hydrogen

1. Effect of hydrogen and LPG addition on the efficiency
and emissions of a dual fuel diesel engine
D.B. Lata a,
*, Ashok Misra a
, S. Medhekar b
a
Department of Mechanical Engineering, Birla Institute of Technology, Mesra, Ranchi 835215, India
b
Department of Applied Physics, Birla Institute of Technology, Mesra, Ranchi 835215, India
a r t i c l e i n f o
Article history:
Received 27 October 2011
Received in revised form
3 January 2012
Accepted 4 January 2012
Available online 10 February 2012
Keywords:
Dual fuel engine
Diesel engine
Alternative fuels
Hydrogen
LPG
a b s t r a c t
This paper presents some experimental investigations on dual fuel operation of a 4
cylinder (turbocharged and intercooled) 62.5 kW gen-set diesel engine with hydrogen,
liquefied petroleum gas (LPG) and mixture of LPG and hydrogen as secondary fuels. Results
on brake thermal efficiency and emissions, namely, un-burnt hydrocarbon (HC), carbon
monoxide (CO), NOx and smoke are presented here. The paper also includes vital infor-
mation regarding performances of the engine at a wide range of load conditions with
different gaseous fuel substitutions. When only hydrogen is used as secondary fuel,
maximum enhancement in the brake thermal efficiency is 17% which is obtained with 30%
of secondary fuel. When only LPG is used as secondary fuel, maximum enhancement in the
brake thermal efficiency (of 6%) is obtained with 40% of secondary fuel. Compared to the
pure diesel operation, proportion of un-burnt HC and CO increases, while, emission of NOx
and smoke reduces in both cases. On the other hand, when 40% of mixture of LPG and
hydrogen is used (in the ratio 70:30) as secondary fuel, brake thermal efficiency enhances
by 27% and HC emission reduces by 68%. Further, shortcoming of low efficiency at lower
load condition in a dual fuel operation is removed when a mixture of hydrogen and LPG is
used as the secondary fuel at higher than 10% load condition.
Copyright ª 2012, Hydrogen Energy Publications, LLC. Published by Elsevier Ltd. All rights
reserved.
1. Introduction
One of the main purposes of improving the combustion
process of conventional internal combustion engine is to find
out useful ways to reduce exhaust emissions without doing
major alteration on their mechanical configuration [1]. With
the increasing need to conserve fossil fuels and minimize
toxic emission, much effort is being focused on the advance-
ment of present combustion technology [2]. This has stimu-
lated exploration and testing of several alternative fuels. The
main exhaust emissions from diesel engine are smoke, NOx
and particulate matter. The only option to reduce these
pollutants is to use alternative fuels which do not have sulfur
dioxide, aldehydes and ketones [3]. Liquid fuels like alcohols,
vegetable oils, bio-diesel, emulsified bio-diesel, dimethyl ether
and gaseous fuels like biogas, liquefied petroleum gas (LPG),
compressed natural gas (CNG), hydrogen, producer gas and
liquefied natural gas (LNG) have been thoroughly explored as
alternative fuels [4e8]. However, gaseous fuels have a domi-
nant position among these [8].
In addition to the single alternative fuel operation, dual
fuel engines (working on dual fuel principle) have been
a subject of high interest due to their potential to reduce
smoke emission with improved performances [9]. They
* Corresponding author: Tel.: þ91 09431382608; fax: þ91 06512275401.
E-mail address: devdaslata@yahoo.com (D.B. Lata).
Available online at www.sciencedirect.com
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i n t e r n a t i o n a l j o u r n a l o f h y d r o g e n e n e r g y 3 7 ( 2 0 1 2 ) 6 0 8 4 e6 0 9 6
0360-3199/$ e see front matter Copyright ª 2012, Hydrogen Energy Publications, LLC. Published by Elsevier Ltd. All rights reserved.
doi:10.1016/j.ijhydene.2012.01.014

2. exhibit good thermal efficiency and low smoke level at high
power output [10]. Most dual fuel engines use gaseous fuel as
secondary fuel (mostly inducted with the air during induction
process). Use of fuels with wide flammability limits and high
flame velocity can reduce these effects. Therefore, hydrogen
(as secondary fuel) becomes a natural choice as it exhibits
wide flammability limits and high flame velocity, and reduced
pollution.
Considerable research works have been done on hydrogen
as an alternative fuel in the I.C. engine as hydrogen is non-
toxic, odorless and undergoes complete combustion. It is
considered to be perfect energy storage medium and is
produced from fossil or non-fossil sources [11]. In addition,
due to availability and emission concern, LPG (as secondary
fuel) has also been a subject of high interest in diesel engines.
Hydrogen and LPG have auto-ignition temperatures of
858 K and 788 K, respectively, and require an ignition source to
be used in an I.C engine [12]. The diesel fuel which has an
auto-ignition temperature of 525 K can be used as a pilot fuel
to ignite hydrogen and LPG.
Several researchers have carried out works on either
hydrogen [3,13e17] or LPG only [2,18e20]. Saravanan et al. [3]
used hydrogen as an air-enrichment medium while diesel as
an ignition source in a stationary diesel engine system. The
stationary diesel engine could be operated with less fuel than
the pure diesel operation and found that it reduces smoke and
emission of particulate matter and NOx. The increase in brake
thermal efficiency results in lower specific energy consump-
tion. Liew et al. [13] used enriched hydrogen as a secondary
fuel and observed that by large amounts of hydrogen at higher
load conditions increased the peak cylinder pressure and peak
heat release rate. It was further observed that as compared to
two-stage combustion process of diesel engine of H2ediesel
dual fuel showed three phases of combustion process.
The hydrogen was injected into the intake port of a single
cylinder diesel engine by Saravanan et al. [14]. The diesel fuel
was injected directly within the combustion chamber during
compression process. The brake thermal efficiency in hydro-
genediesel dual fuel operation was observed to be increased
by 15% as compared to pure diesel operation at 75% load
condition while NOx emission was increased by 1e2% in
hydrogenediesel dual fuel operation at full load condition.
Boretti [15] modified diesel injector of a common rail diesel
engine by incorporating double injector one for hydrogen and
other for diesel. It was observed that the dual fuel hydro-
genediesel engine had higher brake thermal efficiency nearer
to 40% as compared to original diesel engine and it reduces
loss in part load brake thermal efficiency also. It showed mean
effective pressure more than 35 bar as compared to 25 bar of
diesel operation.
Hydrogen assisted diesel combustion was investigated by
Lilik et al. [16] on a DDC/VM Motori 2.5 L, 4-cylinder, turbo-
charged, common rail, direct injection light-duty diesel
engine, with a focus on exhaust emissions. Hydrogen was
replaced for diesel fuel on an energy basis of 0%, 2.5%, 5%,
7.5%, 10% and 15% by aspiration of hydrogen into the engine’s
intake air. Four speeds and load conditions were investigated
(1800 and 3600 rpm at 25% and 75% of maximum output). A
significant retardation of injection timing by the engine’s
electronic control unit (ECU) was observed during the
increased aspiration of hydrogen. The retardation of injection
timing reduces NOx emission considerably; however, the
same emission reductions were attained without aspirated
hydrogen by manually retarding the injection timing. Conse-
quently, hydrogen assisted diesel combustion was studied
with the pilot fuel and main injection timings locked to
examine the effects caused directly by hydrogen addition.
Hydrogen assisted diesel combustion resulted in a modest
increase of NOx emissions and a change in NO/NO2 ratio in
which NO emissions decreased and NO2 emissions increased
with NO2 becoming the main NOx constituent in some
combustion methods.
Liu et al. [17] investigated nitrogen dioxide (NO2) emissions
of a heavy-duty diesel engine operated in hydrogen H2ediesel
dual fuel combustion mode with H2 added into the intake air.
The detailed effects of H2 addition and engine load condition
on NO2 emissions were studied at 1200 RPM. The addition of
a small amount of H2 significantly increased the emissions of
O2 and the NO2/NOx ratio, particularly at lower load condition.
Increasing the engine load was found to slow down the
enhancing effect of H2 on the conversion of NO to NO2 with
the maximum NO2/NOx ratio was observed at lower H2
concentration. The maximum NO2 emission of the H2ediesel
dual fuel operation was three to five times more than that of
diesel operation at 70% and 10% load conditions, respectively.
Gomes Antunes et al. [18] achieved higher fuel efficiency in
hydrogen-fueled engine by approximately 43% as compared to
28% in the conventional diesel engine along with 20% reduction
in nitrogen oxide formation than diesel engine due to direct
injection of hydrogen in a diesel engine. Saravanan et al. [19]
experimentally investigated NOx reduction characteristics by
exhaust gas recirculation in a dual fueled engine using hydrogen
and diesel. Hydrogen was injected in intake port and diesel was
injected directly inside the cylinder. The injection timing and
injection duration of hydrogen were optimized based on the
performance and emissions. The best result was observed at the
start of injection 5
before gas exchange top dead center (BGTDC)
and injection duration of 30
crank angle. The flow rate of
hydrogen was optimized at 7.5 lpm and 25% EGR.
Szwaja et al. [20] investigated both pure hydrogen
combustion under HCCI (homogeneous charge compression
ignition) conditions and hydrogenediesel combustion in
a compression ignition (CI) engine from 0% to 17% with respect
to energy percentage. With 17% of hydrogen, the hydro-
genedieseleair mixture was stoichiometric and provided
favorable conditions for generating combustion knock. When
5% of hydrogen was added to a diesel engine it shortened the
diesel ignition lag and decreased the rate of pressure rise. Bir-
tas et al. [21] conducted investigation on a tractor diesel engine
running with small amounts of gas provided by a water elec-
trolyzer which was inducted in the cylinder. It was found that
the addition of HRG gas has a minor negative impact, up to 2%,
on the engine brake thermal efficiency. Smoke is significantly
reduced, up to 30%, with HRG enrichment, while NOx concen-
tration varied up to 14%, in both ways, depending on the engine
operation mode. A relative small amount of HRG gas could be
used with favorable effects on emissions and with a small
penalty in thermal efficiency.
Poonia et al. [22] experimentally investigated the effects of
intake charge temperature, pilot fuel quantity, and exhaust
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3. gas recirculation on dual fuel diesel engine. It was observed
that throttling of the intake charge could improve the
performance of LPGediesel dual fuel engine. It was further
observed that at low outputs the increase in pilot fuel quantity
and intake temperature showed good results. The HC emis-
sion increases with increase in pilot fuel quantity and intake
temperature, while, exhaust gas recirculation (EGR) combined
with intake heating results in increased brake thermal effi-
ciency and lowered HC emission.
Akansu and Bayrak [23] studied a single-cylinder four-
stroke SI engine which was operated with LPG (C4H10/C3H8,
70:30), hydrogen and methane mixture (H2/CH4, 30:70).
Experiments were conducted at excess air ratio between 0.8
and 1.5. Spark timing was varied from 14 to 35
CA BTDC under
a constant load condition of 6 Nm at 1400 rpm. NO emission
related to CH4/H2 was higher than those of LPG fuel. NO
emission was highest at 30
CA (BTDC) spark timing for the
each fuel. The UHC, CO and CO2 emissions of LPG fuel were
higher than those of CH4/H2 (70/30) fuel.
Vijayabalan et al. [24] modified single cylinder vertical
air-cooled diesel engine to use LPG in dual fuel mode to
study the performance, emission, and combustion charac-
teristics. The LPG was mixed with air, compressed, and
ignited by a small pilot spray of diesel. The dual fuel engine
showed a reduction in oxides of nitrogen and smoke in
entire range of load condition. However, it exhibits poor
brake thermal efficiency and higher hydrocarbon and
carbon monoxide emissions at lower load conditions due to
poor ignition. The lower load performances were improved
by introducing glow plug inside the combustion chamber.
The brake thermal efficiency was improved by 3% in the
glow plug assisted dual fuel mode, while reduction in
hydrocarbon, carbon monoxide, and smoke emissions by
69%, 50% and 9%, respectively, were observed at lower load
condition. However, presence of glow plug left the NOx
emission unaffected.
In all these works, neither hydrogen nor LPG enhances all
desired features of a dual fuel engine and therefore, it is
reasonable to try a mixture of hydrogen and LPG as a secondary
fuel.
Recently, authors reported detailed theoretical and/or
experimental studies on dual fuel engine [25e27] in which
hydrogen alone, LPG alone and a mixture of hydrogen and LPG
were used as secondary fuels and the performances in these
three distinct cases were compared. Theoretical models were
presented to predict pressure, net heat release rate, mean gas
temperature, and brake thermal efficiency. The emphasis was
given on spray mixing characteristics, flame propagation,
equilibrium combustion products and in-cylinder processes,
which were computed using empirical equations and compared
with experimental results [25]. The effect of gaseous fuel
substitutions on the ignition delay period which affects the
performance of dual fuel diesel engine was also analyzed [26]. A
detailed account on maximum rate of pressure rise, peak
cylinder pressure, heat release rate in first phase of combustion
and combustion duration at a wide range of load conditions
with different gaseous fuel substitutions was presented in [27].
However, emission was not investigated in the above
references [25e27]. Moreover, brake thermal efficiency was
investigated only at higher loads.
The understanding of brake thermal efficiency and emis-
sion of unburned hydrocarbon (HC), carbon monoxide (CO),
oxides of nitrogen (NOx) and smoke is indispensable to judge
the overall performance of a dual fuel diesel engine. There-
fore, this paper presents detailed experimental results on the
performance of a dual fuel turbocharged multi-cylinder diesel
engine with hydrogen, LPG and mixture of LPG and hydrogen
as secondary fuels at different load conditions and gaseous
fuel substitutions.
Though dual fuel operation is advantageous in many
respect, it normally possess low efficiency at lower load
condition as low fuel concentration results in increased igni-
tion delay period of the pilot fuel that causes poor combustion
[9]. Poor combustion results in high carbon monoxide and un-
burnt emission [28]. One major finding of the present inves-
tigations is that above mentioned shortcoming is removed
when a mixture of hydrogen and LPG is used as the secondary
fuel at higher than 10% load condition.
There are various techniques for fuel induction in the
cylinder such as carburetion, timed manifold/port injection
(TMI), direct hydrogen injection (DHI) and continuous mani-
fold induction (CMI). DHI requires injector to withstand higher
combustion temperature as well as to prevent injector from
corrosion due to exhaust gases. Further, lubrication between
the injector moving parts also makes the design of DHI and
TMI more complicated. The hydrogen injection system further
requires electronically activated injectors operated by elec-
tronic control units (ECU) which must be strong in design [14].
Moreover, incorporation of TMI and DHI requires considerable
modification in turbocharged multi-cylinder engine for using
two gaseous fuels (LPG and hydrogen). In view of these facts,
CMI for turbocharged multi-cylinder diesel engine (62.5 kW)
has been adopted in the present investigations. Proper care
has been taken to avoid backfire problem.
2. Experimentation
The experimental setup to carryout experimentation in the
present paper is same as in references [25e27] and has been
described here briefly for the sake of clarity. A schematic
layout of the diesel engine test setup used during the experi-
ments is shown in Fig. 1. Table 1 shows the engine geometry
and operating parameters for the present work. The diesel
engine is modified to work on dual fuel mode by attaching
hydrogen and LPG gas cylinders in connection with the intake
manifold through flame traps, mass flow meters, followed by
a one-way non-return valve and common flame arrestor by
keeping turbocharger and its bypass active. The engine was
coupled to a 62.5 kW D.C. generator. The load on the engine
was varied by introducing five water pumps and twelve 3 kW
industrial water heaters in a set of four each. The power is
based on the electrical output of the installation. The engine
was run at a constant speed of 1500 RPM [25e27].
Exhaust gas emissions namely, CO, NOx and un-burnt
hydrocarbons (UHC) were measured by an AVL 5000 DI Gas-
analyzer. The CO was measured in volume percentage basis
whereas NOx and UHC were measured in ppm units. The
cylinder pressure was measured by piezoelectric pressure
transducer (pressure range 0e250 bars) and a charge
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4. amplifier. The pressure data were transferred to data acqui-
sition system for further analysis. A crank angle encoder
(Kistler make) with an accuracy of 1
was used for angle
measurement. After 20 min of engine operation on stabilized
conditions, pressure data were obtained for an average of 100
cycles. The mass flow rate of hydrogen and LPG was measured
by mass flow meters in liters per minute. To ensure repeat-
ability the experiments were carried out for five times.
The experiments were performed on the test engine under
the following four conditions.
(i) Case I: engine runs on diesel only.
(ii) Case II: engine runs on diesel as pilot fuel and hydrogen as
secondary fuel.
(iii) Case III: engine runs on diesel as pilot fuel and LPG as
secondary fuel.
(iv) Case IV: engine runs on diesel as pilot fuel and LPG plus
hydrogen as secondary fuel.
The experimental plan in the form of test matrix is shown
in Tables 2aec.
The uncertainty in the calculated values such as brake
power, brake thermal efficiency was analyzed. The average
percentage uncertainty in brake power was Æ1.44%, while
average percentage uncertainty in brake thermal efficiency
was Æ0.51% with variable brake power, mass of diesel, LPG
and hydrogen. The percentage uncertainties of the measuring
instruments are as follows: mass flow rate of gaseous fuels
Æ4%, mass flow rate of diesel Æ3%, speed Æ1%, HC Æ2%, CO
Æ3%, NOx Æ3%, and crank angle Æ1
. The average and standard
deviation of flow rate of diesel for the Cases I and II at five
different measured trails are shown in Table 3.
3. Results and discussion
The experimental results at rated speed of 1500 rpm, injection
pressure 260 bar and injection timing 16
BTDC are presented
Fig. 1 e Experimental setup. 1- Engine, 2- Gen-set, 3- Diesel tank and measurement system, 4- Air tank and measurement
system, 5- Gas mixture, 6- Gas analyzer, 7- PC based data acquisition system, 8- Charge amplifier, 9- Cylinder pressure sensor,
10- Crank angle encoder, 11- LPG gas cylinder, 12- Hydrogen gas cylinder, 13- Hydrogen gas flame trap, 14- LPG gas flame trap,
15- Gas flow meter, 16- Gas cylinder control valve, 17- Pressure regulator, 18- Solenoid switch valve, 19- Temperature and
Pressure measurement locations.
Table 1 e Engine specifications.
Sr. No. Parameter Engine A specification
1 Make and model Ashok Leyland ALU
WO4CT Turbocharged,
intercooler, Gen- Set
2 General details Four Stroke, Compression
Ignition, Constant Speed,
vertical, water-cooled,
direct injection, turbo
charger, Intercooler,
Gen-Set
3 No. of cylinder 4
4 Bore, mm 104
5 Stroke, mm 113
6 Rated speed, rpm 1500
7 Swept volume, cc 3839.67
8 Clearance volume, cc 84.90
9 Compression ratio 17.5:1
10 Injection pressure, bar 260
11 Injection timing BTDC 16
12 Rated power, kW at 1500 rpm 62.5
13 Inlet pressure, bar 1.06
14 Inlet temperature, K 313
15 Nozzle diameter, mm 0.285
16 Number of hole 5
Table 2a e Test matrix.
Case No. Primary
fuel
Secondary
fuel
Load percentage
(%)
I Diesel e 10, 20, 30, 40, 60, 80
II Diesel Hydrogen 10, 20, 30, 40, 60, 80
III Diesel LPG 10, 20, 30, 40, 60, 80
IV Diesel LPG þ hydrogen 10, 20, 30, 40, 60, 80
i n t e r n a t i o n a l j o u r n a l o f h y d r o g e n e n e r g y 3 7 ( 2 0 1 2 ) 6 0 8 4 e6 0 9 6 6087

5. for Cases II, III, and IV at different load conditions as per
references [25e27]. The quantities of hydrogen and LPG in
Cases II, III and IV were supplied at predetermined rate as
given in text matrix (Tables 2aec) and diesel fuel was auto-
matically catered through fuel metering and governing
system at all load conditions. Zero percent gaseous fuel
substitution represents pure diesel operation. Due to knock-
ing, the maximum amount of substitution by hydrogen and
LPG was limited to 50% and 70%, respectively.
The mixture of LPG and hydrogen was varied in the
following proportions in each combination (M) (Table 2c):
(LPG-90% þ H2-10%), (LPG-80% þ H2-20%), (LPG-70% þ H2-30%),
(LPG-60% þ H2-40%).
3.1. Thermal efficiency
The indicated power output is calculated as
Wi ¼
X720
i¼1
piðVi À ViÀ1Þ (1)
where Wi is work output and pi is pressure and Pi ¼ Wi(N/
2 Â 60) where Pi indicated power, N is RPM.
For good comparisons of measurements taken at different
conditions of atmospheric pressure and temperature ( patm
and Tatm) the normalized indicated power output is used
PiðnormalizedÞ ¼ pi
101325
patm
ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi
273 þ Tatm
273 þ 20
r
(2)
Then the indicated thermal efficiency is calculated from the
measured fuel consumption as
hi ¼
PiðnormalizedÞ Â 22:4
À
mdQd þ mH2
QH2
þ mLPGQLPG
Á
 1000
(3)
where md, mH2
, mLPG are the mass flow rate of diesel, hydrogen
and LPG, respectively, and Qd, QH2
, QLPG are the lower heat of
combustion of diesel, hydrogen and LPG, respectively.
Similarly, brake thermal efficiency is given as
hb ¼
BP
À
mdQd þ mH2
QH2
þ mLPGQLPG
Á
 1000
(4)
The variation of thermal efficiency at different load condi-
tions and for different substitutions of hydrogen (Case II), LPG
(Case III) and mixture of LPG and hydrogen (Case IV) are
shown in Figs. 2e7, respectively. It is observed that the
thermal efficiencies of the Cases II, III and IV are lower than
pure diesel operation (Case I) at lower load conditions. This
may be because of the fact that the pilot diesel fuel (1.72 mg/
cycle/cylinder) is low in quantity at lower load conditions.
Thus, lesser ignition centers are formed due to less quantity of
injected pilot diesel. Addition of gaseous fuel further reduces
pilot diesel fuel (1.2 mg/cycle/cylinder). And hence, poor
ignition of gaseous fuel results in lower brake thermal effi-
ciency. Moreover, the inducted gaseous-air forms too lean
mixture to burn well (at lower load conditions) that leads to
low combustion rate and flame quenching.
The indicated thermal efficiencies at 10% load conditionsand
for 30% gaseous fuel substitution in Cases II and III are found to
be17.3%and20.16%, respectively,as compared to 23.48%of Case
I operation (see Figs. 2 and 3). The thermal efficiency of Case II
(Fig. 2) is least due to higher cooling losses in Case II compared to
Case III. The higher cooling losses in Case II may be due to lower
quenching distance and higher thermal conductivity [29]. LPG
has higher pre-ignition energy release rate as compared to
hydrogen [2]. The lower brake thermal efficiency of Case III
probably results from sluggish combustion [24].
It was observed that the dual fuel diesel engine shows lower
thermal efficiency at lower load conditions as compared to
diesel. This may be due to the fact that even at low concen-
tration of hydrogen or LPG in the intake air, the combustion
spreads throughout the gaseair mixture. This causes high heat
transfer losses to the adjacent walls. While, in case of diesel
engines under light load condition, the penetration of the
diesel spray is such that it does not reach the cylinder walls and
the combustion is confined to piston bowl [28], and also, the
surrounding coatings of air acts as insulation in between burnt
gases and the walls, which reduces heat losses thereby giving
better thermal efficiencies with diesel.
The diesel fuel was introduced into the cylinder through
a conventional individual type diesel injection system. The
pilot diesel quantity injected was found to be very small.
Hence, to enhance the performance at low load condition in
dual fuel engine operation, there is need for careful assess-
ment of the optimum injection characteristics, including
injection pressure, number of jets involved, location of the jet
in the cylinder, jet orifice diameter and type of the nozzle
being used.
Table 2b e Test matrix.
Case No. Primary fuel Secondary fuel Secondary fuel substitution as % of
diesel at each load %
I Diesel e e
II Diesel Hydrogen H2-5, H2-10, H2-15, H2-20, H2-25, H2-30,
H2-40, H2-50.
III Diesel LPG LPG-10, LPG-20, LPG-30, LPG-40, LPG-50,
LPG-60, LPG-70.
IV Diesel LPG þ Hydrogen (M) M ¼ Mixture (%) M-30, M-40, M-50, M-60, M-70.
Table 2c e Test matrix.
Mixture (M) Mixture composition as in Case IV (% of
LPG þ % of H2) in each % mixture (M)
LPG-90 þ H2-10
LPG-80 þ H2-20
LPG-70 þ H2-30
LPG-60 þ H2-40
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6. When the load is increased up to 40%, the thermal effi-
ciency for the Cases II and III are found to be 29.4% and 27.84%,
respectively, which are still less compared to 32.1% of Case I
operation (see Figs. 2 and 3). It is known that the laminar flame
velocity of diesel, hydrogen and LPG are 4.8 cm/s, 117 cm/s
and 5.12 cm/s respectively at overall equivalence ratio of 0.6
[13,14], i.e., gaseous fuels have higher flame velocity and
diffusivity than diesel fuel. Therefore, these gaseous fuels
consume most part of the oxygen from the entrained air
during main part of the combustion. Hence, a part of injected
diesel goes into the exhaust without taking part in combus-
tion (higher unburned hydrocarbon emissions in these cases
confirm it).
One can note that at 80% load condition, the thermal effi-
ciency for Case II (42.14%) and Case III (37.66%) becomes
higher as compared to Case I (36.19%). This may be due to the
increase in combustion rate with larger pilot diesel quantities
(as an average 5.44 mg/cycle/cylinder) as evident from the
cylinder pressure verses crank angle diagram shown in Fig. 8
which leads to stronger ignition sources and hence, more
complete and better combustion of gaseous fuel. The flam-
mability limits of premixed hydrogen flame are much wider
than the diesel fuel. Therefore, cylinder chamber gets fumi-
gated with hydrogen which enhances burning of the diesel
fuel. Furthermore, the temperatures at the time of injection of
pilot diesel fuel due to pre-ignition reaction energy for the
Cases II and III are observed to be 569.36 K and 692.99 K,
respectively. With the increase in hydrogen concentration,
the pressure increases due to high flammability limits and
rate of combustion of hydrogen. Moreover, hydrogen has
a higher flame velocity at stoichiometric conditions which
makes engine combustion operation getting closer to the
thermodynamically ideal engine. This results in accelerated
pre-flame reactions, reduced ignition delay period and rapid
combustion of gaseous fuel due to higher flame velocity [15]. It
was observed by Boretti [15] that the thermal efficiency
increases up to 40% as compared to original diesel engine and
it also reduces loss in part load brake thermal efficiency due to
direct injection of hydrogen inside the common rail diesel
0 10 20 30 40 50 60 70 80 90 100
0
10
20
30
40
50
Diesel
Diesel+10% Hydrogen
Diesel+20% Hydrogen
Diesel+30% Hydrogen
Diesel+40% Hydrogen
Diesel+45% Hydrogen
Diesel+50% Hydrogen
IndicatedThermalEfficiency(%)
Indicated Load (%)
Indicated Thermal Efficiency (%) vs Indicated Load (%)
for Diesel + Hydrogen
Fig. 2 e Indicated thermal efficiency (%) vs. indicated load
(%) for diesel D hydrogen.
0 10 20 30 40 50 60 70 80 90 100
0
10
20
30
40
50
Diesel
Diesel + 10% LPG
Diesel + 20% LPG
Diesel + 30% LPG
Diesel + 40% LPG
Diesel + 50% LPG
Diesel + 60% LPG
IndicatedThermalEfficiency(%)
Indicated Load (%)
Indicated Thermal Efficiency (%) vs Indicated Load (%) for Diesel + LPG
Fig. 3 e Indicated thermal efficiency (%) vs. indicated load
(%) for diesel D LPG.
Table 3 e Measurement of diesel consumption for only diesel operation and diesel plus hydrogen operation.
Load (%) Trail 1
(seconds/100 ml
of diesel)
Trail 2 Trail 3 Trail 4 Trail 5 Average SD Deviation
(1.65 Â SD)
Variance Values lies
between average
(Æ) (1.65 Â SD)
Only diesel
2 171 175 177 178 169 174 3.46 5.72 12 179.72e168.28
12 142 148 145 138 143 143.2 3.31 5.46 10.96 148.66e137.73
36 66 64 68 62 70 66 2.83 4.67 8 70.66e61.33
60 42 46 47 49 45 45.8 2.32 3.82 5.36 49.62e41.97
84 30 26 39 40 37 34.4 5.46 9.01 29.84 43.41e25.38
Diesel þ hydrogen
2 200 212 205 198 210 205 5.44 8.98 29.6 213.97e196.02
12 130 138 137 137 133 135 3.03 5.01 9.2 140.01e129.99
36 50 48 56 54 53 52.2 2.85 4.71 8.16 56.91e47.49
60 43 48 47 49 43 46 2.53 4.17 6.4 50.17e41.82
84 30 35 37 31 29 32.4 3.07 5.07 9.44 37.47e27.33
i n t e r n a t i o n a l j o u r n a l o f h y d r o g e n e n e r g y 3 7 ( 2 0 1 2 ) 6 0 8 4 e6 0 9 6 6089

7. injection system. It may be due to some part of air being
possibly replaced by hydrogen.
It is worth to be mentioned here that the thermal efficiency
starts decreasing beyond 45% of hydrogen substitution in Case
I and 40% of LPG substitution in Case II (Figs. 2 and 3). Since
a higher concentration of hydrogen or LPG leads to higher
flame velocity and higher burning rate resulting in rapid
consumption of oxygen in premixed combustion phase,
therefore, lesser oxygen becomes available in the diffusion
combustion phase, resulting in a decrease in efficiency. Fig. 4
shows comparison of brake thermal efficiency for the Cases II
and III. Case III shows higher thermal efficiency at 10% load
condition, while Case II shows higher brake thermal efficiency
at 80% load condition. LPG shows higher pre-ignition energy
rate at lower concentration and load, while hydrogen shows
the same at higher concentration and load condition [9].
Poonia et al. [22] observed that lower load performance of
LPGediesel dual fuel engine was improved by the intake
charge temperature, increase in pilot fuel quantity, and
exhaust gas recirculation.
In Case IV, results are analyzed here for 30% and above of
mixture (LPG þ hydrogen) in diesel as experiments on the
Cases II and III revealed that brake thermal efficiency drops at
lower gaseous fuel substitution and load conditions.
Figs. 5e7 show variation of brake thermal efficiency with
gaseous (mixture) fuel substitution at 10%, 40% and 80% load
condition, respectively. In each figure, brake thermal effi-
ciency for four different mixture combinations (as mentioned
in Table 2c) has been plotted.
At 10% load, the brake thermal efficiency verses gaseous
fuel substitution for four mentioned combinations is shown in
Fig. 5. It can be seen in the figure that brake thermal efficiency
decreases with gaseous fuel substitution. The brake thermal
efficiency in this case is always less than in Case I. For
Fig. 4 e Comparisons of brake thermal efficiency for the
cases II and III at different load conditions.
Fig. 5 e Brake thermal efficiency (%) vs. gaseous fuel
substitution (%) for diesel D hydrogen D LPG substitution
at 10% load condition.
0 10 20 30 40 50 60 70 80
0
5
10
15
20
25
30
35
LPG 90% + H2
10%
LPG 80% + H2
20%
LPG 70% + H2
30%
LPG 60% + H2
40%
BrakeThermalEfficiency(η)(%)
Gaseous Fuel Substitution (%)
Brake Thermal Efficiency (%) vs. (LPG + H2
) Substitution (%)
at 40% Load (Diesel : Mixture = 70:30)
Fig. 6 e Brake thermal efficiency (%) vs. gaseous fuel
substitution (%) for diesel D hydrogen D LPG substitution
at 40% load condition.
0 10 20 30 40 50 60 70 80
0
5
10
15
20
25
30
35
40
LPG 90% + H2
10%
LPG 80% + H2
20%
LPG 70% + H2
30%
LPG 60% + H2
40%
BrakeThermalEfficiency(η)(%)
Gaseous Fuel Substitution (%)
Brake Thermal Efficiency (%) vs. (LPG + H2
) Substitution (%)
at 80% Load (Diesel: Mixture = 70:30)
Fig. 7 e Brake thermal efficiency (%) vs. gaseous fuel
substitution (%) for Diesel D hydrogen D LPG substitution
at 80% load Condition.
i n t e r n a t i o n a l j o u r n a l o f h y d r o g e n e n e r g y 3 7 ( 2 0 1 2 ) 6 0 8 4 e6 0 9 66090

8. example, brake thermal efficiency of 30% mixture is found to
be 16.1%, 15.7%, 15.3% and 14.9%, respectively, as compared to
19.57% of Case I operation. One can note that 7.45% drop in
brake thermal efficiency is there when hydrogen percentage
in the mixture is increased from 10% to 40%. This may be
because the spray does not penetrate up to the cylinder walls
due to less pilot quantity in pure diesel operation and there-
fore, the combustion is confined to piston chamber (bowl) only
[28]. Moreover, this combustion zone is surrounded by air
which acts as semi-insulator between the burned gases and
cylinder walls. Therefore, lesser amount of heat is transferred
to the cylinder walls. Hence, pure diesel operation has higher
brake thermal efficiency than Cases II, III and IV at lower load
conditions (Figs. 2, 3 and 5).
In previous situation, if load is increased up to 40% (Fig. 6),
brake thermal efficiency increases with mixture substitution
and attains a maximum at 40% of the mixture substitution and
decreases beyond it. This is true for all mixture combinations.
However, it becomes obvious from the Fig. 6 that the combi-
nation LPG:hydrogen ¼ 70:30 is the most suited combination at
this load. One can note from Table 4 that the brake thermal
efficiency obtained in Case IV even at 40% load condition is
higher than Case I (pure diesel operation) at all load conditions.
When thermal efficiency for the most suited combination
LPG: hydrogen ¼ 70:30 (Case IV) is compared with Cases I, II
and III, rise in thermal efficiency (in Case IV) is observed to be
22%, 42% and 43%, respectively. The rise in thermal efficiency
in Case IV may be due to the fact that LPG reduces the laminar
burning velocity of hydrogen and suppresses the propensity of
onset of both diffusional-thermal and hydrodynamic cellular
instabilities in hydrogen air flames. It also retards the reaction
intensity and increases the critical radius [30]. At lower load
conditions, the propane (LPG)eair flame tends to be stable and
hydrogeneair flame becomes unstable and by increasing
hydrogen fraction flame destabilization increases due to
reduction in Markstein length (Markstein length measures the
effect of curvature on a flame; larger the Markstein length,
greater the effect of curvature on burning velocity).
At 80% load condition also the brake thermal efficiency
attains a maximum at 40% of the mixture substitution and
decreases beyond it. These maxima are found to be 37.3%,
37.5%, 38.2% and 36.8%, respectively, for four mixture
combinations (see Fig. 7). The brake thermal efficiency at
same load in Case I is found to be 30.16%.
Fig. 9 compares the plots of brake thermal efficiency verses
gaseous fuel substitution for Cases I, II, III and Case IV at 80%
load condition. The mixture combination in Case IV is LPG:
320 340 360 380 400 420 440
0
20
40
60
80
100
CylinderPressure(bar)
Crank Angle (degree)
LPG
Diesel
Hydrogen
Cylinder Pressure vs. Crank Angle at 80% Load Condition
Fig. 8 e Cylinder pressure (bar) vs. crank angle (degree) at
80% load condition.
Table 4 e Comparison of brake thermal efficiency for different cases.
Case No. Primary fuel Secondary fuel Gaseous fuel
substitution (%)
10% Load 40% Load 80% Load
Case I Diesel e 0 19.57 26.75 30.16
Case II Diesel Hydrogen (H2) 30 14.4 24.5 35.12
40 14.6 23 34.0
50 12.7 22 31.01
Case III Diesel LPG 30 16.8 23.3 31.38
40 16.4 22.98 32.0
50 15.8 22.56 31.5
Case IV Diesel LPG þ H2 30 LPG 90% þ H2 10% 16.1 32.4 37.2
40 16.4 32.6 37.3
50 15 30.9 34.6
30 LPG 80% þ H2 20% 15.7 30.8 37.3
40 17.5 32.5 37.5
50 14.8 28.5 34.9
30 LPG 70% þ H2 30% 15.3 28.5 37.6
40 16.8 32.7 38.2
50 13.0 28.7 35.7
30 LPG 60% þ H2 40% 14.9 28.1 35.8
40 15.5 31.8 36.8
50 12.8 26.8 34.0
The values in bold are the parameters for which the performances of the engine is better.
i n t e r n a t i o n a l j o u r n a l o f h y d r o g e n e n e r g y 3 7 ( 2 0 1 2 ) 6 0 8 4 e6 0 9 6 6091

9. hydrogen ¼ 70:30. The figure shows 25%, 9% and 22% rise in
brake thermal efficiency in Case IV as compared to Cases I, II
and III, respectively. This could be explained as follows. At
80% load conditions, LPG flame tends to become unstable,
while, hydrogeneair flames tend to be stable. Therefore
increase in hydrogen fraction leads to stabilization of flame
[31] (due to increase in Markstein length). This nature may
also results from the opposite diffusion behavior of propane
(LPG) and hydrogen. Diffusivity of hydrogen (0.61 cm2
/s) is
more than LPG (0.12 cm2
/s) in air. Hydrogen has strong buoy-
ancy and high diffusivity while LPG has lower flame speed and
narrower flammability limits as compared to hydrogen.
The comparison of brake thermal efficiency at 10%, 40%
and 80% of load conditions at various percentages of gaseous
fuel substitutions for the Cases IeIV are shown in Table 4. It is
observed that 40% mixture of LPG þ hydrogen in the ratio
70:30 enhance brake thermal efficiency by 22% and 27% at 40%
and 80% load conditions, respectively, as compared to Case I;
however, it drops by 14% at 10% load condition.
It may be noted here that the combustion process in dual
fuel diesel engine depends on the spray and ignition charac-
teristics of the pilot diesel fuel and the type and overall
concentration of gaseous fuel used in the charge. The
combustion energy release characteristics in a dual fuel diesel
engine reveal throughout the relatively complex physical and
chemical interactions that take place between the combustion
processes of the two fuel systems. However, the chemical
kinetics of the diesel fuel and its chemical interactions with
the gaseous fuel component are too complicated and at the
moment is not clearly understood [32].
4. Emissions
4.1. Un-burnt hydrocarbons (UHC)
Fig. 10(a and b) shows variation in unburned hydrocarbon at
10% and 80% load conditions for the Cases II, III and IV. The
nature of curve at 40% load condition shows similar trends,
which represents medium load condition, and hence has not
been shown in the figures to prevent overloading of graphs.
The HC emissions at 40% load condition show similar trend as
that of 10% load, hence emissions are shown at 10% and 80%
load condition only. The over leaning and under mixing are
responsible for HC emission in diesel engine. It was observed
in earlier section that 30% hydrogen in the mixture of LPG and
hydrogen in Case IV gives maximum thermal efficiency.
Therefore, further analysis on emission for the Case IV is
carried on the same composition.
At 10% load condition, Cases II, III and IV show HC emis-
sions of 6.86 g/kWh, 5.9 g/kWh and 6.94 g/kWh, respectively,
as compared to 1.72 g/kWh of Case I. This may be due to
reduction in pilot quantity, which causes poor ignition of
gaseous fuel and inducted mixture is too lean to burn. Further,
at low loads, the gas temperature is lower, while at higher load
this rises due to faster burning of hydrogen. This leaves diesel
fuel injected toward the end of injection period, deficient in
oxygen. Further, during the compression process, the homo-
geneously mixed gaseous fuel undergoes chemical reactions
before diesel pilot fuel injection and the speed of these
chemical reactions may become high due to higher charge
temperature. When diesel pilot fuel is injected into this
environment, the traveling of flame front is complete. Hence
most part of the oxygen available in the combustion chamber
gets consumed, which results in diesel combustion to take
place in the atmosphere of lack of oxygen. Therefore, rate of
HC emissions is higher.
At 80% load condition, Cases II, III and IV show maximum
HC emission of 5.64 g/kWh, 4.57 g/kWh and 1.07 g/kWh,
respectively, as compared to 1.8 g/kWh of Case I operation. At
higher load condition Case IV shows lower HC than Case I due
to better and fast combustion rate leading to more complete
combustion and hence low HC emissions. While, in Cases II
and III, at a low pilot quantity (average 4.15 mg/cycle/cylinder)
HC emission is high at gaseous fuel substitution. High load
condition results in increased ignition delay and cylinder gas
temperature which may lead to dispersion of the pilot fuel
prior to ignition. This will lead to poor combustion of the
gaseous fueleair mixture. Higher gaseous fuel substitutions
0 10 20 30 40 50 60 70 80
0
5
10
15
20
25
30
35
40
BrakeThermalEfficiency(η)(%)
Gaseous Fuel Substitution (%)
Diesel + Hydrogen
Diesel + LPG
Diesel + LPG-70%+ H2
-30%
Brake Thermal Efficiency (%) vs. Gaseous Fuel Substitution (%)
at 80%Load
Fig. 9 e Comparisons of brake thermal efficiency for the
cases II, III and IV at 80% load condition.
Fig. 10 e Un-burnt HC (g/kWh) vs. diesel D gaseous fuels
substitution (%).
i n t e r n a t i o n a l j o u r n a l o f h y d r o g e n e n e r g y 3 7 ( 2 0 1 2 ) 6 0 8 4 e6 0 9 66092

10. which also displace some part of the air in the intake process
may lead to large consumption of oxygen and cause depletion
in the available oxygen for further combustion of pilot diesel
fuel, thus increased HC levels. Moreover, at 80% load condi-
tion, rich supply of pilot fuel does not mix properly with
gaseous fuel air mixture. Furthermore, the trapped gaseous
fuel in the piston land and other crevices may escape to the
atmosphere either as product of partial combustion or
without combustion because of flame quenching.
The increase in HC emission was also observed by Poonia
et al. [22] with increase in pilot fuel quantity and intake
temperature which reduces it by exhaust gas recirculation
(EGR) combined with intake heating. The latter also increase
brake thermal efficiency.
4.2. Carbon monoxide
The formation of carbon monoxide depends on post-oxidation
reaction. In normal diesel engine due to the presence of more
excess air, the carbon oxidation reaction is almost completed
[33]. The considerable amount of CO is not produced until the
smoke limit is reached. The variation in CO is shown in
Fig. 11(a and b). At 10% load condition Cases II, III and IV show
89%, 77% and 75% rise in CO emission, respectively, as
compared to Case I operation. At 10% load condition gaseous
fueleair mixture near the pilot is burned due to less turbu-
lence. Thus some partial oxidation product like carbon
monoxide may come out in the exhaust. At higher concen-
tration of gaseous fuel, the concentration of the partial
oxidation product could increase [34]. Moreover, inducted
mixture becomes rich due to more displacement of air. This is
thought to be the reason for the rise in CO emissions.
At 80% load condition (Fig. 11), maximum rise in CO
emission for the Cases II, III and IV is 76%, 84% and 80%,
respectively, as compared to Case I, due to rich mixture and
prominent temperature at this condition. In Cases II and IV
hydrogen shows different behavior in dual fuel engine due to
presence of liquid hydrocarbon. As soon as the ignition
starts, the spontaneous combustion occurs due to the pres-
ence of higher percentage of hydrogen. Hence, the diesel fuel
is further subjected to higher combustion temperature in an
atmosphere of shortage of oxygen. This leads to fuel cracking
and production of carbon monoxide. The excess availability
of air as in Case I reduces reaction rate due to drop in
temperature [34]. Overall, the presence of hydrogen reduces
CO because it does not contain any carbon particle and
whatever the small percentage of CO is present in the
exhaust is due to the burning of lubricating oil and partial
combustion of diesel fuel. In diesel engine, the diesel fuel is
injected at the end of compression process in the atmosphere
of high temperature and high-pressure air. This forms
stratified charge mixture and it becomes locally rich and
lean. The rate of oxidation reaction is amplified with the
increase in quantity of air which further reduces the
temperature. Further, CO emission is increased at both load
conditions due to delayed ignition period.
4.3. NOx
The nitrogen oxides mainly consist of nitric oxide (NO) and
a small amount of nitrogen dioxide (NO2). These are primarily
formed by the oxidation of atmospheric nitrogen in the
combustion chamber. NO is formed behind the flame front
during combustion of gaseous fuel. The combustion temper-
atures and the availability of oxygen mainly control NO
formation. While, formation of NOx in the dual fuel engine
mainly depends on diesel pilot spray zone. The formation of
NOx increases with the increase in the size and quantity of
pilot diesel fuel. Further, the nitrogen oxide emission
increases with the rise in cylinder temperature, oxygen
concentration and combustion duration.
Fig. 12(a and b) shows variation of NOx for the Cases II, III
and IV at 10% and 80% load conditions. It is observed that dual
fuel operation produces less NOx at all load conditions than
Case I operation. At 10% load condition 60%, 33% and 93% drop
in NOx emission were observed for the Cases II, III and IV,
respectively, as compared to Case I. Similarly, at 80% load
condition 20%, 41% and 84% reduction in NOx for the Cases II,
III and IV were observed, respectively, as compared to Case I.
This may be due to more uniform temperature distribution
obtained with the gaseous fueleair mixture. This causes
reduction in high temperature region around the diesel flame.
Fig. 11 e Carbon monoxide (g/kWh) vs. diesel D gaseous
fuels substitution (%).
Fig. 12 e NOx (g/kWh) vs. diesel D gaseous fuels
substitution (%).
i n t e r n a t i o n a l j o u r n a l o f h y d r o g e n e n e r g y 3 7 ( 2 0 1 2 ) 6 0 8 4 e6 0 9 6 6093

11. Furthermore, the NOx is formed in diesel engine during
diffusion combustion phase and on the weak side of reaction
zone. Dual fuel operation reduces the amount of pilot diesel
fuel during diffusion-controlled combustion phase. It was
observed by Saravanan [14] that NOx emission increased by
1e2% in hydrogenediesel dual fuel operation at full load
condition due to injection of hydrogen into intake manifold of
single cylinder diesel engine. This might be due to higher rise
in temperature during premixed combustion phase due to
non-steadiness of temperature distribution. Further, the start
of combustion and longer duration of premixed combustion
phase may result in higher nitric oxide emissions due to
higher in cylinder temperature and more residential time
available for the formation of NOx [34].
4.4. Smoke
Emission of smoke is due to the combined effect of poorer
liquid fuel preparation and mixing of the less quantity of pilot
diesel fuel injected. However, with increasing load, the smoke
in the exhaust of the engine running on the hydrogen and
diesel decreases as compared to pure diesel operation. The
high rate of hydrogen chain oxidation shortens the slow
combustion period even though the oxygen concentration in
the gaseous pre-mixture is reduced.
Smoke basically consists of combustion generated carbo-
naceous materials (soot) on which some organic compounds
are absorbed. It is from incomplete combustion of hydro-
carbon fuels and some amount of lubricating oil [35]. As
shown in Fig. 13(a and b), dual fuel operation (Cases IIeIV)
reduces smoke at all load conditions as compared to pure
diesel operation (Case I). The heterogeneous and diffusive
modes of combustion are responsible for smoke formation in
diesel engine. The smoke formation increases with load and
equivalence ratio. Dual fuel engine uses gaseous fuel as
a secondary fuel, which improves fueleair mixing. Hence
major part of injected pilot fuel burns in homogeneous
atmosphere of combustion and has lesser diffusion combus-
tion phase. High rate of hydrogen chain oxidation further
reduces combustion duration. Dual fuel operation reduces
smoke by around 30% than Case I operation. This is mainly
due to the fact that hydrogen does not contain any carbon
molecule; moreover, LPG has lower carbon/hydrogen ratio
[36]. In addition, LPG has lower molecular weight as well as
smaller number of carbon to carbon bonds [37]. These facts
keep engine clean and smoke free.
5. Conclusions
Experiments were performed on 4 cylinder turbocharged,
intercooled with 62.5 kW gen-set diesel engine using
hydrogen, LPG and mixture of LPG and hydrogen as secondary
fuels. The experiments were performed to measure brake
thermal efficiency and emissions namely un-burnt hydro-
carbon, carbon monoxide, NOx and smoke at different load
conditions on the following four cases.
(i) Case I: engine runs on diesel only.
(ii) Case II: engine runs on diesel as pilot fuel and hydrogen as
secondary fuel.
(iii) Case III: engine runs on diesel as pilot fuel and LPG as
secondary fuel.
(iv) Case IV: engine runs on diesel as pilot fuel and LPG plus
hydrogen as secondary fuel.
On the basis of the results and discussions presented
above, the following conclusions may be drawn.
1. Use of hydrogen or/and LPG as secondary fuel enhances the
brake thermal efficiency at high load conditions while it
produces reverse effect at low load conditions.
2. A mixture of hydrogen and LPG as secondary fuel reduces
the un-burnt hydrocarbon, NOx and smoke at higher load
conditions.
3. A severe knock was noticed during the dual fuel diesel
engine operation with 50% substitution of hydrogen in
Cases II and 70% of LPG in Cases III at higher loads, which
suggests a limiting value for the substitution by secondary
fuels.
4. Overall, the investigations show that beyond 35% load
condition, Case IV operation is always better than Case I, II
and III operations.
5. The best performances of the engine employed for inves-
tigation are obtained by the substitution of 40% of mixture
in the ratio LPG: hydrogen ¼ 70:30. This situation is most
suited in terms of efficiency and emissions.
6. One major finding of the present investigations is that
shortcoming of a dual fuel operation (low efficiency at
lower load condition) is removed when a mixture of
hydrogen and LPG is used as the secondary fuel at higher
than 10% load condition.
It is worth to add here that there is a need to develop
suitable hydrogen and LPG kits (ECU) to control supply of
gaseous fuels and diesel in a manner so as to optimize the
engine performance over the complete range of operation.
These kits should be simple and rough so that existing diesel
engines may be easily adapted for dual fuel operation.
In short, the most efficient and eco-friendly performance
of the engine is offered by the secondary fuel made by
Fig. 13 e Smoke (HSU) vs. diesel D gaseous fuels
substitution (%).
i n t e r n a t i o n a l j o u r n a l o f h y d r o g e n e n e r g y 3 7 ( 2 0 1 2 ) 6 0 8 4 e6 0 9 66094

12. a mixture of hydrogen and LPG. By implementing a suitable
ECU, brake thermal efficiency may be increased and emission
of HC and CO may be reduced significantly.
Acknowledgment
The authors are grateful to Professor Pramod S. Mehta,
Internal Combustion Engine Laboratory, Indian Institute of
Technology, Madras, Chennai for helpful discussions.
Appendix A. Analysis of uncertainty
Voltage ¼ 230 V.
Current ¼ 14 A.
Voltage error ¼ Æ1.75%, current error ¼ Æ1.5%.
Voltage ¼ 230 Â 1.75/100 ¼ 4.0125, voltage fluctuation ¼
230 Æ 4.025.
Similarly, current ¼ 14 Â 1.5/100 ¼ 0.21, current fluctuation ¼
14 Æ 0.21.
(I)Analysis of uncertainty in brake power by considering
generating efficiency 80%
BP ¼
VI
hg  1000
kW
ðBPÞ ¼ ðV; IÞ
vBP
vV
¼
I
hg  1000
¼
14
0:8 Â 1000
¼ 0:0175
vBP
vI
¼
V
hg  1000
¼
230
0:8 Â 1000
¼ 0:2875
DBP ¼
2
6
4
ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi
vBP
vV
 DV
2
þ
vBP
vI
 DI
s 2
3
7
5
DBP ¼
ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi
ð0:0175 Â 4Þ2
þð0:02873 Â 0:21Þ2
q
DBP ¼
ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi
4:9 Â 10À3 þ 3:6451 Â 10À3
p
DBP ¼ Æ0:09243 kW
BP ¼ 42 Æ 0:09243 ¼ 42:09243
BP ¼
42:09243
42
À 1
!
 100 ¼ 0:22%
Therefore, the uncertainty in brake power ¼ 0.22%.
(II)Total diesel consumption
TDC ¼
Quantity of diesel  3600  density of diesel
time  1000
TDC ¼ fðtÞonly
TDC ¼
ð100 Â 3600 Â 0:83Þ
ð200 Â 1000Þ
¼ 1:494 kg=h
vTDC
vt
¼ À
ð100 Â 3600 Â 0:83Þ
t2 Â 1000
¼ À
ð298800Þ
ð200Þ2
Â1000
¼ À7:43 Â 10À3
DTDC ¼
ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi
ð À 7:47 Â 10À3 Â 0:3Þ2
q
DTDC ¼ 2:241 Â 10À3
Therefore, the uncertainty in diesel consumption
is ¼ 1.494 Æ 2.241 Â 10À3
¼ 0.15%.
(III)Brake thermal efficiency
hth ¼
BP Â 3600 Â 100
TDC Â CV
hth ¼ fðBP; TDCÞ
hth ¼
ð2 Â 3600 Â 100Þ
ð1:494 Â 43500Þ
¼ 11:0787%
Since, diesel consumption has been calculated at 2 kW
(brake power).
vhth
vBP
¼
3600 Â 100
ð1:494 Â 43500Þ
¼ 5:5393
vhth
vTDC
¼ À
BP Â 3600 Â 100
ðTDCÞ2
Â43500
¼ À7:4155%
Dhth ¼
vhth
vBP
 DBP
2
þ
vhth
vTDC
 DTDC
2
#1
2
Dhth ¼
h
ð5:5393 Â 0:09243Þ2
þðÀ7:4155  0:002241Þ2
i1
2
Dhth ¼ 0:51%
Therefore, the uncertainty in Brake thermal efficiency ¼
11.0787 Æ 0.51%.
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