combustion analysis diesel-natural gas

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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,

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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

A.P. Carlucci et al. / Energy 33 (2008) 256–263 257



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.

A.P. Carlucci et al. / Energy 33 (2008) 256–263258



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




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

A.P. Carlucci et al. / Energy 33 (2008) 256–263260



(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_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.




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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