total combustion duration of
CNG fuel is prolonged compared with gasoline fuel.
Methane has a higher isentropic coefficient (Heywood,
1998) than other fuel, which results in a higher final
compression temperature. Additionally, the more advanced

timing for the increase of power output and efficiency
gives more combustion time and, hence, a higher temperature and pressure. The heat transfer to the walls is, however, not greatly changed for these two fuels, as shown in
Figure 8. Due to the higher isentropic coefficient of CNG,
the temperature of the gas in the expansion stroke (working
stroke) for CNG fuel decreases more rapidly than for gasoline fuel. Finally, the exhaust gas temperature of gasoline
fuel is higher than CNG exhaust gas fuel, as shown in
Figure 9.
The CNG’s lack of latent heat of evaporation increases
the temperature of the chamber wall during the induction
stroke.
If the total mass of fuel (gasoline) evaporates, it can cool
the air by up to 20ºC (Equation 2).
m· air C p ΔT=m· fuel L
L ≈ 300 kj/kg
C p – air ≈ 1.0049 kj/kg
m· air /m· fuel =14.39

(2)

The value of L (Latent heat of evaporation) is calculated
from the Clapeyron formula (Guibet, 1999).
The temperature difference between CNG and gasoline

fuel at Top Dead Center and at 1/4 stroke for the intake side
of the cylinder block wall is shown in Figure 10 and 11.
These temperatures are the same for both fuels.


20

O. S. ABIANEH, M. MIRSALIM and F. OMMI

Figure 10. Temperature and temperature difference of
cylinder block material (TCNG cylinder block−TGasoline cylinder block),
cylinder 3 at TDC, intake side, 2000 rpm, lambda=1, 90ºC
coolant.

Figure 13. Temperature and temperature difference of
cylinder block material (TCNG cylinder block−TGasoline cylinder block),
cylinder 3 at 1/4 stroke, exhaust side, 2000 rpm, lambda=1,
90ºC coolant.

Figure 11. Temperature and temperature difference of
cylinder block material (TCNG cylinder block−TGasoline cylinder block),
cylinder 3 at 1/4 stroke, intake side, 2000 rpm, lambda=1,
90ºC coolant.

Figure 14. Temperature and temperature difference of
cylinder head (TCNG cylinder block−TGasoline cylinder block), cylinder 3,
exhaust side, 2000 rpm, lambda=1, 90ºC coolant.

Figure 12. Temperature and temperature difference of
cylinder block material (TCNG cylinder block−TGasoline cylinder block),

cylinder 3 at TDC, exhaust side, 2000 rpm, lambda=1,
90ºC coolant.
Some types of matter affect the temperature of the
cylinder block, including radiation, latent heat of evaporation of gasoline and heat transfer from gases, which was
explained previously.
As shown in Figures 12 and 13, the temperature difference between CNG and gasoline fuels for the exhaust side
of cylinder block wall at TDC and 1/4 stroke is lower than

Figure 15. Temperature of cylinder head material and
BMEP, on cylinder 2, between exhaust valves, at Full load,
90ºC coolants.
intake side.
The temperature of the cylinder block at 1/4 stroke on
the exhaust side is lower than on the intake side because of
the oil return line.
as can be seen in Figures 11 to 16, the difference in the
temperature of the walls between the fuels in cylinder head
and cylinder block are the same.
Figure 17 shows the influence of lambda variation on
cylinder head temperature for CNG fuel. The maximum


COMBUSTION DEVELOPMENT OF A BI-FUEL ENGINE

Figure 16. Temperature of cylinder head material, on
cylinder 2, between exhaust valves, at Full load, 90ºC
coolants.

21

Figure 19. Pollution, 2000 rpm, lambda=1, BMEP=9 bar,
for different coolant temperatures in CNG mode.
that when the temperature of the cooling water is increased,
the amount of hydrocarbon pollutant is decreased. This is
because of the thickness of the quench area and the heat
transfers to the walls are decreasing but the amount of NOx
is increasing because of the higher combustion chamber
temperature resulting from lower heat transfer.

3. IGNITION AND FLAME PROPAGATION

Figure 17. Temperature of cylinder head and cylinder block
material, between exhaust valves, on cylinder 2 and 3,
2000 rpm, BMEP=9 bar, 90ºC coolant, for different
lambda in CNG mode.
temperature for CNG fuel at full load occurred at lambda
=0.94~0.96. At this point, flame speed and engine torque
are at their maximum.
The influence of water temperature on wall temperature
and nitrogen’s hydrocarbons and oxides is shown in Figures
18 and 19.
As shown in Figure 18, when the temperature of the
cooling water is decreased by 10ºC, then the temperature of
the cylinder head is also reduced by 10ºC. This relationship
is nearly linear also for other tested points. Figure 19 shows

The minimum spark energy required for methane ignition
is markedly higher than for other hydrocarbons. As a result,
the conversion of an engine to natural gas requires a highperformance ignition system (Guibet, 1999). In the fuel
tested in this study, other hydrocarbons (e.g. ethane) are

presented in Table 1. In the case of ethane, the coil can
ignite the CNG at 38 (mJ) energy. It was also observed that
the required energy for pure methane is 100 to 120 mJ
(Guibet, 1999).
Methane also has a wider flammability range than other
hydrocarbons. This allows an engine to operate on a lean
mixture, which is advantageous in some applications, such
as industrial vehicles.
Methane combustion is relatively slow. This can cause
deterioration in performance due to an increase in heat
transfer to the combustion chamber walls.
An interesting observation is that methane's slower
combustion contributes to less combustion noise due to a
less aggressive pressure gradient.

4. FULL LOAD INVESTIGATION

Figure 18. Temperature of cylinder head and cylinder block
material, cylinder 3, 2000 rpm, lambda=1, BMEP=9 bar,
for different coolant temperatures in CNG mode.

4.1. Torque and Power
The higher compression ratio increases torque and power
in CNG mode. In addition, the closed timing intake valve
was advanced in the CNG mode with CVVT2. As a result,
torque at lower engine speeds was nearly restored to levels
found during gasoline use.
Torque and Power for CNG decrease compared to
gasoline (Figure 20 and 21). In the worst case, power at
Continues Variable Valve Timing

2


22

O. S. ABIANEH, M. MIRSALIM and F. OMMI

Figure 20. Corrected torque.

Figure 21. Corrected power.
6000 rpm for CNG decreases by 6% compared to gasoline.
Torque was found to decrease 22% at 3500 rpm.
Torque, Volumetric efficiency and Power for CNG decrease compared to gasoline because of some reasons:
− The volumetric efficiency is reduced because of the
gaseous state of the fuel. Gaseous fuel occupies a
larger volume per unit energy than liquid fuel.
− The gasoline fuel mixture temperature is reduced
because of liquid fuel latent heat of evaporation and
therefore the volumetric efficiency is increased compare to gas fuel.
− The other main factor that reduces the power output is
the low flame speed of natural gas, which requires
more advanced spark timing to achieve complete combustion within the correct portion of the engine cycle
(Figure 7). This can cause a further reduction in the
power output of the engine.
4.2. BSFC
As shown in Figures 22 and 23, the BSFC (Brake Specific
Fuel Consumption) of CNG is lower than the BSFC of
gasoline. In addition, the normalized BSFC of CNG is
lower than the normalized BSFC of gasoline, but the difference is less than BSFC.

The formula for the normalized BSFC is: (Equation 3)
HHV CNG Norm-BSFCCNG =BSFCCNG × ----------------------HHV Gasoline

Figure 22. BSFC.

Figure 23. Normalized BSFC.

Figure 24. Lambda variation at full load, 90ºC coolant
temperature.
The normalized BSFC of gasoline is higher than for CNG
fuel because:
− A high compression ratio leads to slower ignition
timing for gasoline fuel and results in a BSFC (Brake
Specific Fuel Consumption) increase.
− As shown in Figures 24 and 25, because of the high
temperatures found at high rpm values at full load, the
fuel/air ratio must be richer to decrease the temperature
below the limit (800°C), this limit comes from material
temperature strength. As a consequence, the BSFC is
increased.
4.3. Peak Pressure
As shown in Figure 26, the position of peak pressure for


COMBUSTION DEVELOPMENT OF A BI-FUEL ENGINE

Figure 25. Exhaust gas temperature.

23

Figure 27. O2, full load condition, λ gasoline ≈ 0.9 and
λ CNG ≈ 0.96 .

Figure 26. Position of peak pressure.
CNG is between 12-16 CA-ATDC. For gasoline, this
position is between 16-28 CA-ATDC because of knocking.
4.4. Pollution
Pollution consists of Carbon monoxide, unburned hydrocarbons and nitrogen oxides; and these pollution from an
engine with CNG fuel follow a pattern similar to gasoline
fuel. The position of emissions sampling is before the
catalyst converter.
4.5. Hydrocarbons (HC)
The amount of unburned hydrocarbons for CNG fuel is
lower than for gasoline fuel. However, it is very much
dependent on engine specifications, like compression ratios
and running conditions (Duan, 1996).
Substantial oxidation of the hydrocarbons that escape the
primary combustion process by any of the processes can
occur during expansion and exhaust. The amount of oxidation depends on the time course of the temperature and
oxygen concentrations of these HCs as they mix with the
bulk gases (Duan, 1996), as shown in Figure 27. The O2
concentration in the exhaust of CNG fuel is higher than
gasoline fuel. This is a result of the lean CNG mixture used
in full load conditions and because the unburned hydrocarbons for CNG fuel can find oxygen and burn during the
expansion and exhaust stroke.
As shown in Figure 28, the amount of HC pollution in
exhaust gas is lower for CNG fuel than gasoline fuel in the
conditions tested in this study.

Figure 28. THC emissions (based on C1), full load

condition, λ gasoline ≈ 0.9 and λ CNG ≈ 0.96 .
Theoretically, the HC emissions from gas engines should
be lower due to the gaseous form of the fuel, which
provides excellent mixing. However, in some engines with
early exhaust valve opening and improper ignition timing,
because of the slow flame speed and slow reaction of
methane, the combustion was not completed before the
exhaust valve opened. This can be attributed to a high level
of HC.
It should be noted that UHC (Un burned Hydro Carbon)
from gas engines mainly consists of methane (approximately 85%), which is not photochemically reactive. Therefore, the non-methane hydrocarbons (NHMC) from gas
engines are extremely lower than from gasoline engines.
4.6. NOx
As shown in Figure 29, the amount of NOx in the exhaust
gas of CNG fuel is higher than for gasoline fuel.
The most important engine variables that affect NO
emissions are the fuel/air equivalence ratio, the burned gas
fraction of the in-cylinder unburned mixture, and the spark
timing. The burned gas fraction depends on the amount of
diluents, such as recycled exhaust gas (EGR) and the
residual gas fraction, used for emission controls. Fuel properties will affect burned gas conditions. The effect of
normal variations in gasoline properties is, however, modest.


24

O. S. ABIANEH, M. MIRSALIM and F. OMMI

Figure 29. NOx emission, full load conditions, λ gasoline ≈ 0.9
and λ CNG ≈ 0.96 .

Changes in the time course of temperature and oxygen
concentration in the burned gases during the entire combustion process and early part of the expansion stroke are
the important factors.
The peak pressure of CNG fuel is higher than the peak
pressure of gasoline fuel because of the advanced timing of
the ignition. This causes the amount of NOx to increase.
As the burned gases cool during the expansion stroke,
the reactions involving NO cease and leave NO concentrations far in excess of levels corresponding to equilibrium
at exhaust conditions. The isentropic coefficient of CNG is
higher than for gasoline and this causes the temperature of
the gas in the expansion stroke in CNG mode to decrease
more rapidly than the gasoline fuel. Therefore, the NOx
freezes rapidly in the CNG mode. Furthermore, advanced
ignition timing and a faster heat release rate in the sparkignited gas engine further increase peak temperatures
(Duan, 1996).
4.7. Carbon Monoxide (CO)
The carbon monoxide for CNG fuel is lower than for the
gasoline engine for all loads and speeds, due to its low
carbon content, and leaner mixture in full load condition, as
shown in Figures 30 and 31.
Carbon monoxide also forms during the combustion
process in rich fuel-air mixtures. In these mixtures, there is

Figure 30. CO emissions versus engine speed in full load
condition.

Figure 31. CO emissions versus BMEP (Brake Mean
Effective Pressure.
insufficient oxygen to fully burn the fuel. In addition, in
high-temperature products and lean mixtures, dissociation

ensures there are significant CO levels. Later, in the expansion stroke, the CO oxidation process also ceases as the
burned gas temperature falls. For fuel-rich mixtures, CO
concentrations in the exhaust increase steadily with increasing equivalence ratios, as the amount of excess fuel
increases.
The amount of CO in the exhaust gas of CNG is less
than in gasoline fuel because of its leaner mixture (Figure
31).

5. CONCLUSIONS
Natural gas is a type of hydrocarbon fuel with a simple
chemical structure. It is a clean fuel that is very suitable for
use in spark ignition engines. Its large reserve around the
world further encourages its application in transport and
power generation markets.
As a result, the following conclusions were reached:
(1) The decrease in power output usually found in the CNG
engine, resulting from the use of gaseous CNG fuel,
was minimized by increasing the high compression
ratio, adding increasing valve lift, optimizing valve timing
and reducing engine backpressure. However, the power
output needs to be improved further to equal that of
gasoline.
(2) The durability of the base gasoline engine is insufficient
for use with CNG because of the characteristic nature
of natural gas. Therefore, improvements were made to
the pistons, cylinder head, valves and valve seats for
both intake and exhaust systems. These characteristics
are not related to the high temperature of the combustion chamber wall, since a large difference between
the wall temperature of the two fuels, CNG and gasoline, was not seen in this study.
CNG is not an oily fuel and therefore causes damage to

the seat material, increasing the blow by (the oil and
fuel vapor) of the engine. This problem can be solved
by adding a passage for lubrication of the seat valve or
by changing the seat material.


COMBUSTION DEVELOPMENT OF A BI-FUEL ENGINE

(3) Combustion in full load mode of the engine using CNG
fuel is leaner than the combustion of the engine using
gasoline fuel.
(4) The CO and NMHC (Non Methane Hydro Carbon)
emissions from the engine running on natural gas are
lower than those of an engine running on gasoline
under similar conditions.
(5) It was demonstrated that, with careful design, natural
gas engines can achieve good performance, low emissions and are efficient enough to be realistic alternatives to diesel and gasoline engines.
(6) Decreasing the water temperature causes a decrease in
the temperature of the combustion chamber and cylinder walls. Most importantly, it causes a decrease in the
amount of NOx emitted and increases the amount of
HC pollutant.
(7) The wall temperature of CNG fuel is hotter than for
gasoline fuel but not enough to damage the cylinder or
cylinder head materials.
(8) The combustion duration of CNG fuel is more prolonged than for gasoline fuel.
(9) The normalized brake specific fuel consumption of the
engine tested with CNG fuel is lower than with gasoline fuel.

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