Renewable and Sustainable Energy Reviews 13 (2009) 1151–1184

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Renewable and Sustainable Energy Reviews

journal homepage: www.elsevier.com/locate/rser

Effect of engine parameters and type of gaseous fuel on the performance

of dual-fuel gas diesel engines—A critical review
B.B. Sahoo a, N. Sahoo b, U.K. Saha b,*
a
Centre for Energy, Indian Institute of Technology, Guwahati 781039, India
b
Department of Mechanical Engineering, Indian Institute of Technology, Guwahati 781039, India

A R T I C L E I N F O A B S T R A C T

Article history: Petroleum resources are finite and, therefore, search for their alternative non-petroleum fuels for internal
Received 7 January 2008 combustion engines is continuing all over the world. Moreover gases emitted by petroleum fuel driven
Accepted 20 August 2008 vehicles have an adverse effect on the environment and human health. There is universal acceptance of
the need to reduce such emissions. Towards this, scientists have proposed various solutions for diesel
Keywords: engines, one of which is the use of gaseous fuels as a supplement for liquid diesel fuel. These engines,
Combustion
which use conventional diesel fuel and gaseous fuel, are referred to as ‘dual-fuel engines’. Natural gas and
Carbon dioxide
bio-derived gas appear more attractive alternative fuels for dual-fuel engines in view of their friendly
Dual-fuel
Efficiency
environmental nature. In the gas-fumigated dual-fuel engine, the primary fuel is mixed outside the
Emissions cylinder before it is inducted into the cylinder. A pilot quantity of liquid fuel is injected towards the end of
Gaseous fuel the compression stroke to initiate combustion. When considering a gaseous fuel for use in existing diesel
engines, a number of issues which include, the effects of engine operating and design parameters, and
type of gaseous fuel, on the performance of the dual-fuel engines, are important. This paper reviews the
research on above issues carried out by various scientists in different diesel engines. This paper touches
upon performance, combustion and emission characteristics of dual-fuel engines which use natural gas,
biogas, producer gas, methane, liquefied petroleum gas, propane, etc. as gaseous fuel. It reveals that ‘dual-
fuel concept’ is a promising technique for controlling both NOx and soot emissions even on existing diesel
engine. But, HC, CO emissions and ‘bsfc’ are higher for part load gas diesel engine operations. Thermal
efficiency of dual-fuel engines improve either with increased engine speed, or with advanced injection
timings, or with increased amount of pilot fuel. The ignition characteristics of the gaseous fuels need
more research for a long-term use in a dual-fuel engine. It is found that, the selection of engine operating
and design parameters play a vital role in minimizing the performance divergences between an existing
diesel engine and a ‘gas diesel engine’.
ß 2008 Elsevier Ltd. All rights reserved.

Contents

1. Introduction . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 1152
2. Dual-fuel concept. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 1153
3. Modification in internal combustion gas engines. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 1153
4. Combustion processes in gas diesel engine and conventional diesel engine . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 1153
5. Performance of ‘‘gas diesel engines’’ . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 1155

* Corresponding author. Tel.: +91 361 2582663; fax: +91 361 2690762/2582699.
E-mail address: saha@iitg.ernet.in (U.K. Saha).
Abbreviations: ATDC, after top dead centre; BMEP, brake mean effective pressure; bsfc, brake specific fuel consumption; BTDC, before top dead centre; C, centigrade; CA, crank
angle; DI, direct injection; DG, diesel genset; FD, fossil-diesel; HRR, heat release rate; IDI, indirect injection; J, joule; kW, kilo Watt; m, mass, meter; NOx, oxides of nitrogen; P,
pressure; Pa, pascal; rpm, revolutions per minute; T, torque; UBHC, unburned hydrocarbon; BSU, Bosch smoke unit; BTE, brake thermal efficiency; CR, compression ratio; BDC,
bottom dead centre; CO, carbon monoxide; db, decibel; deg., degree; h, hour; HC, hydrocarbon; IT, injection timing; K, Kelvin; l, litre; M, mega; N, Newton; NG, natural gas;
ppm, parts per million; rev, revolutions; RHR, rate of heat release; SOx, oxides of sulphur; TDC, top dead centre.

1364-0321/$ – see front matter ß 2008 Elsevier Ltd. All rights reserved.
doi:10.1016/j.rser.2008.08.003
1152 B.B. Sahoo et al. / Renewable and Sustainable Energy Reviews 13 (2009) 1151–1184

5.1. Effect of engine load. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 1155

5.2. Effect of engine speed . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 1159
5.3. Effect of pilot fuel injection timing . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 1161
5.4. Effect of pilot fuel mass inducted . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 1169
5.5. Effect of engine compression ratio . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 1173
5.6. Effect of intake manifold conditions . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 1173
5.7. Effect of type of gaseous fuel. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 1174
6. Conclusion . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 1182
6.1. Effect of engine load. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 1182
6.2. Effect of engine speed . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 1182
6.3. Effect of pilot fuel injection timing . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 1182
6.4. Effect of mass of pilot fuel inducted . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 1182
6.5. Effect of engine compression ratio . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 1182
6.6. Effect of engine intake manifold conditions . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 1182
6.7. Effect of type of gaseous fuel. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 1183
Acknowledgements . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 1183
References . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 1183

1. Introduction and combustion. Therefore, it is more economical and of

environmental advantage to use gaseous fuel in diesel engines
National interest in generating alternative fuels for internal that use the ‘dual-fuel concept’ [4,5].
combustion (IC) engines continues to be strong due to environ- Due to limited resources of fossil fuels, alternative solutions have
mental concerns and/or the uncertainties associated with the been proposed by many scientists. The ‘‘dual-fuel concept’’ is one of
future availability of fossil fuel. Mostly, the interest lies in them that use both conventional diesel and gaseous fuels. The lower
identifying alternative sources of fuel energy supply. The motiva- running costs and the use of alternative fuel sources with dual-fuel
tion for the need for diversification has prompted the research engine operation have attracted many investigators to use this
worldwide into potential alternative sources of fuel energy for IC engine in different areas of interest. The earliest experiments on
engines. One feature of alternative sources of fuel energy is that dual-fuel system were performed by Cave in 1929, and Helmore and
they are well suited for decentralised development to meet the Sokes in 1930, in which burning hydrogen is induced as a secondary
needs for social and economic progress, especially in rural fuel in diesel engines. When hydrogen burn completely, there was
communities where fossil fuels may be difficult or expensive to reduction of liquid fuel load and saving of 20% diesel fuel [6].
obtain. Natural gas, bio-derived gas and liquids appear more However, at that time, the dual-fuel engine was not used
attractive in view of their friendly environmental nature [1,2]. commercially due to its mechanical complexity and rough running
The gaseous fuels are getting more positive response from caused by auto-ignition and knocking at relatively low compression
researchers and end-users compared to past because of current ratios. In 1939, the first commercial dual-fuel engine, fuelled by
unfolding developments. Gas is clearly the fossil fuel of least town gas or other types of gaseous fuels, was produced by the
environmental impact. When burnt, it produces virtually insig- National Gas and Oil Engine Co. in Great Britain. This type of engine
nificant SOx and relatively little NOx, the main constituents of acid was relatively simple in operation and was mainly employed in
rain, and substantially less CO2, a key culprit in the greenhouse some areas where cheap stationary power production was required.
debate, than most oil products and coal [3]. During the Second World War, scientists in Great Britain, Germany
Most combustion devices are easily adaptable to the use of and Italy paid more attention to the possible application of dual-fuel
gaseous fuels for power production. High compression ratio piston engines in civil and military areas due to the shortage of liquid fuels.
engines are generally suitable when high thermal efficiency is Some vehicles with diesel engines were successfully converted to
desired. Gaseous fuels have high octane numbers, and therefore, dual-fuel running. Different kinds of gaseous fuels, such as coal gas,
suitable for engines with relatively high compression ratio. sewage gas or methane, were employed in conventional diesel
Gaseous fuels also promise to be suitable for higher compression engines [7]. After the Second World War, due to economical and
engines, since it is known that they resist knock more than environmental reasons, dual-fuel engines have been further
conventional liquid fuels, as well as producing less polluting developed and employed in a very wide range of applications from
exhaust gases if appropriate conditions are satisfied for its mixing stationary power production to road and marine transport, such as in
long and short haul trucks and buses. Some conversions from the
original compression ignition diesel engines to dual-fuel operation
Nomenclature are made by manufacturers utilizing a double plunger system or two
pumps in the injection system of the engine to handle the small
quantity of diesel fuel required for ignition [8].
Greek letters Diesel engines, with appropriate relatively simple conversion,
h efficiency can be made to operate on gaseous fuels efficiently. Such engines,
u crank angle which are called ‘dual-fuel engines’ or ‘gas diesel engines’, usually
8 degree have the gaseous fuel mixed with the air in the engine cylinders,
either through direct mixing in the intake manifold with air or
through injection directly into the cylinder. The resulting mixture
Subscripts
after compression is then ignited through the injection of a small
d dual-fuel amount of diesel fuel (the pilot) in the usual way. This pilot liquid
g gas fuel auto-ignites to provide ignition sources for subsequent flame
propagation within the surrounding gaseous fuel mixture. Unlike
 B.B. Sahoo et al. / Renewable and Sustainable Energy Reviews 13 (2009) 1151–1184 1153

the spark ignited gas engine, which requires an adequate and mixture is then ignited by energy from the combustion of the
uninterrupted gas supply, most current dual-fuel engines are made diesel fuel spray, which is called pilot fuel. The amount of pilot fuel
to operate interchangeably, either on gaseous fuels with diesel needed for this ignition is between 10% and 20% of the operation on
pilot ignition or wholly on liquid fuel injection as a diesel engine. diesel alone at normal working loads and the amount differs with
Accordingly, a dual-fuel engine tends to retain most of the positive the point of engine operation and its design parameters. During
features of diesel operation [9]. Even it surpasses occasionally part load engine operation, the fuel gas supply is reduced by means
those of the diesels, producing higher power outputs and of a gas control valve. However, a simultaneous reduction of the air
efficiencies. This is achieved without significant smoke or supply decreases the air quantity induced. Hence, the compression
particulates emission and with reduced NOx production [10], pressure and the mean effective pressure of the engine are
while having reduced peak cylinder pressures and quieter decreased. This would eventually lead to a drop in power and
operation. efficiency. The drastic reduction in the compression conditions
When considering a gaseous fuel for use in existing diesel might even become too weak for the mixture to effect self-ignition.
engines, a number of issues are important. These issues include the Therefore, dual-fuel engines should not to be throttled/controlled
engine operating and design parameters, and type of gaseous fuel on the air side.
supply to the engine. The purpose of this review is to discuss the Ideally, there is a need for optimum variation in the liquid pilot
effect of engine operating and design parameters, and type of fuel quantity used any time in relation to the gaseous fuel supply so
gaseous fuel on the performance of the gas diesel engine. The as to provide for any specific engine the best performance over the
engine operating and design parameters include load, speed, whole load range desired [12]. Usually, the main goal, for both
compression ratio, pilot fuel injection timing, pilot fuel mass emissions and economic reasons, is to minimize the use of the
inducted, intake manifold conditions, and type of gaseous fuel. diesel fuel and maximize its replacement by the cheaper gaseous
The effects of these parameters on performance, combustion and fuel throughout the whole engine load range. The dual-fuel engine
emission characteristics of dual-fuel diesel engines are also can operate effectively on a wide range of different gaseous fuels
examined. This paper touches upon dual-fuel engines that use while maintaining the capacity for operation as a conventional
natural gas, biogas, producer gas, methane, liquefied petroleum diesel engine. Normally, the change over from dual-fuel to diesel
gas, propane, etc. as gaseous fuel. An attempt is made here to operation and vice versa can be made automatically, even under
review the previous studies to look into further improvement of load conditions [3].
‘‘gas diesel engines’’ from the viewpoint of performance, combus-
tion and emission. 4. Combustion processes in gas diesel engine and conventional
diesel engine
2. Dual-fuel concept
The combustion processes in CI engines running on pure diesel
Available technologies for reciprocating IC engines are gen- fuel can be divided into four stages as shown in Fig. 1. They are, A–
erally divided in two categories: compression-ignition (CI) and B: period of ignition delay; B–C: premixed (rapid pressure rise)
spark-ignition (SI) engines. In CI engines (diesel engines), air is combustion; C–D: controlled (normal) combustion; and D–E: late
compressed at pressures and temperatures at which the injected combustion. Point ‘A’ is the start of fuel injection and ‘B’ for start of
liquid fuel fires easily and burns progressively after ignition. combustion.
Whereas, SI engines (Otto engines) that runs according to the Beau However, the combustion processes in gas-fumigated dual-fuel
de Rochas cycle [11], the carburetted mixture of air and vaporized engines using pilot injection have been identified to take place in
fuel (high octane index) is compressed under its ignition point and five stages as shown in Fig. 2. They are the pilot ignition delay (AB),
then fired at a chosen instant by an independent means. pilot premixed combustion (BC), primary fuel ignition delay (CD),
In a dual-fuel engine, both types of above combustion coexist rapid combustion of primary fuel (DE) and the diffusion combus-
together, i.e. a carburetted mixture of air and high octane index tion stage (EF).
gaseous fuel is compressed like in a conventional diesel engine. The Ignition delay (AB) of injected pilot fuel exists longer than the
compressed mixture of air and gaseous fuel does not auto-ignite pure diesel fuel operation. This is due to the reduction in oxygen
due to its high auto-ignition temperature. Hence, it is fired by a concentration resulting from gaseous fuel substitution for air. The
small liquid fuel injection which ignites spontaneously at the end pressure rise (BC) is moderately low as compared to pure diesel
of compression phase. The advantage of this type of engine is that,
it uses the difference of flammability of two used fuels. Again, in
case of lack of gaseous fuel, this engine runs according to the diesel
cycle by switching from dual-fuel mode. The disadvantage is the
necessity to have liquid diesel fuel available for the dual-fuel
engine operation [12].

3. Modification in internal combustion gas engines

A SI gas engine modification is comparatively easy as it is

already designed to operate on air–fuel mixture with spark
ignition. The basic modification needed in this type of engine is the
provision of a gas–air mixer instead of the carburettor. The engine
control is performed by the mixture supply variation, i.e. throttle
valve position like in the case with petrol fuel.
In dual-fuel gas diesel engines, the normal diesel fuel injection
system still supplies a certain amount of diesel fuel. The engine
however induces and compresses a mixture of air and gaseous fuel
that is prepared in the external mixing device. The compressed Fig. 1. Details of combustion processes in diesel engine [13].
1154 B.B. Sahoo et al. / Renewable and Sustainable Energy Reviews 13 (2009) 1151–1184

Fig. 2. Dual-fuel pilot injection pressure–crank angle diagram [13].

Fig. 4. Dual-fuel pilot injection heat release diagram (engine speed: 3600 rev/min,
engine torque output, N-m: 5.15) [13].

However, this ignition delay is short as compared with the initial

delay period due to the pilot fuel injection. The pressure decreases
slowly (CD) until the actual combustion of the fumigated gas starts.
The phase of combustion (DE) is very unstable because it started
with flame propagation that has been initiated by the spontaneous
ignition of pilot fuel. The pressure rise here does not cause any
operating problem since it occurs in an increasing cylinder volume.
Diffusion combustion stage (EF) starts at the end of rapid pressure
rise and continues well into the expansion stroke. This is due to the
slower burning rate of gaseous fuel and the presence of diluents
from the pilot fuel. Some gas–air mixture may escape combustion
under this phase due to low oxygen concentration, valve overlap,
flame quenching on the walls or the effects of crevices. The success
of this phase primarily depends on the length of ignition delay.
Fig. 3. Heat release diagram of diesel fuel operation (engine speed: 3000 rev/min, The heat release diagram of diesel fuel operation (Fig. 3) shows
engine torque output, N-m: 9.65) [13]. an apparent negative heat release prior to the main start of
combustion. This is due to the cooling effect of the injected liquid
fuel operation due to the ignition of small quantity of pilot fuel. fuel. This effect is not apparent with the dual-fuel combustion
Again, there is a finite time lag between the development of the processes as pre-oxidation of the gaseous fuel starts before pilot
first and second pressure rises due to a longer ignition delay of gas– fuel injection (Fig. 4). The pilot fuel is then flattened the heat
air mixture, a result of the high self-ignition temperature. release rate within this region. Fig. 4 shows the initial heat release
Table 1
Summary of engine type and dual-fuel used in the experimental investigation by various researcher(s).

Researcher(s) Test engine Pilot fuel Primary fuel

Nwafor [2,13,37] Petter model AC1 single cylinder, air-cooled, high speed, IDI, four-stroke diesel engine Diesel NG
Bari [40] Two cylinder four stroke cycle diesel engine (16.8 kW at 1500 rpm, Model-2105 Nang Diesel Biogas
Chang Company, China), water cooled, naturally aspirated with double swirl
combustion chamber
Singh et al. [30] A naturally aspirated multi cylinder DG with matching alternator FD Producer gas
Mansour et al. [12] Naturally aspirated, V-8 Deutz FL8 413F four cycle diesel engine Diesel NG
Papagiannakis and Hountalas [19] Single cylinder, naturally aspirated, four stroke, air cooled, direct injection, high speed, Diesel NG
Lister LV1 DI diesel engine with a bowl in piston combustion chamber
Selim [14] Ricardo E6 single cylinder variable compression IDI diesel engine Diesel Compressed NG
Selim [4] Ricardo E6 single cylinder variable compression IDI diesel engine Diesel CH4, CNG, LPG
Nwafor and Rice [38] Petter model AC1 single cylinder, air-cooled, high speed, IDI, four-stroke diesel engine Diesel NG
Henham and Makkar [3] Two-cylinder, four-stroke, water-cooled, IDI Lister Petter LPWS2 diesel engine Gasoil Biogas
Papagiannakis and Hountalas [24] Single cylinder, naturally aspirated, four stroke, air cooled, DI, high speed, Lister LV1 Diesel NG
diesel engine with a bowl in piston combustion chamber
Krishnan et al. [35] Single-cylinder DI, CI engine Diesel NG
Kusaka et. al. [39] Water cooled, 4-stroke-cycle, and 4-cylinder conventional DI diesel engine Diesel NG
Uma et al. [16] Direct injected six cylinder, vertical, four stroke engine with mechanical injector Diesel Producer gas
Badr et al. [9] Two single cylinder, 4-stroke, water cooled, DI, normally aspirated laboratory Diesel Methane
dual-fuel engines
Abd et al. [33,36] Single cylinder, four stroke, water cooled engine (Ricardo E6) Diesel Methane, propane
 B.B. Sahoo et al. / Renewable and Sustainable Energy Reviews 13 (2009) 1151–1184 1155

due to pilot injection. This figure also indicates that dual-fuel

combustion is continued well into the expansion stroke [13].

5. Performance of ‘‘gas diesel engines’’

Gaseous fuels are considered to be good alternative fuels for

passenger cars, truck transportation and stationary engines that
can provide both good environmental effect and energy security
[4]. However, as some of the engine operating and design
parameters namely, load, speed, compression ratio, pilot fuel
injection timing, pilot fuel mass, inlet manifold condition,
composition of gaseous fuel candidates vary, the performance of
the dual-fuel gaseous engines are affected. Numerous studies have
been carried out by many researchers to examine the effect of the
above-mentioned parameters on the emission, performance and
combustion characteristics of dual-fuel engines. These investiga-
tions by the researchers are conducted in different test engines
with various gaseous primary fuels and pilot fuels. The type of test
engine and fuels for dual fuel operation used by several
researcher(s) are given in Table 1.

Fig. 5. Effect of engine load on pressure rise rate for the diesel and dual-fuel engines
5.1. Effect of engine load [14].

The effect of load on combustion noise for the diesel and dual
fuel engine (Fig. 5), at an engine speed of 1200 rpm is examined by pilot diesel which then auto-ignites and starts burning the gaseous
Selim [14]. The pressure rise rate (combustion noise) for the diesel fuel at a higher rate of pressure rise. This is also shown by Nielsen
engine increases slightly when the load increases. For the dual-fuel et al. [15] on dual-fuel engine where natural gas is admitted in the
mode, the combustion noise also increases when the load increases inlet air manifold.
and is always higher than that for the diesel fuel case. Combustion Concentration of pollutants is investigated with diesel alone
noise for the diesel case is about 4 bar/8C A and it only fluctuates and dual-fuel mode (producer gas) at different loads (10, 20, 30 and
slightly around this value. For the dual-fuel engine, it increases 40 kW) by Uma et al. [16] and is given in Table 2. NOx emissions in
from 4 bar/8C A at a load of 4.5 N-m to 15.5 bar/8C A at 18.5 N-m. dual-fuel mode are lower than the emissions from diesel engine in
Increasing the load at constant speed results in an increase in the diesel alone mode. SO2 levels are low in dual fuel mode. This is due
mass of gaseous fuel admitted to the engine, since the pilot mass to low sulphur content in biomass fuel. The CO emission in dual-
injected remains constant at all loads. This increase in the mass of fuel mode is higher than that of diesel alone. High concentration of
methane then causes an increase in the ignition delay period of CO in the dual-fuel exhaust is an indication of incomplete

Table 2
Concentration of pollutants from diesel engine in diesel alone and dual-fuel mode [16].

Parameter Load (kW)

10 20 30 40

Diesel Dual-fuel Diesel Dual-fuel Diesel Dual-fuel Diesel Dual-fuel

CO (ppm) 181 635 207 640 284 734 323 904

174 661 218 681 275 693 303 941
172 672 232 696 294 705 336 922

CO2 (ppm) 3.1 6.2 4.2 7.1 5.7 9.2 6.1 11.0
3.0 6.0 4.1 7.2 5.8 9.1 6.1 10.9
3.2 6.2 4.2 7.2 5.9 9.3 6.2 11.1

HC (ppm) 109 119 132 141 180 182 270 262

112 127 125 136 188 178 271 288
103 113 116 121 187 192 276 284

CH4 (ppm) 7.0 18 8.4 24 10.2 21 12.1 32

7.2 12 9.2 27 10.1 32 12.3 30
7.0 14 8.7 30 10.2 25 11.8 36

SO2 (ppm) 4.6 1.1 5.4 1.2 6.8 1.5 9.6 2.3
4.2 1.2 5.0 1.2 6.9 1.6 9.6 2.3
3.9 1.2 5.8 1.3 8.4 1.9 10.3 1.9

NO2 (ppm) 172 93 230 140 279 170 412 240

181 98 229 145 282 161 403 249
188 101 224 137 279 171 405 253

Particulates (mg/m3) 22 18 26 24 29 24 33 28
20 20 20 27 23 18 27 22
27 22 32 24 32 29 36 40
1156 B.B. Sahoo et al. / Renewable and Sustainable Energy Reviews 13 (2009) 1151–1184

Table 3
Fuel consumption and SEC of diesel engine in diesel alone and dual-fuel mode [16].

Load (kW) Fuel consumption SEC (MJ/kWh)

Diesel mode Dual-fuel mode Diesel mode Dual-fuel mode

Diesel (kg/h) Diesel (kg/h) Producer gas (Nm3/h)

10 5.3 1.9 57 22.8 34

20 7.2 1.3 66 15.5 18
30 9.8 1.5 81 14.0 15
40 12.2 3.7 112 13.1 16

combustion. At part load condition, concentration of CO increases. An experimental investigation has been conducted by Papa-
This also suggests the need for lower load limit for dual-fuel giannakis and Hountalas [19] to examine the effect of dual fuel
operation. At part load condition, the specific energy consumption combustion on the performance and pollutant emissions of a DI
(SEC) also increases. High CO emission in dual-fuel mode operation diesel engine. The engine has been properly modified to operate
is due to combination of factors such as low heating value of gas, under dual-fuel operation. The air inlet temperature is kept 23 8C
low adiabatic flame temperatures, and low mean effective for all cases. Measurements are taken at four different engine loads
pressures. Additionally, the engines are not actually designed for corresponding to 20%, 40%, 60%, and 80% of full load and three
producer gas operation. But these issues can be resolved if efforts engine speeds of 1500, 2000 and 2500 rpm under both normal
are focused on development of new engine designs for producer diesel operation (only diesel fuel) and dual-fuel operation with NG
gas operation. Apart from injector design, other parameters such as and pilot injection of diesel fuel. Under dual-fuel operation, an
compression ratio, ignition advance, combustion chamber design, effort has been made to keep the pilot amount of diesel fuel
etc. will have to be optimized to produce low CO emissions. HC constant, while the power output of the engine is adjusted through
emissions in dual-fuel mode are little lower than HC emissions in the amount of gaseous fuel. Referring to Fig. 6, the term ‘z’ refers to
diesel alone mode. the percentage of gaseous fuel. At part engine load, cylinder
The SEC increases with decreasing load both in diesel alone and pressure under dual-fuel operation diverges from the respective
dual-fuel mode (Table 3). This implies the considerable efficiency values under normal diesel operation. The lower cylinder pressures
loss at low-load condition. SEC in dual-fuel mode is higher than the observed under dual-fuel operation during the compression stroke
diesel mode throughout the tested load condition. Increased SEC are the result of the higher specific heat capacity of the NG–air
indicates the efficiency reduction in the dual-fuel mode. This is due mixture. The total heat release rate under dual-fuel operation is
to reduced heating value of the producer gas–air mixture and drop slightly higher compared to the one under normal diesel operation
in the pressure of the gas entering the air inlet and lower flame (Fig. 6); revealing late combustion of the gaseous fuel. But, the
velocity. The results are similar to the earlier studies [17,18] which effect on the cylinder pressure is small since it is in the expansion
also reported de-rating of diesel engine operated in dual-fuel stroke. At high engine load (Fig. 7), the cylinder pressure traces
mode. under dual-fuel operation diverge again from the respective values
under normal diesel operation during the compression stroke and
the initial stages of combustion. This difference is again more
evident at low engine speed where the combustion rate under

Fig. 6. Cylinder pressure and total heat release traces under normal diesel (z = 0%) Fig. 7. Cylinder pressure and total heat release traces under normal diesel (z = 0%)
and dual-fuel (z > 0%) operation for 1500 and 2500 rpm engine speed at 40% load and dual-fuel (z > 0%) operation for 1500 and 2500 rpm engine speed at 80% load
[19]. [19].
 B.B. Sahoo et al. / Renewable and Sustainable Energy Reviews 13 (2009) 1151–1184 1157

Fig. 10. Variation of nitric oxide under normal diesel and dual-fuel operation versus
Fig. 8. Variation of combustion duration under normal diesel and dual-fuel load at 1500 and 2500 rpm engine speed [19].
operation as function of load at 1500 and 2500 rpm engine speed [19].

dual-fuel operation during the premixed controlled combustion mainly, due to the lower temperature and air–fuel ratio inside the
phase is significantly lower compared to the one under normal combustion chamber of the engines, resulting in a slower
diesel operation. It is revealed that the total rate of heat release combustion rate as observed from the results of the heat release
under dual-fuel operation is obviously higher compared to the one rate analysis [11,20]. On the other hand, at high load, the
under normal diesel operation (Fig. 7). The effect is stronger at low improvement of gaseous fuel utilization leads to a relevant
engine speed, revealing later combustion of the gaseous fuel and improvement of the total ‘bsfc’ under dual-fuel operation, which
this obviously has an effect on the ‘bsfc’. tends to converge to the one under normal diesel operation. But its
The combustion duration is higher under dual-fuel operation at value continues to be higher compared to the one under normal
low engine load but drops with the increase of load (Fig. 8). diesel operation.
Especially, at low engine loads, the combustion duration, even The formation of nitric oxide (NO) is favored by high oxygen
tends to become lower compared to normal diesel operation. At concentration and high charge temperature [11,21,22]. NOx
low engine loads, the total ‘bsfc’ for dual-fuel operation is concentration is affected considerably by the presence of gaseous
considerably higher compared to the one under normal diesel fuel–air mixture. The concentration of NOx under dual-fuel
operation (Fig. 9). This reveals a poor utilization of the gaseous fuel, operation is lower compared to the one under normal diesel

Fig. 9. Variation of total ‘bsfc’ under normal diesel and dual-fuel operation as Fig. 11. CO under normal diesel and dual-fuel operation versus load at 1500 and
function of load at 1500 and 2500 rpm engine speed [19]. 2500 rpm engine speed [19].
1158 B.B. Sahoo et al. / Renewable and Sustainable Energy Reviews 13 (2009) 1151–1184

Fig. 12. UBHC under normal diesel and dual-fuel operation versus load at 1500 and
2500 rpm engine speed [19].

Fig. 13. Soot emissions under normal diesel and dual-fuel operation versus load at
operation at the same engine operating conditions (i.e. engine 1500 and 2500 rpm engine speed [19].

speed and load). At low engine loads, the NOx concentration under
dual-fuel operation is slightly lower compared to the one under
normal diesel operation (Fig. 10). This is mainly as a result of the With the increase of engine load, there is a sharp decrease of HC
lower rate of premixed controlled combustion of the gaseous fuel, emissions under dual-fuel operation. This is the result of the
which results in lower charge temperature inside the combustion increase of burned gas temperature that helps oxidize efficiently
chamber compared to normal diesel operation. At higher load, the the UBHCs. But for all cases examined, the HC emissions are
NOx concentration under dual-fuel operation is considerably lower considerably higher under dual-fuel operation compared to
compared to the one under normal diesel operation (Fig. 10). As far normal diesel operation. From Fig. 13, it can be seen that soot
as the effect of engine speed is concerned, the increase of engine emissions under dual-fuel operation are considerably lower
speed under dual-fuel operation results in a further decrease of compared to the ones under normal diesel operation for all cases
NOx values compared to normal diesel operation. The rate of CO examined. Under normal diesel operation, soot emissions increase
formation is a function of the unburned gaseous fuel availability with increasing engine load. On the other hand, under dual-fuel
and mixture temperature which controls the rate of fuel operation and for all cases examined, the soot emissions do not
decomposition and oxidation [11,21,23]. The CO emissions under follow the same trend since a reduction of soot emissions is
dual-fuel operation are significantly higher as shown in Fig. 11. At observed with increasing engine load. This is to be expected since
low engine speed, CO concentration under dual-fuel operation the increase of engine load is accomplished by increasing the
clearly decreases with the increase of engine load. This is the result amount of gaseous fuel that forms no soot while the increasing
of the improvement of gaseous fuel utilization especially during charge temperature contributes to its oxidation [24–29].
the second phase of combustion. At high engine speed, the increase Performance of DG set in dual-fuel mode (FD + producer gas) at
of engine load does not seem to affect the concentration of CO due different engine load conditions is investigated by Singh et al. [30].
to the less time available for combustion. Performance of DG is evaluated in terms of SEC, brake thermal
At low engine load, HC emissions under dual-fuel operation are efficiency, engine output and sound pressure level. In operating CI
considerably higher compared to the ones under normal diesel engine on dual-fuel mode it is noted that there is a minor reduction
operation (Fig. 12). This is mainly due to the lower charge in engine output about 1–2% (Table 4). This reduction is expected,
temperature and air–fuel ratio, resulting in slower combustion and because the calorific value of air–producer gas mixture is lower than
allowing small quantities of fuel to escape the combustion process. that of air–liquid fuel vapour mixture. The engine performance is

Table 4
Effect of fuel and load on engine output, SEC, brake thermal efficiency and sound pressure [30].
Engine load (%) 63 63 84 84 98 98
Mode of operation FD Dual-fuel FD Dual-fuel FD Dual-fuel
RPM of engine 1487 1490 1493 1483 1486 1489
Engine output (kW) 14.08 14.00 19.34 18.93 22.59 22.00
LFCR (kg h 1) 3.586 1.336 4.517 1.445 5.370 4.183
SEC (MJkW h 1) 11.07 19.55 10.15 15.39 10.35 11.61
hbth (%) 32.53 18.41 35.45 23.38 34.77 31
PGFR (m3 h 1) – 50 – 53 – 17
LFR (%) – 62.74 – 68.00 – 22.24
Sound pressure (db) 100.5 96.90 101.50 102.15 99.50 100.40

FD, fossil-diesel; LFR, liquid fuel replacement (%); PGFR, producer gas flow rate (m3 h 1
); LFCR, liquid fuel consumption rate (kg h 1
).
 B.B. Sahoo et al. / Renewable and Sustainable Energy Reviews 13 (2009) 1151–1184 1159

Fig. 16. Effect of fuels and engine loads on CO [30].

Fig. 14. Effect of fuels and engine loads on HC [30].

found to be extremely sensitive to the variation in the gas quality.

When the engine operates with a maximum liquid fuel replacement
the deterioration in the gas quality has an adverse effect on
performance of the engine. This is due to the fact that during such
operation the air intake is reduced to a level which is just enough for
the combustion of liquid fuel and producer gas. As the quality of
producer gas deteriorates with a consequent reduction in the engine
speed, the governor is trying to keep the speed up, while allowing
more liquid fuel in to the engine. This in turn causes incomplete
combustion and misfiring due to insufficient combustion air. Even,
the stalling of the engine will follow this if the air intake is not
increased.
At 84% engine load with 18.4:1 CR in dual-fuel mode, the
Fig. 17. Effect of fuels and engine loads on NO2 [30].
percentage reduction in concentration of pollutants like CO, CO2,
NO, NO2 is 55.5%, 19.7%, 82% and 83%, respectively, while HC
concentration increases by 67.2% (Figs. 14–18). The exhaust gas
temperature increases when CR and index of compression are
independently increased (Fig. 19).

Fig. 18. Effect of fuels and engine loads on NO [30].

5.2. Effect of engine speed

The emission and performance levels are measured on a chassis

Fig. 15. Effect of fuels and engine loads on CO2 [30]. dynamometer of a diesel engine under steady and unsteady
1160 B.B. Sahoo et al. / Renewable and Sustainable Energy Reviews 13 (2009) 1151–1184

Fig. 21. ‘bsfc’ and brake specific power as a function of engine speed for full load
Fig. 19. Effect of fuels and engine loads on engine exhaust temperature [30]. [12].

between 78 and 158 crank angles (depending on engine load and

speed) after the corresponding pressure peak for diesel fuelling.
The double hump showed on the pressure trace, is the combustion
of the pilot diesel charge followed by the NG combustion. The

Fig. 20. Brake power and brake torque as a function of engine speed for full load
[12].

conditions by Mansour et al. [12]. The compression ratio is kept

constant for the two versions engine test. Under unsteady
conditions, a small power and torque losses in dual-fuel version
is observed as compared to the diesel one, except around the speed Fig. 22. Fuel–air equivalence ratio versus engine speed for full load [12].
of 2400 rpm (Fig. 20). This difference is due to the system response
time of gaseous fuel injection in the admission collector. When the
engine speed rises to 2450 rpm, the brake torque and the brake
power in dual-fuel version becomes slightly higher than in the
diesel mode. Fig. 21 shows the evolution of the specific power and
fuel consumption versus engine speed for full load. It is noticed
that at low engine speed the difference in brake torque and power
is important between the two versions (loss of power and higher
consumption), but in high regimes the gap becomes less important.
In this analysis, the evaluation of the specific consumption is
based on the experimental results of the brake power and the
consumption results under steady conditions.
The NG fuelling compared to diesel fuelling leads to slightly
higher (less lean) equivalence ratios for a given speed condition
(Fig. 22). This occurs for two reasons; first, the NG is aspirated into
the engine where it mixes with air, thereby displacing some
portion of air, which could have moved into the cylinder. As less air
is induced, the equivalence ratio increases. Secondly, load is
decreased, the engine is less efficient using NG, and hence more NG
must be added to produce the fixed load-speed condition. The
increased fuelling then increases the equivalence ratio of the
engine. The cylinder pressure data measured with NG fuelling Fig. 23. Cylinder pressure data for diesel and dual-fuel gas fuelling for full loads at
show (Figs. 23 and 24) the second pressure peak occurring 1000 rpm engine speed [12].
 B.B. Sahoo et al. / Renewable and Sustainable Energy Reviews 13 (2009) 1151–1184 1161

are; pilot fuel injection timing 358 BTDC, mass of pilot fuel 0.37–
0.47 kg/h and compression ratio 22. From Fig. 26(a) and (b), it is
seen that the LPG produces the lowest torque output and thermal
efficiency as compared to methane or the NG mixture. The torque
output and efficiency is highest for methane gas. The thermal
efficiency also improves with increasing engine speed. Fig. 26(c)
and (d) shows that, the pressure rise rate (dP/du) decreases as the
engine speed increases for the three dual-fuel cases. However,
the pressure rise rate is highest for the dual-fuel engine and follows
the same trend with methane, followed by LPG and NG at almost all
engine speeds.

5.3. Effect of pilot fuel injection timing

The injection timing of the pilot fuel is an important factor that

influences the performance of dual-fuel engines. For a fixed total
equivalence ratio, advancing the injection timing increase the peak
cylinder pressure because more fuel is burned before TDC and the
peak pressure moves closer to TDC. Retarding the injection timing
decreases the peak cylinder pressure because more of the fuel
burns after TDC. This is because, the pilot fuel combustion is
Fig. 24. Cylinder pressure data for diesel and dual-fuel gas fuelling for full loads at delayed and thus, the temperature of the mixture is not enough to
2000 rpm engine speed [12]. propagate the flame in the whole gaseous fuel–air mixture; and
consequently, incomplete combustion of the gaseous fuel mixture
maximum combustion pressure for NG fuelling is slightly higher takes place. The charge temperature increases with advancing the
for all engine speeds than the diesel fuelling level. The general injection timing of the pilot fuel and the associated higher energy
trend is governed by decreasing pressure and temperature levels release rates of the mixture. Similarly, the rates of pressure rise
with increasing speeds. during the combustion of the gaseous fuel increases with
Selim [14] has examined the effect of engine speed on advancing the injection timing of the pilot fuel.
combustion noise for diesel and dual-fuel cases (Fig. 25). The The effect of advanced injection timing on the performance of
maximum pressure rise rate during combustion is presented as a NG used as primary fuel in dual-fuel combustion has been
measure of combustion noise. Generally, as the engine speed examined by Nwafor [2]. The injection is first advanced by 5.58 (i.e.
increases, the pressure rise rate (dP/du) decreases for both mode of 35.58 BTDC). The engine runs for about 5 min at this timing and
operation. In the diesel engine case, the pressure rise rate drops then stops and with subsequent attempts, he fails to start the
from about 5.5 bar/8C A at speed of 990 rpm to around 2.93 bar/ engine. But after changing the injection to 33.58 BTDC, the engine
8C A at 1890 rpm. However, for the dual-fuel case the pressure rise runs smoothly, but seems to incur penalty on fuel consumption
rate is higher than that for the diesel case, almost at all engine especially at high load levels. Table 5 compares fuel consumption
speeds, and it follows a similar trend to the diesel mode. It drops measurements for the two speeds, i.e. 3000 and 2400 rev/min of
from 6.6 bar/8C A at 980 rpm to 2.95 bar/8C A at 1880 rpm. It is also investigation. The data show that the advanced timing system
been shown in an IDI diesel engine [31] that the combustion noise incurs penalty on fuel consumption for a given operating
decreased when the engine speed increases. Further, Selim [4] has condition. The governor injects more pilot fuel than when running
presented the effects of engine speed for a dual-fuel engine on standard time units. The poor performance of gas engines at low
(Fig. 26a–d). During these experiments, the constant parameters load levels is due to the effect of gas residuals and low cylinder
temperature. It is also due to the reduction in combustion
efficiency caused by reduced flame propagation speed and
increased compression work resulting from the large amount of
air–gas inducted. The diesel fuel operation produces the highest
BTE at the two speeds (Fig. 27a and b). Standard timing shows little
improvement over the advanced system at 3000 rev/min. How-
ever, at 2400 rev/min the dual-fuel mode standard and advanced
timing show similar trends at low and intermediate load levels.
Fig. 28(a) and (b) shows that diesel fuel operation produces the
shortest delay periods at both test speeds. Standard dual-fuel
timing at 3000 rev/min shows longer delay periods at high loads
than the advanced injection timing operation. There is a very
significant difference between the ignition delays of diesel fuel and
dual-fuel operations at 2400 rev/min. Standard dual-fuel timing
also shows the longest delay period. At high loads and combustion
temperatures, the ignition delay is reduced and combustion is
dominated by the system temperature. At low loads with longer
delay periods, greater proportion of pilot fuel takes part in
premixed combustion, thus increases the tendency of diesel to
knock. Very poor atomization results in a relatively long delay
Fig. 25. Effect of engine speed on pressure rise rate for the diesel and dual-fuel period, due to the slow development of very fine droplets. Dual-
engines [14]. fuel operation with advanced timing shows the highest exhaust
1162 B.B. Sahoo et al. / Renewable and Sustainable Energy Reviews 13 (2009) 1151–1184

Fig. 26. Effects of engine speed on performance and noise: IT = 358 BTDC, CR = 22, md = (0.37–0.47)kg/h [4].

temperatures at 3000 rev/min (Fig. 29a and b). The dual-fuel mode lean combustion, wall wetting, cold starting and poor mixture
with the standard timing also shows a marginal increase over the preparation. For both test conditions, the HC levels are relatively
operation on pure diesel fuel. The standard timing dual-fuel unit high in dual-fuel operations and stay reasonably high throughout
produced the highest cylinder wall temperatures while the the load range.
advanced system showed the lowest values at this speed. However, Selim [14] has compared the effect of pilot diesel injection
the diesel fuel operation produces the lowest cylinder wall timing on the combustion noise (maximum pressure rise rate
temperatures, whereas the standard unit offers the highest values. during combustion) of a dual-fuel engine to 100% diesel case is
The best fuel economy is realised when running on pure diesel fuel shown in Fig. 31. For the late injection of pilot (20–258 BTDC), the
and hence, the thermal efficiency of the gas engine is less than that combustion noise is comparable for diesel and dual-fuel cases.
of pure diesel fuel case. The results from Fig. 30(a) and (b) indicate However, when the injection advance increases (25–408 BTDC), the
that the HC emissions of the gas-fuelled engine are higher than that dual-fuel engine produced a higher rate of pressure rise (dP/du).
in pure diesel fuel operation. Diesel fuel operation gives the lowest With the presence of gaseous fuel in the mixture, any advance in
HC emissions at both speeds. Dual-fuel standard timing shows pilot fuel injection results longer ignition delay period and increase
higher concentration of HC in the exhaust at low load levels over in pressure rise rate. Similar results are also shown by same author
the advanced injection unit. However, this system shows for LPG fuel [32].
decreased HC emission levels at high load level operation. HC At lower loads, dual-fuel engines suffer from lower thermal
emissions increase due to several factors, including quenching, efficiency and higher unburned percentages of fuel. In order to
 B.B. Sahoo et al. / Renewable and Sustainable Energy Reviews 13 (2009) 1151–1184 1163

Table 5
Effect of advanced injection timing on the performance of NG [2].

Engine load (N) Diesel fuel operation Standard timing, pilot Standard timing, gas Advanced timing, Advanced timing, gas
(kg/h) fuel consumed (kg/h) consumed (l/min) pilot fuel consumed (kg/min) consumed (l/min)

Engine speed = 3000 rev/min

6.61 0.539 0.368 10.0 0.408 10.0
13.18 0.642 0.402 11.0 0.456 11.0
17.72 0.696 0.442 12.4 0.478 12.5
23.78 0.818 0.447 14.0 0.496 14.0
29.33 0.894 0.470 16.0 0.509 16.0
33.87 0.975 0.493 17.0 0.535 17.0
38.92 1.045 0.529 18.0 0.594 18.0
44.47 1.154 0.563 19.0 0.642 19.0
49.52 1.266 0.570 20.2 0.631 20.2

Engine speed = 2400 rev/min

8.63 0.463 0.333 8.0 0.349 8.0
13.68 0.509 0.299 12.0 0.319 12.0
18.73 0.592 0.278 14.0 0.298 14.0
24.79 0.680 0.279 15.2 0.295 15.2
29.83 0.735 0.303 16.0 0.324 16.0
35.89 0.829 0.312 17.0 0.366 17.0
41.44 0.887 0.328 17.5 0.383 17.5
46.80 1.013 0.344 18.0 0.456 18.0
53.56 1.132 – – – –

Fig. 27. Injection advanced effect on gas combustion—BMEP (kN/m2) versus BTE (%) [2].

improve this, the effects of injection timings of 258, 27.58 and 308 longer ignition delay of the mixture with the increased timing
BTDC on the performance of an IDI diesel engine are investigated advance. The longer ignition delay allows a fuller spray penetration
by Abd et al. [33]. Figs. 32 and 33 show that, advancing the pilot and development, creating a larger amount of the pilot fuel–air–
fuel injection timing reduces the UBHC emissions. This is due to a gaseous fuel mixture (or flame propagation region) prior to

Fig. 28. Injection advanced effect on gas combustion—BMEP (kN/m2) versus ignition delay (deg.) [2].
1164 B.B. Sahoo et al. / Renewable and Sustainable Energy Reviews 13 (2009) 1151–1184

Fig. 29. Injection advanced effect on gas combustion—BMEP (kN/m2) versus temperature (8C) [2].

Fig. 30. Injection advanced effect on gas combustion—BMEP (kN/m2) versus UBHC (ppm) [2].

ignition. The higher combustion rates of this larger premixed For a fixed total equivalence ratio, amount of pilot fuel and intake
regions yields higher combustion temperatures and thus, lowers temperature, advancing the injection timing has a great effect on
the UBHC emissions. With bigger injection advance, better overall the maximum charge temperature in the cylinder. For any total
combustion and the activity of the partial oxidation reactions equivalence ratio, as the injection is retarded, the maximum charge
reduce the CO emissions (Figs. 34 and 35). It also widens the lower temperature decreases. The net effect is a reduction in NOx as
combustion limit boundary of the overall lean mixture effectively. shown in Figs. 36 and 37. Higher combustion temperatures due to

Fig. 31. Effect of pilot fuel injection timing on pressure rise rate for the diesel and Fig. 32. Variations of experimental results of UBHC concentration with total
dual-fuel engine [14]. equivalence ratio for different values of injection timings for methane [33].
 B.B. Sahoo et al. / Renewable and Sustainable Energy Reviews 13 (2009) 1151–1184 1165

Fig. 33. Variations of experimental results of UBHC concentration with total

equivalence ratio for different values of injection timings for propane [33].
Fig. 36. Variations of experimental results of NOx with total equivalence ratio for
different values of injection timings for methane [33].

Fig. 37. Variations of experimental results of NOx with total equivalence ratio for
Fig. 34. Variations of experimental results of CO concentration with total different values of injection timings for propane [33].
equivalence ratio for different values of injection timings for methane [33].

Fig. 35. Variations of experimental results of CO concentration with total Fig. 38. Variations of experimental results of thermal efficiency with total
equivalence ratio for different values of injection timings for propane [33]. equivalence ratio for different values of injection timings for methane [33].
1166 B.B. Sahoo et al. / Renewable and Sustainable Energy Reviews 13 (2009) 1151–1184

Fig. 39. Variations of experimental results of thermal efficiency with total

equivalence ratio for different values of injection timings for propane [33].
Fig. 40. Effect of injection timing on knocking torque for a dual-fuel engine fuelled
with methane (CH4) and propane (C3H8) [33].

Fig. 41. Effects of pilot fuel injection timing on performance and noise: N = 1300 rpm, CR = 22, md = 0.37 kg/h [4].
 B.B. Sahoo et al. / Renewable and Sustainable Energy Reviews 13 (2009) 1151–1184 1167

Fig. 44. Onset of combustion and combustion duration (10–90%) versus injection
timing [35].

Fig. 42. Fuel conversion efficiency and BSEC versus injection timing [35].
would result in a higher pressure rise rate (dP/du). This is also
shown in by Abdalla et al. [34] for 100% diesel and all diesel–
advancing the pilot fuel injection timing lead to increases in the methanol blends.
brake power and, consequently, thermal efficiency (Figs. 38 and To achieve low NOx and good fuel efficiency from pilot-ignited
39), respectively. Further advances in the injection timing will natural gas combustion, Krishnan et al. [35] have tried different
increase the tendency to knocking early for medium and high pilot injection timings in a single cylinder CI engine. The overall
engine loads (Fig. 40). set of experiments involved engine testing at a constant speed of
Selim [4] has investigated the effects of pilot fuel injection 1700 rev/min, full load (42 kW, 1220 kPa bmep) engine operation
timing for a dual-fuel diesel engine. The results are shown at at fixed pilot quantity and inlet conditions. Fuel conversion
constant speed of 1300 rpm and compression ratio of 22 (Fig. 41a– efficiency (defined by the ratio of the brake power to the total rate
d). The highest torque output for methane and natural gas occurs of energy input into the engine) is calculated from the measured
when the injection timing is 258 BTDC, while for LPG it occurs at total diesel and natural gas flow rates multiplied by their
308 BTDC. The torque output and hence the thermal efficiency is respective lower heating values. Brake specific energy consump-
highest at certain timings, and it decreases at earlier or later timing tion (BSEC) is then calculated which is the reciprocal of the fuel
(Fig. 41a and b). Earlier injection of pilot fuel causes the maximum conversion efficiency. As injection timing is advanced from 158 to
pressure to increase, Fig. 41(c), and occurs BTDC in the compres- 458 BTDC, fuel conversion efficiency increased from 38% to about
sion stroke. This in turn, reduces the maximum pressure during the 43% (Fig. 42). Upon further advance of the injection timing, the
expansion stroke and torque output. The combustion noise, as efficiency decreases to about 40% at 608 BTDC. When brake power
shown in Fig. 41(d), increases as the pilot diesel injection advance and engine speed are held constant for all injection timings, the
increases for all dual-fuel cases. This is attributed to the increase in particular heat release pattern that provides for greater work per
ignition delay of the diesel fuel, since the liquid fuel injected earlier unit mass of fuel translates into higher fuel conversion efficiency.
in lower air pressure and temperature. The longer delay period As injection timing is retarded from 358 to 158 BTDC, the start of

Fig. 43. (a) Gross heat release rate versus crank angle for injection timings between 158 and 358 TDC. (b) Gross heat release rate versus crank angle for injection timings
between 408 and 608 BTDC [35].
1168 B.B. Sahoo et al. / Renewable and Sustainable Energy Reviews 13 (2009) 1151–1184

Fig. 45. NOx-fuel conversion efficiency trade-off curve at different injection timings
Fig. 46. HC and CO emissions versus injection timing [35].
[35].

heat release is delayed and the overall heat release process is compared to 608 BTDC. NOx emissions for the maximum efficiency
shifted away from TDC. Also the duration and peak heat release timing (458 BTDC) are much higher than those for very retarded
are increased (Fig. 43a). These effects contributed to the loss in (158 BTDC) or very advanced (608 BTDC) timings (Fig. 45). Along
fuel conversion efficiency between 358 and 158 BTDC. As injection with NOx reduction, efficiency also decreases when injection
timing is advanced from 408 to 608 BTDC, the duration of heat advances or retards from 458 BTDC. However, it is interesting that
release does not change significantly even though the start of heat the efficiency for 608 BTDC is about 3% greater than that for 158
release is delayed (Fig. 43b). The delayed start of heat release has BTDC and NOx emissions are lower as well. Thus, it is clearly
led to a decrease in fuel conversion efficiency for timings more beneficial to advance the pilot injection timing to
advanced beyond 458 BTDC. Fig. 44 shows the onset of combustion reduce NOx emissions and maintain minimal loss in fuel
in degrees ATDC and duration of combustion for the results shown conversion efficiency. Fig. 46 shows that for timings retarded
in Fig. 43(a) and (b). The duration of combustion shown here is the beyond 258 BTDC, both HC and CO emissions are found to decrease
period in crank angles from 10 to 90% mass burn. The onset of especially between 208 and 158 BTDC timings. From Fig. 43(a) and
combustion is seen to be earlier for injection timings between 308 44, it is seen that for both 208 and 158 BTDC timings, the heat
and 458 BTDC. For timings advanced beyond 458 BTDC or retarded release peaks are lower and the combustion durations are longer
beyond 308 BTDC, combustion begins progressively closer to TDC. compared to 258 BTDC. Although slower combustion have
However, combustion duration does not follow the same trend. resulted in more wall-quenched HC for 208 and 158 BTDC, the
For instance, although combustion has started at about the same time available for oxidation also increased since combustion
time for both 158 and 608 BTDC, the combustion duration for the duration increased. For the longest combustion duration (158
former is seen to be much longer than that of the latter. So, BTDC timing), the time for which high temperatures persist in the
prolonged combustion time for 158 BTDC progresses into the cylinder is increased, and hence more complete oxidation of HC
expansion stroke, thereby reduces the fuel conversion efficiency and CO have occurred.

Fig. 47. The variations of the exhaust gas concentrations of methane and CO with total equivalence ratio for different pilot fuel quantities at ambient intake conditions and
1000 rpm [9].
 B.B. Sahoo et al. / Renewable and Sustainable Energy Reviews 13 (2009) 1151–1184 1169

Fig. 48. Variations of the experimentally established flame spread limits (FSL) with
the pilot quantity employed for methane operation at a compression ratio of 14.2:1
and 1000 rpm. The corresponding flame initiation limit values are also shown [9].

5.4. Effect of pilot fuel mass inducted

Fig. 51. Variations of experimental results of UBHC concentration with total
equivalence ratio for different values of pilot fuel quantities (main fuel: propane) [36].
The pilot fuel quantity is one of the most important variables
that have a controlling influence on the performance of dual-fuel
engines, especially at light loads. It is known that most diesel fuel

Fig. 52. Variations of experimental results of CO concentration with total

Fig. 49. Variations of experimental results of UBHC concentration with total equivalence ratio for different values of pilot fuel quantities (main fuel:
equivalence ratio for different values of pilot fuel quantities (main fuel: methane) propane) [36].
[36].

Fig. 50. Variations of experimental results of CO concentration with total equivalence Fig. 53. Variations of experimental results of NOx with total equivalence ratio for
ratio for different values of pilot fuel quantities (main fuel: methane) [36]. different values of pilot fuel quantities (main fuel: methane) [36].
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Fig. 54. Variations of experimental results of NOx with total equivalence ratio for Fig. 57. Effect of pilot fuel flow rate on the knocking torque [36].
different values of pilot fuel quantities (main fuel: propane) [36].

Fig. 58. Variations of thermal efficiency with total equivalence ratio with different
Fig. 55. Variations of experimental results of brake power with total equivalence
values of pilot fuel quantities [36].
ratio for different values of pilot fuel quantities (main fuel: methane) [36].

Fig. 56. Variations of experimental results of brake power with total equivalence
ratio for different values of pilot fuel quantities (main fuel: propane) [36]. Fig. 59. Effect of pilot fuel mass on pressure rise rate for the dual-fuel engine [14].
 B.B. Sahoo et al. / Renewable and Sustainable Energy Reviews 13 (2009) 1151–1184 1171

Fig. 60. Pressure–crank diagram for dual-fuel engine, at different pilot fuel mass
[14]. Fig. 62. Pressure–crank angle diagram of diesel fuel operation [37].

verted methane become virtually unaffected by the pilot quantity.

This is indicative of the equivalence ratio limit for successful flame
propagation from the pilot ignition centres. The variations of the
flame spread limits (FSL) with changes in pilot quantity are shown
in Fig. 48. The increase in pilot quantity lowers the FSL due to
number of contributing factors. These include; greater energy
release on ignition, correspondingly improved pilot fuel injection
characteristics, larger pilot-mixture envelope size, larger ignition
centres, higher rate of heat transfer to the unburned gaseous
mixture and increased contribution of hot residual gases. The
flame initiation limit exhibits a similar trend is also shown in
Fig. 48.
Abd et al. [36] have investigated the effect of three pilot fuel
quantities, 0.15, 0.20 and 0.25 kg/h, on the performance and
emissions of an indirect injection diesel engine fuelled with
gaseous fuel. Fig. 49 shows that at very light loads with a small
pilot fuel quantity, the concentrations of UBHCs measured are
relatively high. This is because, in an excessively lean mixture, the
flame originating from the pilot ignition (if there is any) cannot
propagate throughout the whole combustion chamber; only
Fig. 61. Dual-fuel pilot injection pressure–crank angle diagram [37]. partial oxidation occurs, and thus UBHC emissions as well as CO
emissions are relatively higher, as shown in Fig. 50. At higher loads,
injection systems experience poor atomization and combustion when the gaseous fuel concentration in the air charge is above the
when the amount of fuel injected per cycle is reduced below 5–10% lean combustion limit, the flame is able to propagate through most
of the maximum design level. of the combustion chamber unaided, and varying the pilot fuel
The variations of the exhaust gas concentrations of methane quantity has little effect. The change in oxidation reactions from
and carbon monoxide with total equivalence ratio for different unsuccessful to successful flame propagation reduces the HCs and
pilot fuel quantities is investigated by Badr et al. [9]. Fig. 47(a) and CO emissions slightly as shown in Figs. 51 and 52.
(b) shows that there is a limiting equivalence ratio beyond which In dual-fuel engines, the effective size of the combustion zone,
the exhaust emissions of the carbon monoxide and the uncon- which relates to the size of the pilot fuel zone, is another important

Table 6
Effect of pilot fuel/gas ratio on knock characteristics of dual-fuel engine.

Load (N) Pilot fuel (kg/h) Gas supply (kg/h) Total fuel (kg/h) % pilot fuel % gas supply Mixture strength

6.61 0.305 0.626 0.931 32.76 67.24 0.557

13.18 0.292 0.720 1.012 28.85 71.15 0.620
17.72 0.271 0.792 1.063 25.49 74.51 0.667
23.78 0.295 0.846 1.141 25.85 74.15 0.730
28.82 0.313 0.882 1.195 26.19 73.81 0.779
33.87 0.331 0.907 1.238 26.74 73.26 0.823
38.92 0.343 0.936 1.279 26.82 73.18 0.866
44.47 0.307 1.008 1.315 23.35 76.65 0.922

Engine speed = 3000 rev/min [37].

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factor, besides to the mechanism of oxidation of nitrogen that 8.1 bar/8C A at 0.38 kg/h pilot mass to about 5.12 bar/8C A at
determines the quantity of nitrogen oxides produced. For the same 0.52 kg/h. Pressure rise rate then increases to 7.42 bar/8C A as the
total equivalence ratio, increasing the pilot fuel quantity increases pilot fuel mass increases to 0.63 kg/h. This in turn, increases in
the charge temperature which tends to increase the production of flame volume resulting smooth burn of methane gas smoothly and
NOx, as shown in Fig. 53. For relatively high loads, the combustion at a lower rate of combustion. The maximum pressure (Fig. 60) and
of gaseous fuel is more complete and less affected by the pilot fuel the maximum rate of pressure rise (dP/du) are lower for a pilot fuel
quantity, and thus, has a mild effect. Similarly, for a given pilot fuel mass of 0.52 kg/h, and it is increased for a lower or higher amount
quantity when higher gaseous fuel concentrations in the cylinder of diesel pilot fuel.
charge are employed, the significant increases in the size of the The effect of pilot fuel/gas ratio on knock characteristics of dual-
combustion zone lead to correspondingly increased higher fuel engine is examined by Nwafor [37]. The knock characteristics
production of NOx. Fig. 54 shows that the production of NOx is for pure diesel and dual fuel operations are compared in Figs. 61
influenced markedly by both the quantity of the pilot fuel and 62. Similar figures are also presented by the author for natural
employed and the overall equivalence ratio. The use of large pilot gas in and unmodified CI engine [38]. The ripples in Fig. 61 are the
fuel quantities and high charge equivalence ratios results in a indication of combustion knock. Dual-fuel operation shows longer
significant increase in the production of NOx. Figs. 55 and 56 ignition delay as measured by the author. The degree of knock in
indicate that the employment of a large pilot fuel quantity produce this phase depends on the ratio of the alternative fuel (NG) to the
higher power output. The increase of pilot fuel quantity leads to pilot fuel as given in Table 6 and thus on the load and speed of
successful flame propagation and, consequently, increases the operation. The increase in speed increases the ignition delay when
output power. Fig. 57 indicates that using a greater pilot fuel running on pure diesel fuel, hence the quantity of premixed pilot
quantity, to enhance the combustion process at low loads, leads to fuel that takes part in combustion increases. Increasing the pilot
increase the tendency to knock at high loads. The thermal fuel and reducing primary fuel reduces the knocking phenomena in
efficiency for methane is seen to be greater than that for propane dual-fuel engines.
at low loads (Fig. 58). This is due to the higher ignition delay of Selim [4] has examined the effects of pilot fuel quantity of a dual
propane at low loads. fuel engine are shown in Fig. 63(a)–(d). During these experiments
The experimental investigation conducted by Selim [14] shows the following parameters are kept constant: engine speed
that increasing the pilot diesel fuel mass can resulted in increase in 1300 rpm, pilot fuel injection timing 358 BTDC and compression
the engine torque (Fig. 59). This is postulated to the increase in the ratio 22. Increasing the quantity of pilot diesel fuel increases the
heat released from burning more fuel. The combustion noise (dP/ torque output, Fig. 63(a) and (b), and, hence, thermal efficiency for
du), however, dropped with the increase of pilot mass from around the three gaseous fuels used. Increasing the pilot diesel fuel for the

Fig. 63. Effects of pilot fuel mass on performance and noise: N = 1300 rpm, IT = 358 BTDC, CR = 22 [4].
 B.B. Sahoo et al. / Renewable and Sustainable Energy Reviews 13 (2009) 1151–1184 1173

three cases results in greater energy release on ignition, improved (from 8.1 to 17.6 to 20 N-m) and also extended the ignition limits
pilot injection characteristics, larger size of pilot mixture envelope greatly (from 7.85 to 17 to 18.5 N-m) as shown in Fig. 64a–c. This is
with greater entrainment of the gaseous fuel, a larger number of due to the early knocking at high compression ratios associated
ignition centres requiring shorter flame travels and a higher rate of with higher pressures and temperatures and lower self-ignition
heat transfer to the unburned gaseous fuel–air mixture [9]. These temperatures of LPG. For extended ignition limits and knock free
factors tend to increase the power output and thermal efficiency of operation of the dual-fuel engine, the compression ratio has to be
the dual-fuel engine [36]. Increasing the pilot fuel mass also reduced to lower values. For the NG mixture and methane,
resulted in higher maximum combustion pressure as may be seen Fig. 64(b)–(c), have similar trend to LPG with the only difference at
in Fig. 63c. For the dual-fuel engine, the maximum pressure is the compression ratio of 22. As LPG has the lowest self-ignition
always higher than in the diesel fuel case due to the combustion temperature (about 400 8C), it starts knocking and ignition fails at
and extra heat released from gaseous fuels. The maximum pressure lower engine torque compared to the other two gases namely,
rise rate, as seen in Fig. 63d, is generally reduced when the pilot methane (about 650 8C) and CNG (about 500 8C).
fuel quantity is increased. The decrease in the combustion noise The maximum pressure rise rate at different compression ratios
(dP/du) when the pilot fuel mass is first increased is postulated to for the three fuels is illustrated in Fig. 65(a–c). It is seen that
be due to the increase in flame volume resulting from the increase increasing the compression ratio generally increases the combus-
in pilot fuel mass, which burns the gaseous fuel smoothly and at a tion noise due to the higher self-ignition possibility of the gaseous
lower rate of combustion. However, when the pilot fuel mass fuels at higher pressures and temperatures. As the compression
increases beyond a certain amount, the ignition delay period of the ratio is reduced, the combustion noise is also reduced, and the
pilot diesel increases and hence, the pressure rise rate (dP/du) for ignition limits are extended.
the gas–air mixture increases [14].
5.6. Effect of intake manifold conditions
5.5. Effect of engine compression ratio
In order to improve exhaust emissions at part load, Kusaka et al.
Selim [4] has examined the effects of compression ratio of a [39] have modified the intake charge condition including intake
dual-fuel engine: First, on knock onset and ignition failure temperature and exhaust gas recirculation (EGR) as given in
(Fig. 64a–c) and secondly, on the maximum pressure rise rate Table 7. They have installed a heat exchanger in the intake system
(Fig. 65a–c). For LPG, reduction in the compression ratio results in and a Pt-catalyst is used in the exhaust system to reduce unburned
retarding the occurrence of knock onset in the dual-fuel engine natural gas emission.

Fig. 64. Effect of compression ratio on knock and ignition limits for LPG, CH4 and NG: N = 1300 rpm, IT = 358 BTDC, md = 0.37 kg/h [4].
1174 B.B. Sahoo et al. / Renewable and Sustainable Energy Reviews 13 (2009) 1151–1184

Fig. 65. Effect of compression ratio on combustion noise for LPG, CH4 and NG: N = 1300 rpm, IT = 358 BTDC, md = 0.37 kg/h [4].

The base line condition is taken here as; intake temperature of Fig. 67 shows the exhaust emissions characteristics under
20 8C and EGR 0%. In the case of intake heating of 165 8C, the start of various conditions. In this figure, ‘‘Diesel’’ means diesel operation
RHR becomes earlier than that of other conditions due to higher without dual-fuelling, while ‘‘Base’’ represents baseline condition.
cylinder temperature (Fig. 66). On the other hand, the combination As shown in this figure, EGR with intake heating can favourably
of 50% EGR and intake heating of 210 8C at the start of RHR becomes reduce THC emissions compared to the baseline and therefore,
retarded compared to that of intake heating of 165 8C without EGR. thermal efficiency is improved. NOx is drastically increased in the
Ignition reaction proceeds faster by intake heating, while the inert case of intake heating without EGR. However, when EGR is
gases included in re-circulated exhaust gas impede progress of combined with intake heating, NOx is reduced drastically. In the
ignition reactions. As a result of the competition of these above two dual-fuel system, the engine can operate with high EGR ratio
conflicting effects, the ignition timing is determined as shown in because dual-fuel diesel engine produces little soot, leading to
Fig. 66. It is also seen from this figure that EGR can favourably reduced NOx emission. This is because of the fact that the inert gas,
control the rate of pressure rise. The reason why the shape of RHR which has a large heat capacity, lowers combustion temperature,
at intake heating of 165 and 210 8C has two peaks is that the while EGR reduces the oxygen concentration in the cylinder. From
natural gas combustion takes place after the combustion of pilot these above points of view, EGR combined with intake heating
diesel is activated, due to the increased natural gas mixture reduces NOx and THC emissions without deteriorating engine
temperature. thermal efficiency.

5.7. Effect of type of gaseous fuel

Table 7
Engine tests conditions for intake heating combined with EGR (engine speed: Laboratory investigation is carried out by Bari [40] to see the
1280 rpm, load: 1/5, NG fraction: 80%) [39]. effect of carbon dioxide (CO2) on the performance of biogas-diesel
EGR ratio (%) Diesel injection Intake temperature (8C)
duel-fuel engine. Mixing of NG with CO2 at different level of
timing (degree BTDC) compositions is the main part of the experiment. At certain
Before the After the heat
percentage of diesel substitution by NG (15, 30, 50 and 75%), the
heat exchanger exchanger
pure CO2 is introduced with the NG. The speed and power are
30 13 60 170 maintained at constant level by varying the quantity of diesel flow.
50 13 110 210
When the percentage of CO2 in biogas increases, the biogas supply
60 13 145 230
to produce the same power also increases (Fig. 68). The rate of
 B.B. Sahoo et al. / Renewable and Sustainable Energy Reviews 13 (2009) 1151–1184 1175

particles. Under the above conditions, the engine performance is

found comparatively better with lower ‘bsfc’s and diesel flow rates.
Therefore, the trend of the ‘bsfc’s curve is falling up to 20–30% CO2
in the gaseous mixture. When the percentage of CO2 becomes
higher in the gaseous mixture, CO2 remains undissociated. This
undissociated CO2 acts as inert gas. Addition of such inert gas
affects the burning velocity of gas–air mixture [42]. Therefore, it
results incomplete combustion and increases ‘bsfc’. Biogas
containing more than 40% CO2 needs scrubbing, because it is
found in this research that the engine ran harshly with biogas
containing high CO2 (>40%).
Henham and Makkar [3] have conducted tests on the dual-fuel
diesel engine at various proportions of gas mixture comprising of
NG and CO2. First, gasoil is substituted by NG from the British
mains (94% methane, 34.8 MJ/m3) at four constant levels (22%, 37%,
45% and 58%). Then, taking each constant level of NG as 100%, it is
mixed with CO2 to vary the composition of gas mixture. Test
results in Figs. 71–73 are at engine speed 2000 rev/min and torque
40 N-m using NG:CO2 mixture for a range of 100:0–40:60 at four
constant NG substitution levels. The overall efficiency falls with NG
substitution at all constant levels (Fig. 71). On mixing NG with CO2
Fig. 66. The effect of intake heating and EGR on (a) R.H.R and (b) rate of pressure rise efficiency is not much affected up to 37% NG substitution. With
(engine speed: 1280 rpm, load: 1/5, NG fraction: 80%) [39]. higher NG substitution, efficiency decreases with increasing CO2 in
gas mixture. At 58% NG substitution level, efficiency decreases
from 28.2% to 26.2% with increasing CO2 in gas mixture. With
higher gas substitution a greater proportion of air is replaced by
increase of biogas at higher substitution is higher. Figs. 69 and 70 gas. So, volumetric efficiency of the engine is lowered and it results
show the variation of diesel flow and ‘bsfc’ with variation of CO2 in in less power. Fig. 72 indicates that exhaust temperature is affected
the biogas. The CO2 in the mixture dissociate into CO and oxygen, more by NG substitution up to 45%. At 58% NG substitution,
because the flame temperature of diesel is very high to initiate exhaust temperature increases with increasing CO2 in gas mixture,
dissociation. It is evident that CO is comparatively fast burning gas from 382 to 402 8C. The CO is affected mainly by NG substitution
than other alternative fuels [41]. Thus, the burning rate of the total and not so much by the proportion of CO2 in the gas mixture
gas–air mixture accelerates for the presence of CO. The amount of (Fig. 73). The increase in CO as compared to that with gasoil only
oxygen found by dissociation of CO2 increases the concentration of (when it is only 0.04%) is caused by lower effective air–fuel ratio as
oxygen in the gas air mixture. Therefore, it reduces the ignition gas mixture replaces more air. Test results in Figs. 74–76 are at
delay as well as enhances the combustion of unburned carbon engine speed 2800 rev/min and torque 40 N-m using NG:CO2

Fig. 67. The effect of intake heating and EGR on emissions and thermal efficiency (engine speed: 1280 rpm, load: 1/5, NG fraction: 80%) [39].
1176 B.B. Sahoo et al. / Renewable and Sustainable Energy Reviews 13 (2009) 1151–1184

Fig. 71. IDI dual-fuel efficiency variation using gas mixture at 2000 rev/min and
Fig. 68. Variation of biogas flow with the increase of CO2 in biogas [40].
40 N-m [3].

Fig. 72. IDI dual-fuel exhaust temperature variation using gas mixture at 2000 rev/
min and 40 N-m [3].

Fig. 69. Variation of diesel flow with the increase of CO2 in biogas [40].

Fig. 73. IDI dual-fuel CO variation using gas mixture at 2000 rev/min and 40 N-m
[3].

Fig. 70. Variation of ‘bsfc’ with the increase of CO2 in biogas [40].

mixture for a range of 100:0–30:70 at five constant NG substitution this condition, the combustion is less controlled and knock is
levels. Fig. 74 indicates that the overall efficiency decreases with noticed during the test run. The in-cylinder pressure character-
increase in CO2 in gas mixture at all substitution level where as the istics of the test engine for gasoil only, gasoil and 58% NG
exhaust temperature and CO follow the same patterns as at substitution, and gasoil and gas mixture (NG:CO2, 1:1), respec-
2000 rev/min except at 65% NG substitution (Figs. 75 and 76). In tively are presented in Figs. 77–82. At 2000 rev/min, peak pressure
 B.B. Sahoo et al. / Renewable and Sustainable Energy Reviews 13 (2009) 1151–1184 1177

Fig. 74. IDI dual-fuel efficiency variation using gas mixture at 2800 rev/min and
Fig. 77. P–u diagram for gasoil only @ 2000 rev/min, 40 N-m [3].
40 N-m [3].

Fig. 78. P–u diagram for gasoil and 60% NG substitution @ 2000 rev/min, 40 N-m [3].
Fig. 75. IDI dual-fuel exhaust temperature variation using gas mixture at 2800 rpm
and 40 N-m [3].
Papagiannakis and Hountalas [24] have examined the dual-
rises from 70 bar to 83 bar at 60% NG substitution and falls to fuel engine performance and exhaust emission characteristics by
77 bar for gas mixture of NG:CO2 (1:1). Sharper peaks are observed varying the mass ratio of gaseous fuel (NG). The engine is
in Figs. 78 and 79 compared to Fig. 77 are due to be the result more supplied with NG from the local low-pressure distribution
fuel availability at the initiation of combustion. For 2800 rev/min, network after making the appropriate modifications. Measure-
peak pressure rises from 57 bar to 70 bar at 60% NG substitution ments are taken at three different engine loads corresponding to
and falls to 67 bar for gas mixture of NG:CO2 (1:1) indicated in 40%, 60% and 80% of full load and three engine speeds of 1500,
Figs. 80–82.

Fig. 76. IDI dual-fuel CO variation using gas mixture at 2800 rev/min and 40 N-m
[3]. Fig. 79. P–u diagram for gasoil and NG:CO2 (1:1) @ 2000 rev/min, 40 N-m [3].
1178 B.B. Sahoo et al. / Renewable and Sustainable Energy Reviews 13 (2009) 1151–1184

Fig. 80. P–u diagram for gasoil only @ 2800 rev/min, 40 N-m [3].

Fig. 83. Experimental cylinder pressure and total heat release traces under normal
diesel and dual-fuel operation for 2000 rpm engine speed at 40% of engine load [24].

dual-fuel operation strongly diverge from the respective values

under normal diesel operation as NG mass ratio increases. The
second peak of heat release observed in this figure is due to
combustion of NG. But, it does not affect strongly the pressure
curve since it is in the expansion stroke and causes only a slight
increase of its value. For all cases examined the cylinder pressure
under dual-fuel operation decreases when increasing the amount
of NG. This is observed during compression and the initial stages
of combustion and is mainly the result of the higher specific heat
capacity of the NG–air mixture and its slower combustion rate
compared to that of diesel fuel.
Referring to Fig. 85, maximum combustion pressure is strongly
affected by the presence of gaseous fuel. As the quantity of
Fig. 81. P–u diagram for gasoil and 60% NG substitution @ 2800 rev/min, 40 N-m [3].
gaseous fuel increases, keeping engine load constant, peak
cylinder pressure decreases significantly while the slope remains
2000 and 2500 rpm under both, normal diesel operation (100% almost the same regardless of engine load. Ignition delay
diesel fuel) and dual-fuel operation (NG and diesel fuel). At part increases with increase in the percentage mass of NG (Fig. 86).
load condition, increasing the NG mass ratio over 50%, the The increase is due to the reduction of charge temperature close to
cylinder pressure after ignition increases at a slower rate the point of fuel injection and high overall specific heat capacity
compared to normal diesel operation (Fig. 83). The initial heat [11,27]. Combustion duration under dual-fuel operation is
release rate for mass ratios up to 50% is almost the same while for
higher values it is lower, revealing less premixed combustion.
The effect of NG combustion on the total heat release becomes
evident only at high mass ratios around 80%. At high engine load
(80%) as shown in Fig. 84, the cylinder pressure traces under

Fig. 84. Experimental cylinder pressure and total heat release traces under normal
Fig. 82. P–u diagram for gasoil and NG:CO2 (1:1) @ 2800 rev/min, 40 N-m [3]. diesel and dual-fuel operation for 2000 rpm engine speed at 80% of engine load [24].
 B.B. Sahoo et al. / Renewable and Sustainable Energy Reviews 13 (2009) 1151–1184 1179

Fig. 85. Maximum combustion pressure as function of NG mass ratio at 2000 rpm
for various engine loads [24].
Fig. 87. Variation of ‘bsfc’ as function of NG mass ratio at 2000 rpm for various
engine loads [24].
generally longer compared to the one under normal diesel
operation for all cases examined (Fig. 86). Under dual-fuel
operation, there is an increase of combustion duration with that the LHV of NG is higher compared to the one of diesel fuel
increasing NG mass ratio at a rate of which is slightly more intense used, its total ‘bsfc’ is even higher if these are corrected to the LHV
at low mass ratios of NG. At part load under dual-fuel operation, of the diesel fuel. This reveals a poor utilization of the gaseous fuel
the total ‘bsfc’ is considerably higher compared to the one under due mainly to the lower temperature inside the combustion
normal diesel operation, mainly as a result of the low combustion chamber of the engine and the late start of ignition because of the
rate of gaseous fuel (Fig. 87). A similar trend is observed at high higher ignition delay [43].
load under dual-fuel operation, but in this case the slope of ‘bsfc’ At part load as the NG mass ratio increases, soot concentration
increment with NG mass ratio is lower compared to the one at part decreases sharply since less liquid fuel is injected on a percentage
load. Here, the total ‘bsfc’ is estimated from the brake power basis and thus less soot is formed (Fig. 88). With high engine load
output of the engine and the measured mass flow rate of fuels. and low gaseous fuel mass ratios, the charge temperature is
Thus, no correction is made to consider for the difference in lower
heating values (LHV) of NG and diesel fuel. Considering the fact

Fig. 86. Variation of ignition delay and combustion duration as function of NG mass Fig. 88. Soot opacity and nitric oxide as function of NG mass ratio at 2000 rpm for
ratio at 2000 rpm for various engine loads [24]. various engine loads [20].
1180 B.B. Sahoo et al. / Renewable and Sustainable Energy Reviews 13 (2009) 1151–1184

inferior and it results a slight increase in soot emissions compared

to the one under normal diesel operation. But, at higher mass
ratios soot decreases and this reduction compared to normal
diesel operation is sharper. This is due to the higher gas
temperature that promotes soot oxidation rate contributing to
a further decrease of soot concentration. Taking these into
consideration for dual-fuel operation, use of a high percentage
of NG is an efficient way to reduce soot concentration. Practically,
the gaseous fuel produces no soot while it contributes to the
oxidation of the one formed from the combustion of the liquid
fuel. It is widely recognised that the high oxygen concentrations
and high charge temperatures favour the formation of nitric oxide
(NO) [22,23,44,45]. The increase of NG mass ratio results to a
decrease of NO concentration compared to the one under normal
diesel operation (Fig. 88). At high engine load and low mass ratios
of NG there is a sharper decrease of NO concentration, compared
to part load conditions with increasing gaseous fuel percentage.
The reduction of NO concentration is due to several factors; the
less intense premixed combustion, the reduction of gas tempera-
ture due to increase of the specific heat capacity, the slower
combustion and finally, the reduction of oxygen concentration
due to presence of the NG mass ratio which replaces an equal
amount of air in the cylinder charge. The variation of UBHC in the
exhaust gases is consistent with the quality of the combustion
process of the engine [11,25,27,45]. At low load, HC emissions
increase with increase of gaseous fuel percentage (Fig. 89). The
low charge temperature combined with a slower burning rate
Fig. 89. UBHC and CO as function of NG mass ratio at 2000 rpm for various engine leads to an increase of gaseous fuel, which does not burn
loads [24]. completely. At high engine load there is an increase of HC

Table 8
Comparison of results by researcher(s) on the performance of dual-fuel diesel engines.

Researcher(s) Performance of dual-fuel diesel engine

Effect of engine load

Selim [14] Combustion noise increases with increase in load and is always higher than that of for the diesel fuel case
Uma et al. [16] The engine performance decreases at part load conditions both at diesel and dual-fuel mode. CO emissions are higher than diesel
engines at all operated load conditions. HC emissions are little higher than diesel case. But, dual-fuel operation reduces
NOx and SO2 without increasing particulate emissions
Papagiannakis and Hountalas [19] Total BSFC is inferior compared to the normal diesel operation at the same engine operating conditions. Lower peak cylinder
pressure compared to the normal diesel operation at a given load condition. At low load, the combustion duration is longer
compared to normal diesel operation, while at high load, it is shorter. Lower NOx, drastic decrease in soot emissions, but
considerably higher CO and HC emissions
Singh et al. [30] A minor reduction in engine output (1–2%) compared to the diesel case. The CO, CO2, NO and NO2 reduces, while HC and
exhaust gas temperature increases

Effect of engine speed

Mansour et al. [12] A light power and torque losses with regard to the diesel one, except around the speed of 2400 rpm. Slightly higher
equivalence ratios for a given speed condition. Maximum combustion pressure is slightly higher for all engine speeds
than the diesel fuelling level
Selim [14] Pressure rise rate decreases with increase in engine speed and is higher than that for diesel case
Selim [4] hth improves with increasing engine speed. The pressure rise rate decreases as the engine speed increases for the three
dual-fuel cases

Effect of pilot fuel injection timing

Nwafor [2] Offers higher BSFC and lower power output. Lower BTE with both advanced and standard timing compared to diesel case.
Standard dual timing shows longer delay periods at high loads than the advanced injection timing operation. Higher HC
emissions. High exhaust temperatures with advanced timing
Selim [14] Pressure rise rate is higher than that for the 100% diesel engine as the injection advance increases (25–408 BTDC). For late
injection of pilot (20–258 BTDC), combustion noise is comparable for diesel and dual-fuel cases
Abd et al. [33] An improvement in thermal efficiency is achieved by advancing the injection timing. Advancing the injection timing at
medium and high loads led to early knocking. Increase in NOx, reduction in CO and UBHC emissions with advance
injection timing
Selim [4] Advancing the pilot fuel injection timing reduces the torque output and the thermal efficiency but, it increases the
maximum pressure and maximum pressure rise rate
Krishnan et al. [35] Fuel conversion efficiency increases with injection timing of 15–458 BTDC and decreases afterwards. Peak heat release rates
are decreased with retard injection timing. Higher NOx emissions for timing 458 BTDC than those for very retarded
(158 BTDC) or very advanced (608 BTDC) timings

Effect of mass of pilot fuel inducted

Badr et al. [9] Exhaust emissions of the CO and the unconverted CH4 are unaffected by the pilot quantity within a limiting equivalence ratio
Abd et al. [36] An improvement in thermal efficiency by increasing the amount of pilot fuel. Increasing the amount of pilot fuel at high loads
leads to early knocking. Increase in NOx, reduction in CO and UBHC by increasing the amount of pilot fuel
 B.B. Sahoo et al. / Renewable and Sustainable Energy Reviews 13 (2009) 1151–1184 1181

Table 8 (Continued )
Researcher(s) Performance of dual-fuel diesel engine

Selim [14] Increasing the pilot fuel diesel mass increases the engine torque. The maximum pressure and the maximum rate of pressure
rise is minimum for a pilot fuel mass of 0.52 kg/h, and it is increased for a lower or higher amount of diesel pilot fuel
Nwafor [37] Increasing the pilot fuel and reducing primary fuel reduces the knocking phenomena
Selim [4] Increasing the quantity of pilot diesel fuel increases the torque output and thermal efficiency for all fuels. Increasing the
pilot fuel mass results in higher maximum combustion pressure but reduced maximum pressure rise rate

Effect of engine compression ratio

Selim [4] Knock starts earlier when a high compression ratio is used in the dual-fuel engine, and this is more notable for LPG.
Increasing the compression ratio generally increases the combustion noise

Effect of engine intake manifold conditions

Kusaka et al. [39] EGR with intake heating improves thermal efficiency. EGR controls the rate of pressure rise. Excessive EGR ratio (>50%)
causes the deterioration of combustion characteristics. EGR with intake heating reduces THC and NOx emissions

Effect of type of gaseous fuel

Bari [40] 40% CO2 in biogas do not deteriorate the engine performance much as compared to the engine with NG (96% methane). But,
30% CO2 in biogas improves the engine performance as compared to the same running with NG
Henham and Makkar [3] Overall efficiency falls with gas mixture substitution and adding CO2 affects this more at higher speed. 60% gasoil substitution
is possible without knock. Exhaust temperature is affected more by NG substitution than by CO2 addition except at
maximum NG substitution. CO is affected mainly by NG substitution and less by gas quality. More rapid pressure rise on
combustion with dual-fuel operation
Papagiannakis and Hountalas [24] At high load slope of ‘bsfc’ increment with NG mass ratio is lower compared to the one at part load. Cylinder pressure
decreases with increasing in the amount of NG. Peak cylinder pressure decreases with increase in quantity of gaseous
fuel at constant load. Combustion duration increases with increasing NG mass ratio. Sharp decrease in soot concentration,
lower NO, and high HC and CO emissions with increase in NG mass ratio
Selim [4] Increasing the mass of gaseous fuel increases the combustion noise and maximum pressure for all the three gaseous fuels

Fig. 90. Effects of mass of gaseous fuel used on performance and noise: N = 1300 rpm, IT = 358 BTDC, CR = 22, md = 0.37 kg/h [4].
1182 B.B. Sahoo et al. / Renewable and Sustainable Energy Reviews 13 (2009) 1151–1184

emissions with increasing gaseous fuel mass ratio until a certain role in the performance of dual-fuel diesel engines. Some of the
limit where they start to decrease. This is due to the increase of salient points showing the effect of above listed parameters on the
burnt gas temperature, which promotes the oxidation of UBHC. performance of dual-fuel engines are listed below.
Under dual-fuel operation, the filling of the crevice volumes with
unburned mixture of air and gaseous fuel during compression and 6.1. Effect of engine load
combustion while the cylinder pressure continues to rise, is an
important source dominating the formation of HC emissions.  The dual-fuel engine performance decreases at part load
With a close look at Fig. 89, it is revealed that at part load, conditions. There is a minor reduction in power output and
increasing the amount of gaseous fuel leads to a sharp increase of higher BSFC for the engines.
CO concentration. This is due to the slow combustion rate of  Lower peak cylinder pressure is for a dual-fuel engine compared
gaseous fuel, which maintains the charge temperature at low to the normal diesel engine at a given load condition, which is
levels resulting to a reduction of the oxidation process of CO. At encouraging since no danger exists for the engine structure.
high engine load, CO emissions increase with increasing NG mass Pressure rise rate (dP/du) increases with increase in load and is
ratio and beyond a certain value of gaseous fuel mass ratio they always higher than that of diesel fuel case.
start to decrease, as a result of the high gas temperature and faster  Combustion duration is longer compared to diesel operation at
combustion rate. In general, CO emission values under dual-fuel low load.
operation are considerably higher compared to normal diesel  Lower NOx and drastic decrease in soot emissions with all
operation. gaseous fuels. But, at all load conditions, CO and HC emissions are
Selim [4] has examined the effects of the amount of gaseous considerably high compared to the diesel case.
fuel, as a fraction of the total amount of fuel, in a dual-fuel diesel
engine. During these experiments the constant parameters are; 6.2. Effect of engine speed
engine speed 1300 rpm, pilot fuel injection timing 358 BTDC, mass
of pilot fuel 0.37 kg/h, compression ratio 22 and the amount of  Thermal efficiency improves with increasing engine speed.
liquid diesel fuel. Slightly higher equivalence ratios for a given speed condition
Fig. 90a shows that for all the three gases the output torque of dual-fuel engines.
increases with increasing the amount of gaseous fuel. It is noticed  Maximum combustion pressure is slightly higher than the diesel
that the output torque and the thermal efficiency for the dual-fuel fuelling level at constant engine speed.
engine using pure methane is higher than that of the NG mixture  Pressure rise rate decreases with increase in engine speed and is
which is higher than LPG. This is due to the higher LHV for methane higher than that for diesel case.
(50 MJ/kg) compared to the natural gas (47.7 MJ/kg) mixture and
LPG (46.1 MJ/kg). The dual-fuel engine, for all fuels used, however, 6.3. Effect of pilot fuel injection timing
suffers from low thermal efficiency at part load, and then it
increases with increasing load by increasing the mass of gaseous  An improvement in thermal efficiency is achieved by advancing
fuel admitted (Fig. 90b). As the amount of gaseous fuel increases, the injection timing.
the maximum combustion pressure and pressure rise rate increase  Maximum pressure and pressure rise rate is higher for the
for all three gaseous fuels (Fig. 90c). Increasing the load at constant advanced injection timing compared with diesel case.
speed resulted in an increase in the mass of gaseous fuel admitted  Advancing the injection timing at medium and high loads led to
to the engine, since the pilot mass injected is constant at all loads. early knocking.
This increase in the mass of gaseous fuel causes an increase in the  Increase in NOx, and a reduction in CO and UBHC emissions with
ignition delay period of the pilot diesel. Then, the pilot fuel auto- advance injection timing.
ignites and starts burning the gaseous fuel at a higher rate of
pressure rise. LPG produces the highest pressure rise rate as 6.4. Effect of mass of pilot fuel inducted
compared to methane and the NG mixture prior to knocking,
Fig. 90c, because of its high tendency to self-ignite, and produce  There is an improvement in thermal efficiency and torque output
knocking combustion. The maximum pressure for the LPG case also by increasing the amount of pilot fuel.
appears to be the highest, followed by methane and then NG  Increasing the pilot fuel mass results in higher maximum
mixture (Fig. 90d). This is due to the early ignition of the LPG that combustion pressure but reduced maximum pressure rise rate.
produces higher maximum pressure BTDC, which tends to reduce  Early knocking with increase in the amount of pilot fuel at high
the torque output produced for LPG (Fig. 90a). loads.
The above literature studies, by different researcher(s), have  Increasing the pilot fuel and reducing primary fuel reduces the
revealed the effect of engine parameters and type of gaseous fuel knocking phenomena.
on the performance of different dual-fuel diesel engines. The  Higher NOx and reductions in CO and UBHC by increasing the
comparison of results concluded by these researcher(s) from their amount of pilot fuel.
experimental programs is summarised in Table 8.
6.5. Effect of engine compression ratio
6. Conclusion
 Knock starts earlier when a high compression ratio is used.
Researchers in various countries have carried out many  Increasing the compression ratio generally increases the
experimental works using gaseous fuels as diesel engine fuel combustion noise.
substitute in a dual-fuel mode of operation. An attempt has been
made here to review the previous studies on dual-fuel concept. The 6.6. Effect of engine intake manifold conditions
overall observation from these experimental results is that the
engine operating and design parameters, namely, load, speed, pilot  EGR with intake heating improves thermal efficiency.
fuel injection timing, pilot fuel mass, compression ratio, inlet  Excessive EGR ratio (>50%) causes the deterioration of combus-
manifold conditions, and type of gaseous fuel play an important tion characteristics.
 B.B. Sahoo et al. / Renewable and Sustainable Energy Reviews 13 (2009) 1151–1184 1183

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