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Energy 33 (2008) 256–263
Experimental investigation and combustion analysis of a direct injection
dual-fuel diesel–natural gas engine
A.P. Carlucci, A. de RisiÃ, D. Laforgia, F. Naccarato
Department of Engineering for Innovation, University of Salento, CREA, via per Arnesano, 73100 Lecce, Italy
Received 11 December 2006
Abstract
A single-cylinder diesel engine has been converted into a dual-fuel engine to operate with natural gas together with a pilot injection of
diesel fuel used to ignite the CNG–air charge. The CNG was injected into the intake manifold via a gas injector on purpose designed for
this application. The main performance of the gas injector, such as flow coefficient, instantaneous mass flow rate, delay time between
electrical signal and opening of the injector, have been characterized by testing the injector in a constant-volume optical vessel. The CNG
jet structure has also been characterized by means of shadowgraphy technique.
The engine, operating in dual-fuel mode, has been tested on a wide range of operating conditions spanning different values of engine
load and speed. For all the tested operating conditions, the effect of CNG and diesel fuel injection pressure, together with the amount of
fuel injected during the pilot injection, were analyzed on the combustion development and, as a consequence, on the engine performance,
in terms of specific emission levels and fuel consumption.
r 2007 Elsevier Ltd. All rights reserved.
Keywords: Dual fuel; CNG injection; Pilot injection; Emissions; Fuel consumption
1. Introduction
Nowadays the internal combustion engines are spread to
the extent that they represent the main cause of pollutant
production. Nevertheless, it is well known that the stocks
of fuels traditionally used in this kind of engines will be
able to satisfy the world’s needs for few more decades. This
explains the massive research activity, drawn all over the
world, addressed to the utilization of innovative fuels and
injection concepts in order to either replace the traditional
ones or obtain a more efficient and clean combustion.Compared to diesel engines, characterized by a high
efficiency but at the same time high levels of particulate,
and to premixed charge gasoline engines, characterized by
a low efficiency because of knock limitations and pumping
losses, lean burn engines can reach a higher efficiency
thanks to lower pumping losses and heat transfer [1–5]. On
the contrary, lean mixtures generally imply higher levels of
both total unburned hydrocarbons (THC) and carbon
monoxide. Nevertheless, mixing the fuel with an increasing
quantity of air, a flame instability, sometimes leading to
misfiring, is observed.
Among the alternative fuels, methane is considered very
promising either because it can work with high compres-
sion ratios without experiencing the knock phenomenon or
because of its clean combustion. However, it is necessary to
prime the combustion. This can be obtained either using a
spark plug, similar to what happens in gasoline engines or
spraying a certain quantity of diesel fuel, whose ignitionand combustion sets the combustion of methane [6]. The
latter allows using methane either to supply most of the
thermal power required, therefore in percentages equal to
about 80–95% of the total required thermal power, or just
to ‘‘clean’’ the diffusive combustion phase of diesel fuel,
therefore in percentages not higher than 30%.
Previous works have shown that, using diesel fuel and
methane (dual-fuel) at the same time, allows to consider-
ably improve the NOx —particulate trade-off, keeping
substantially unchanged the total efficiency, but increasing,
ARTICLE IN PRESS
www.elsevier.com/locate/energy
0360-5442/$ - see front matterr 2007 Elsevier Ltd. All rights reserved.
doi:10.1016/j.energy.2007.06.005
ÃCorresponding author. Tel.: +390832 297756; fax: +39 0832 297777.
E-mail address: [email protected] (A. de Risi).
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and sometimes excessively, the levels of THC and CO in
the exhausts [7–9]. This inconvenient, more evident running
at low load is due to an air/methane mixture too lean and
therefore not able to propagate sufficiently fast in the
whole combustion chamber [10].
In order to avoid this problem, Huang et al. in Ref. [11]
have studied the effect of injection timing with respect to
the timing of the spark priming on the combustion of a
direct injection of natural gas inside a rapid compressionmachine. Tests have been done for different equivalence
ratios. The authors observed that, when the injection is
relatively advanced, the combustion is slow at the
beginning and then it becomes fast at the end. Assuming
that the combustion development is regulated either by the
stratification of the fuel charge and by the decrease of the
turbulence generated by the fuel jet, it has been inferred
that this combustion, similar to the combustion observed
in homogeneous charge engines, is produced by a low
stratification and after the turbulence generated by the fuel
jet is decreased. The authors observed a combustion
development more similar to the diesel engine combustion
if the ignition priming happened when the injection was
still taking place. On the other hand, they observed the
fastest combustion when the ignition priming coincided
with the injection end. It was inferred, then, that the
temporal interval between the end of the injection and
the ignition priming is a control parameter for both the
development and the lean limit of the combustion.
The aim of the present work is to study the effect of
compressed natural gas (CNG) and diesel fuel injection
pressure, and the diesel fuel injected quantity, for different
engine operating conditions, on combustion development
and engine emission levels, when both the fuels are used to
feed a direct injection diesel engine, and, in particular, the
diesel fuel is used in small quantities to ignite the indirect
injection of natural gas.
2. Characterization of the injector
Fig. 1 shows the scheme of the gas injection system used
during the experiments. It consists of a commercial gas
injector, a connecting steel duct and a spring-mountedpoppet assembly. The poppet is held by a spring and it
opens when the differential pressure overcomes the preload
of the spring. The duration of the valve opening, the mass
of the poppet and the characteristics of the injected gases
determine the duration of the injection and the poppet
oscillations. The value of the flow coefficient C d was
determined pumping air in the injector in steady-state
conditions. The air flow rate was then measured for
different values of valve opening and inlet air pressure ( p0in Fig. 2). The flow coefficient was then estimated as the
ratio between the real air flow rate and the theoretical one,_M th, calculated under the hypothesis of isentropic flow [12].
The reference flow cross-section, A, is equal to
A ¼ pDL sinW
2, (1)
where D is the poppet diameter, L the poppet lift and W is
the angle between the direction of the flow and the axis
of the injector. W/2 has been measured from pictures (see
Fig. 3) of the flux taken by means of a CCD camera using
the shadowgraphy technique. The dynamic behavior of the
injection system has been characterized as well. In
particular, the system has been mounted on the top of a
constant volume vessel hereafter referred to as bomb as
shown in the experimental layout of Fig. 2, and the
ARTICLE IN PRESS
Nomenclature
A CNG injector cross-reference exit area (m2)
Aht heat transfer area (m2)
dQnet/dCAD rate of net heat release (J/CAD)
dQw/dCAD rate of heat transferred to the walls(J/CAD)
D CNG injector poppet diameter (m2)
hc convective thermal coefficient (W/m2 K)
k specific heats ratio
L CNG injector poppet lift (m2)
p combustion chamber pressure (N/m2)
pinj CNG injection pressure (bar)
pbomb bomb pressure (bar)
T combustion chamber bulk temperature (K)
T w combustion chamber walls temperature (K)
V combustion chamber volume (m3)
W CNG flux direction (deg)
Abbreviations
1ATDC crank angle degrees After Top Dead Center (1)
1BTDC crank angle degrees Before Top Dead
Center (1)
BSCO brake specific carbon monoxide (g/kWh)BSTHC brake-specific total unburned hydrocarbons
(g/kWh)
BSNOx brake-specific nitric oxides (g/kWh)
BMEP brake mean effective pressure (bar)
CAD crank angle degree (deg)
CNG compressed natural gas
CO carbon monoxide
ECU electronic control unit
IVC inlet valve closing (deg)
NOx nitric oxides
ROHR rate of heat release (J/CAD)
SOI start of injection
THC total unburned hydrocarbons
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solenoid valve of the injector has been driven with a
square-like electric signal.
The pressure at the exit of the gas injector was then
measured by means of a piezoresistive absolute pressure
sensor, for different values of injection pressure, bomb
pressure and opening duration.
Fig. 4 shows the pressure traces measured by the
piezoresistive sensor when the injection pressure is equal
to 10 bar and the bomb pressure is varied in the range
2–6 bar. Because of leaks on the poppet valve, the pressure
in the injector body is equal to that downstream of the
poppet. Fig. 5 shows the delay time between the start of the
electric driving signal and the injector opening, as appeared
on the CCD shootings, for different injection pressures and
pressure within the pressurized vessel (bomb).
3. Engine tests
For the engine tests, the injector was mounted on the
intake manifold of a diesel engine about 80 mm upstream
of the intake valves as shown in Fig. 6.
The engine used for the tests is a four-valve single-
cylinder research engine. This engine was equipped with an
electronically controlled last-generation common-rail in-
jection system that allowed a full control of the injection
parameters.
ARTICLE IN PRESS
High pressure
air line
p0 T0
p
Piezoresistive
pressure sensor Injector tip
Fig. 2. Experimental layout for the dynamic behavior of the injectionsystem.
Fig. 3. Picture of the methane flux taken with a CCD camera using the
shadowgraphy technique.
0
2
4
6
8
10
12
0 0.02 0.04 0.06 0.08 0.1
time [sec]
p r e s s u r e
[ b a r ]
2 bar 6 bar
4 bar
Electric
driving signal
5 bar 3 bar
Fig. 4. Electric driving signal and pressure traces measured by a
piezoresistive sensor ( pinj ¼ 10 bar, pbomb ¼ 2–6bar).
Gas injector
Connecting
steel duct
Spring
mounted
poppetassembly
Fig. 1. Scheme of the gas injection system used during the experiments.
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The main characteristics of the engine are reported in
Table 1.
During the experimental tests, beside all engine para-
meters, methane and diesel fuel injector current and
injection pressures have been measured. The electronic
control of the methane injector was carried out by means of
an FPGA device whose controlling code was on purposerealized to properly work together with the main engine
electronic control unit (ECU). As for the main ECU, the
FPGA module establishes the crankshaft angular position
based on the signals of the same two inductive sensors used
by the main ECU and mounted on the flywheel and on the
camshaft, respectively.
To investigate the effect of phased CNG injection on the
combustion development and performance of the engine,
two among the most recurrent operating conditions in the
New European Driving Cycle (NEDC) (e.g. 1500 rpm with
4 bar brake mean effective pressure (bmep) and 2000 rpmwith 8 bar bmep) were investigated. In the following, each
operating condition will be referred with the notation
‘‘engine speedÂbmep’’.
For each of the two operating cases, three different
operating parameters were varied. In particular, diesel fuel
injection pressure and quantity and methane injection
pressure were varied on three levels, for two different
engine operating conditions. Therefore, the total number of
tested parameter sets was equal to 54. Table 2 reports the
matrix of experiments. For each engine parameter set, the
start of injection (SOI) of the diesel pilot injection, when
the engine was operated in the dual-fuel operating
conditions, was varied until the cylinder pressure peak
ARTICLE IN PRESS
Fig. 5. Delay between the electric driving signal and the injector opening
( pinj ¼ 8–10bar, pbomb ¼ 2–6bar).
Eddy-Current
Dynamometer
Angular
Reference
Cylinder
Pressure
E.C.U.
Data Acquisition BoardPC Based
Data Storage
and
Post-processing
System PC Based
Injection
Control and
E.C.U.
Monitoring
PC Based
Dynamometer
Monitoring
and
Control
Natural Gas
Compressed
Pressure valveControl
Gas Injector
Fig. 6. Experimental layout for tests on the engine.
Table 1
Engine characteristics
Bore (mm) 90
Stroke (mm) 85
Compression ratio 17.1:1
Injection system Common rail
Max. injection pressure (bar) 1300
Number of nozzles per injector 5
Nozzle diameter (mm) 0.170
Spray angle (deg) 142
Valve timing Opening ClosingIntake 13.51 BTDC 46.51 ABDC
Exhaust 51.51 BBDC 16.51 ATDC
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occurred at 101 crank angle degrees After Top Dead Center
(deg) (ATDC).
A test with the normal diesel operation (e.g. without
methane) was also carried out for the two selected engine
operating conditions. The injection strategy for those tests
consisted of a pilot and a main injection. The SOI of the
injections, for the 1500 rpm test, were at 241 crank angle
degrees Before Top Dead Center (deg) (BTDC) for the
pilot injection and 7.51 BTDC for the main injection, while
for the test at 2000 rpm, the SOI of the injections were at
391 BTDC for the pilot injection and 111 BTDC for the
main injection.
For all the investigated methane injection strategies, the
end of the electrical signal driving the injector was set at
2101 BTDC while the SOI was adjusted to mach the desired
injected methane mass. This was established in order to
make sure that all the methane injected into the intake
manifold could enter the cylinder. Please note that, as
reported in Table 1, the inlet valve closing (IVC) is at
133.51 BTDC, thus the 76.5 crank angle degree (CAD)
advance at 2000 rpm was just enough to compensate the
8.5 ms of delay between the end of the electrical signal and
the effective closure of the methane poppet valve.
Note that, hereafter, the results will be reportedaccording to the following notation. Three numbers have
been used to define a test carried out at a particular
parameter setup. The first number indicates the methane
injection pressure bar; the second the diesel injection
pressure in bar and the third the injected diesel quantity
mm3/stroke. For example, the notation 10_1000_6 indi-
cates the test was carried out setting the methane injection
pressure at 10 bar, the diesel injection pressure at 1000 bar
and injecting 6 mm3 of diesel fuel per stroke. During tests,
commercial diesel fuel and CNG with a methane percen-
tage of 99% were used.
Engine performance has been characterized in terms of
exhaust emissions. In particular, NOx, CO and THC
emission levels have been measured by sampling the
exhaust gases and analyzing them by means of an AVL
Digas 4000 exhaust gas analyzer. Smoke emissions have
been characterized in terms of opacity of the exhaust gases;
the opacity was measured by an AVL DiSmoke opaci-
meter. The error in measuring both THC and NOx levels
was 1 ppm vol., while for the opacity measurement it was
0.1%. The quantity of diesel fuel to be injected in each
cycle was obtained, once fixed the injection pressure,
estimating the energizing time of the injector opening valve
on the basis of the characterization of the same injector
done on a test bench. The low quantity of fuel injected, in
fact, did not allow a measure of fuel consumption based on
the weighting method. The CNG mass injected per stroke
was, on the contrary, estimated applying, during the
transitory injector working conditions, the data obtained
during the characterization of the injector in quasi-
stationary operations.
The in-cylinder pressure was measured using a Kistler
6053 piezoelectric pressure transducer. A Kistler 5044
charge amplifier was used to convert the electrical charge
yielded by the sensor into a proportional voltage. The
sensor sensitivity was À19 pC/bar, while the calibration
factor was 20Â105 Pa/V. The signal of the cylinder
pressure was digitized every 0.11 CA using an NI PCI
6052 data acquisition board. A mean combustion cycle was
obtained by averaging 50 cycles acquired in sequence. This
cycle was therefore filtered with a low-pass numeric filter
and then the rate of heat release (ROHR) was estimated
adding the net heat release rate, evaluated by means of the
traditional single-zone first law equation:
dQnet
dðCADÞ¼
k
k À 1 p
dV
dðCADÞþ
1
k À 1V
d p
dðCADÞ(2)
to the rate of heat transferred to the walls:
dQw
dðCADÞ¼ AhthcðT ÀT wÞ, (3)
where hc was estimated by means of Woschni model.
4. Results and discussion
Analyzing the data reported in the following, it can be
assumed that the methane/air mixture is homogeneous,
because of the methane injector location. Moreover, it is
important to mention, as previously said, that the diesel
fuel injection was performed only in order to ignite the
methane charge; consequently, tests related to 1500Â 4
operating condition differ from the ones related to 2000Â 8
operating condition not only for the different engine speed,
but also because of the methane/air ratio, which, for the
2000rpm tests, is almost double than in the tests at
1500 rpm, being the diesel injected quantity the same for
both tests.
Brake specific emissions for both the 1500Â 4 and
2000Â 8 tests are compared in Fig. 7. Standard case
emission levels (e.g. with the engine working only with
diesel fuel) are reported for comparison in the figure on the
left-end side of each graph with big spots.
Fig. 7 shows that the variation of brake-specific carbon
monoxide (BSCO) plot (a) and brake-specific nitric oxide
ARTICLE IN PRESS
Table 2
Engine parameter set matrix
Engine operating condition (engine speedÂbmep) Diesel injection pressure (bar) Gas oil inj. quantity (mm3) Methane inj. pressure (bar)
1500Â4 1000–800–600 8–6–4 10–7–5
2000Â8 1000–800–600 8–6–4 10–7–5
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(BSNOx) plot (b) levels, when varying the fuel quantity
injected during the pilot injection, is opposite. Brake
specific total unburned hydrocarbons (BSTHC) levels,
reported in plot (c), increase when decreasing the pilot
fuel amount for tests carried out at 1400Â 5, while they
seem to be insensitive to the same parameter when the
engine operates at 2000Â 8. BSPM emission levels, finally,
reported in the plot (d), do not show well defined variations
when varying the pilot fuel quantity.
ARTICLE IN PRESS
0
20
40
60
80
100
120
1500x4_BSCO (g/kWhr)
0
2
4
6
8
10
12
0
2
4
6
8
10
12
14
1500x4_BSPM
B S C
O [ g / k W h ]
0
20
40
60
80
100
120
B S C O
[ g / k W h ]
s t a n
d a r d
1 0
_ 1 0 0 0
_ 8
1 0
_ 1 0 0 0
_ 6
1 0
_ 1 0 0 0
_ 4
1 0
_ 8 0 0
_ 8
1 0
_ 8 0 0
_ 6
1 0
_ 8 0 0
_ 4
1 0
_ 6 0 0
_ 8
1 0
_ 6 0 0
_ 6
1 0
_ 6 0 0
_ 4
7_
1 0 0 0
_ 8
7_
1 0 0 0
_ 6
7_
1 0 0 0
_ 4
7_
8 0 0
_ 8
7_
8 0 0
_ 6
7_
8 0 0
_ 4
7_
6 0 0
_ 8
7_
6 0 0
_ 6
7_
6 0 0
_ 4
5_
1 0 0 0
_ 8
5_
1 0 0 0
_ 6
5_
1 0 0 0
_ 4
5_
8 0 0
_ 8
5_
8 0 0
_ 6
5_
8 0 0
_ 4
5_
6 0 0
_ 8
5_
6 0 0
_ 6
5_
6 0 0
_ 4
s t a n
d a r d
1 0
_ 1 0 0 0
_ 8
1 0
_ 1 0 0 0
_ 6
1 0
_ 1 0 0 0
_ 4
1 0
_ 8 0 0
_ 8
1 0
_ 8 0 0
_ 6
1 0
_ 8 0 0
_ 4
1 0
_ 6 0 0
_ 8
1 0
_ 6 0 0
_ 6
1 0
_ 6 0 0
_ 4
7_
1 0 0 0
_ 8
7_
1 0 0 0
_ 6
7_
1 0 0 0
_ 4
7_
8 0 0
_ 8
7_
8 0 0
_ 6
7_
8 0 0
_ 4
7_
6 0 0
_ 8
7_
6 0 0
_ 6
7_
6 0 0
_ 4
5_
1 0 0 0
_ 8
5_
1 0 0 0
_ 6
5_
1 0 0 0
_ 4
5_
8 0 0
_ 8
5_
8 0 0
_ 6
5_
8 0 0
_ 4
5_
6 0 0
_ 8
5_
6 0 0
_ 6
5_
6 0 0
_ 4
s t a n
d a r d
1 0
_ 1 0 0 0
_ 8
1 0
_ 1 0 0 0
_ 6
1 0
_ 1 0 0 0
_ 4
1 0
_ 8 0 0
_ 8
1 0
_ 8 0 0
_ 6
1 0
_ 8 0 0
_ 4
1 0
_ 6 0 0
_ 8
1 0
_ 6 0 0
_ 6
1 0
_ 6 0 0
_ 4
7_
1 0 0 0
_ 8
7_
1 0 0 0
_ 6
7_
1 0 0 0
_ 4
7_
8 0 0
_ 8
7_
8 0 0
_ 6
7_
8 0 0
_ 4
7_
6 0 0
_ 8
7_
6 0 0
_ 6
7_
6 0 0
_ 4
5_
1 0 0 0
_ 8
5_
1 0 0 0
_ 6
5_
1 0 0 0
_ 4
5_
8 0 0
_ 8
5_
8 0 0
_ 6
5_
8 0 0
_ 4
5_
6 0 0
_ 8
5_
6 0 0
_ 6
5_
6 0 0
_ 4
s t a n
d a r d
1 0
_ 1 0 0 0
_ 8
1 0
_ 1 0 0 0
_ 6
1 0
_ 1 0 0 0
_ 4
1 0
_ 8 0 0
_ 8
1 0
_ 8 0 0
_ 6
1 0
_ 8 0 0
_ 4
1 0
_ 6 0 0
_ 8
1 0
_ 6 0 0
_ 6
1 0
_ 6 0 0
_ 4
7_
1 0 0 0
_ 8
7_
1 0 0 0
_ 6
7_
1 0 0 0
_ 4
7_
8 0 0
_ 8
7_
8 0 0
_ 6
7_
8 0 0
_ 4
7_
6 0 0
_ 8
7_
6 0 0
_ 6
7_
6 0 0
_ 4
5_
1 0 0 0
_ 8
5_
1 0 0 0
_ 6
5_
1 0 0 0
_ 4
5_
8 0 0
_ 8
5_
8 0 0
_ 6
5_
8 0 0
_ 4
5_
6 0 0
_ 8
5_
6 0 0
_ 6
5_
6 0 0
_ 4
B S P M [ g / k W h ]
B S H C
[ g / k W h ]
2000x8_BSCO (g/kWhr)
1500x4_BSCO (g/kWhr)
2000x8_BSCO (g/kWhr)
1500x4_BSCO (g/kWhr)
2000x8_BSCO (g/kWhr)
2000x8_BSPM
Fig. 7. Brake-specific emissions for the tested operating conditions at
1500 rpm (black big spots refer to baseline—only diesel fuel—case).
30
60
90
0 20
30
40
50
60
70
10_600_410_600_610_600_810_600_410_600_610_600_8
10_600_4
10_600_6
10_600_8
10_600_4
10_600_6
10_600_8
7_1000_4
7_1000_6
7_1000_87_1000_4
7_1000_6
7_1000_8
7_1000_4
7_1000_6
7_1000_8
7_1000_4
7_1000_6
7_1000_8
150
120
R O H R [ J / C A D ]
50-10 -5
Crank Angle [°]
201510
50-10 -5
Crank Angle [°]
201510
50-10 -5
Crank Angle [°]
201510
c y
l i n d e r p r e s s u r e
[ b a r ]
30
60
90
0 20
30
40
50
60
70150
120
R O H
R [ J / C A D ]
30
60
90
0
150
120
R O H R [ J / C A D ]
c y
l i n d e r
p r e s s u r e
[ b a r ]
20
30
40
50
60
70
c y
l i n d e r p r e s s u r e [
b a r ]
7_600_4
7_600_6
7_600_8
7_600_4
7_600_6
7_600_8
Fig. 8. Pressure and ROHR histories for the experiments in dual-fuel
configuration (n ¼ 1500rpm) when varying the diesel fuel quantity and
pressure.
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The variation, previously described, of BSNOx levels as a
function of pilot and methane injection pressure suggests
that, although the combustion development in dual-fuel
engines depends on the combustion of both diesel fuel and
CNG–air mixture, BSNOx emissions are mainly produced
during the combustion of the pilot injection. In particular,
increasing the pressure of the pilot injection, the fuel, aswell known, is better atomized.
Therefore, when pilot fuel ignition takes place—the
ignition angle appears to be constant once the engine
operating condition is fixed—more ignition nuclei are
ready to burn, globally raising the ROHR peak of the
combustion of the premixed phase.
All these experimental evidences lead to the conclusion
that the quantity of methane is barely enough to sustain a
stable combustion for the tests at 1500 rpm, while it is rich
enough to have stable flame propagation in the test at
2000 rpm.
As a consequence, for the tests at 1500 rpm, the methane
combustion can only happen in the proximity of the
combustion locations of the diesel fuel.
This is supported by the plots of Fig. 8, in which the
initial slope of the ROHR curves does not change by
changing the diesel fuel injection parameters. On the other
hand, in the test at 2000 rpm, the ROHR curves of Fig. 9
show that the premixed combustion phase is not only
dependent on the amount of diesel fuel injected, but also on
the methane combustion which, in these tests, is able to
self-propagate after ignition occurs, thus modifying the
initial slope of the ROHR plots. The significant overall
increment of NOx is due to an increase in the combustion
temperature. In fact, the mean temperature during thecombustion, estimated with a ‘‘single-zone’’ model, was
equal to about 1700 K for the 1500 rpm tests and about
2400 K for the 2000 rpm tests.
5. Conclusions
Tests were carried out in order to study the combustion
development and its implications on the engine perfor-
mance, in terms of pollutant emission levels and fuel
consumption, on a dual-fuel CNG–air engine. During tests,
the engine was operated at two different conditions and,
for each of them, methane and diesel fuel injection
pressure, together with pilot fuel amount, were varied.
It was observed that an analysis of the ROHR is not
sufficient to explain the effect of each of the injection
parameters on the pollutant emissions. In the case of NOx,
it was found that the penetration of the jet holds the same
importance as the quantity of pilot fuel injected. The more
the jet penetrates into the combustion chamber, the more
its combustion will spread into the same chamber, and then
the local temperatures will be closer in value to the bulk
temperature. Similar conclusions can be drawn for the CO
and HC emission levels, although the latter specie seems
sometimes not to be sensitive to the injection parameters.PM levels do not show a well defined dependence on the
tested variables, but the corresponding levels, when
operating the engine in dual-fuel mode, are remarkably
lower than those observed with diesel fuel only.
Acknowledgment
This research has been supported by the Ministry of
Instruction, University and Research (MIUR).
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