Article Title 1
Impact of diesel-hythane dual-fuel combustion on engine performance and 2
emissions in a heavy-duty engine at low-load condition 3
4
5
Authors: 6
K. Longo
1
, X. Wang
1
and H. Zhao
1
7
1
Centre for Advanced Powertrain and Fuels, Brunel University London, Kingston 8
Lane, Uxbridge, Middlesex, UB8 3PH, UK 9
10
11
Corresponding author: 12
K. Longo13
Email: kevin.longo@brunel.ac.uk 14
15
16
Acknowledgments: 17
Mr K. Longo acknowledges the Guangxi Yuchai Machinery Company for 18
supporting his PhD study supervised by Prof. Hua Zhao and Dr Xinyan Wang 19
at Brunel University London. 20
21
22
Copyright © 2023 I Mech E. The definitive version was published in Longo, K., Wang, X. and Zhao, H. (2023) 'Impact of diesel-hythane dual-fuel
combustion on engine performance and emissions in a heavy-duty engine at low-load condition', International Journal of Engine Research, 0 (ahead-
of-print), pp. 1 - 17. doi: 10.1177/14680874231170651.. This is the author’s version of the work. It is posted here by permission of SAGE Publications
on behalf of Institution of Mechanical Engineers for personal use, not for redistribution (see: https://sagepub.com/journals-permissions).
Abstract 23
Heavy-duty diesel vehicles are currently a significant part of the transportation sector, as well as one of the 24
major sources of carbon dioxide (CO
2
) emissions. International commitments to reduce greenhouse gas (GHG) 25
emissions, particularly CO
2
and methane (CH
4
) highlight the need to diversify towards cleaner and more 26
sustainable fuels. Hythane, a 20% hydrogen and 80% methane mixture, can be a potential solution to this 27
problem in the near future. This research was focused on an experimental evaluation of partially replacing diesel 28
with hythane fuel in a single-cylinder 2.0 litre heavy-duty diesel engine operating in the diesel-gas dual fuel 29
combustion mode. The study investigated different gas substitution fractions (0%, 38% and 76%) of hythane 30
provided by port fuel injections at 0.6 MPa indicated mean effective pressure (IMEP) and a fixed engine speed 31
of 1200 rpm. Various engine control strategies, such as diesel injection timing optimisation, intake air pressure 32
and exhaust gas recirculation (EGR) were investigated in order to optimise the dual-fuel combustion mode. The 33
results indicated that by using hythane energy fraction (HEF) of 76% combined with 125 kPa intake air boost 34
and 25% EGR dilution, CO
2
emissions could be decreased by up to 23%, while indicated thermal efficiency 35
(ITE) was compromised by 1.5 percentage points, equivalent to a 3% reduction. Furthermore, soot was 36
maintained below Euro VI limit and nitrogen oxides (NOx) level was held below the Euro VI regulation limit of 37
8.5 g/kWh assuming a NOx conversion efficiency of 95% in a selective catalyst reduction (SCR) system. 38
Nevertheless, carbon monoxide (CO), unburned hydrocarbon (HC) and methane slip levels were considerably 39
higher, compared to the diesel-only baseline. The use of a pre-injection prior to the diesel main injection was 40
essential to control the heat release and pressure rise rates under such conditions. 41
1. Introduction42
Transportation energy demands account for approximately 20% of global energy consumption and are43
anticipated to rise by 25% between 2019 and 2050. This is due to an expected increase in the number of 44
vehicles, in particular heavy-duty (HD) vehicles as a result of economic growth [1]. 45
According to the Intergovernmental Panel on Climate Change (IPCC) [2], the combustion of fossil fuels is 46
a major contributor to the global warming by releasing substantial concentration of GHG, such as carbon dioxide 47
(CO
2
) into the atmosphere. In 2017, HD vehicles were responsible for about 6% of the CO
2
emissions in 48
European Union (EU) [3]. Therefore, this increasing concern about CO
2
has prompted the implementation of 49
new regulations to limit the CO
2
generation in the transportation sector. 50
Currently, the criterion for the evaluation of internal combustion (IC) engines is their tailpipe emissions [4]. 51
Thereby, a conventional diesel combustion (CDC) engines will thus no longer be able to meet the upcoming 52
strict emission regulations, requiring the employment of new technologies and alternative low and zero carbon 53
fuels. At present, the most intensive research is being conducted on two possibilities. The first is an attempt to 54
completely eliminate the use of fossil fuels in internal combustion engines (ICE), while the second is to burn 55
more efficiently with particular attention to exhaust emissions. The latter has been the most common approach 56
in recent years and has contributed to the substantial reduction in pollutant emissions. 57
Co-combustion of fuels with different properties, often known as dual-fuel (DF) combustion, are capable of 58
reducing both pollutants and CO
2
emissions when a low carbon fuel is used [5], however, this technology has 59
limited engine operation map, mainly at lower and higher loads due to incomplete or knocking combustion [6]. 60
In particular, diesel-natural gas dual-fuel compression ignition (CI) combustion has been demonstrated as an 61
effective solution for HD applications thanks to their simplicity of adaptation to existing ICEs [7]. Compressed 62
natural gas or bio-gas can be fed through a port fuel injection (PFI) system in a dual-fuel CI engine to provide a 63
lean and homogeneous distribution of the low reactivity fuel in the combustion chamber, resulting in multiple 64
ignition spots [8]. When compared to a diesel-only operation, this method allows for reduced local fuel-air 65
equivalence ratios and combustion temperatures, resulting in lower soot and NOx formation [9]. Another reason 66
for the simultaneous decrease in soot and NOx suggested by Iorio et al. [10] was this combustion mode has a 67
low flame temperature due to a higher ratio of heat capacity of CH
4
. 68
According to Stettler et al. [11], when compared to diesel-only vehicles, lean-burn compressed natural gas 69
(CNG) dual-fuel vehicles reduced CO
2
emissions by up to 9%. This conclusion was obtained after studying the 70
energy consumption, greenhouse gas emissions, and pollutants produced by five aftermarket dual-fuel engine 71
configurations in two vehicle platforms. 72
In fact, both the diesel injection timing and the properties of low reactivity fuel have a significant impact on 73
DF combustion operation, affecting both engine performance and exhaust emissions [12, 13]. With increasing 74
diesel injection advance, NOx increased while carbon monoxide (CO) and soot emissions were reduced [12]. 75
Moreover, Pedrozo et al. [14] concluded that the combination of reactivity-controlled compression ignition 76
(RCCI) and late intake valve closing (LIVC) can reduce unburned methane and also NOx emissions up to 80% 77
in a diesel-CNG combustion, 78
Though, due to the properties of methane (CH
4
), generally the main compound of natural gas, diesel-CNG 79
dual-fuel combustion has some drawbacks, such as slower flame propagation, which results in longer 80
combustion duration and, as a result, lower efficiency [15]. Also, this combustion mode is accompanied by 81
unburned CH
4
emission, also known as methane slip [14]. CH
4
is a GHG with 27 times global warming potential 82
(GWP) of the CO
2
emission over a 100-year lifetime [16]. Furthermore, the combination of natural gas and diesel 83
enable the DF technology to achieve similar thermal efficiency to that of the conventional diesel engines only at 84
high loads, as reported in [17, 18]. 85
When produced from renewable sources, hydrogen, on the other hand, has no carbon and is a clean and 86
environmentally friendly fuel [19]. Nonetheless, the use of pure hydrogen as a fuel in a DF engine, which 87
provides increased efficiency with respect to the CDC mode, it demonstrates certain limitations on the input 88
energy fraction, due to the problem of pre-ignition and backfire occurring before the diesel fuel injection. 89
Likewise, hydrogen is also associated with other undesirable effect, such as engine knocking owning to its 90
intensity, as reported in [20]. 91
By that, the usage of hydrogen blended with methane, commonly known as hythane, has the potential to 92
mitigate the problems associated with separate methane and hydrogen combustion [15, 21]. The higher 93
reactivity of the hydrogen improves combustion stability, resulting is faster and more complete combustion of 94
methane, and lower unburned CH
4
[22, 21]. Graham et al. [23] indicated that hythane can provide a 10%-20% 95
decrease in GHG levels, namely CO
2
emissions at the tailpipe when compared with diesel. However, this 96
reduction is only relevant when the hydrogen is produced from renewable sources [21]. Therefore, hythane with 97
hydrogen content up to 20% by volume can be deployed with existing CNG infrastructure and on-board gas 98
supply system without significant modification, effectively reducing CO
2
emissions at quite moderate financial 99
costs [21, 24]. 100
De Simio et al. [6] investigated a wide variety of diesel injection timings for diesel DF operation with natural 101
gas and hythane mixtures containing up to 25% hydrogen by volume in a four-cylinder CI light-duty engine at 102
low and medium loads. Although the highest CO
2
reduction and brake thermal efficiency (BTE) combination has 103
been evidenced at very advanced diesel start of injection (SOI) for 72% of hythane energy fraction (HEF) with 104
15% hydrogen by volume at low engine load, when RCCI DF is deployed, it is possible to reach 20-25% CO
2
105
reduction at the cost of a roughly 23% drop in BTE for later diesel SOIs (conventional DF). Conversely, 106
equivalent CO
2
, which combines CO
2
, CH
4
and non-methane HC emissions, has increased significantly when 107
compared to CDC. 108
Because of the higher flame temperature of hydrogen, NOx concentration increases with higher hydrogen 109
addition, whereas CO and HC levels decrease [25, 26]. Nevertheless, Talibi et al. [15] has noted a different 110
trend by investigating the effect of hythane enrichment with diesel pilot injection in a single-cylinder light-duty 111
CI engine. It was found that CO and HC were significantly higher while employing diesel-hythane dual-fuel 112
(DHDF) mode. Furthermore, a considerable reduction of particulate matter (PM) emissions was achieved 113
compared to CDC. Tutak et al. [27] tested various compositions of hydrogen and CNG in a single-cylinder light-114
duty diesel engine and concluded that the addition of hydrogen accelerated combustion, shortening the duration 115
of the combustion event. Additionally, it was also found that higher hydrogen and CNG fractions resulted in an 116
increase in peak pressure and temperature as well as higher NOx emissions. 117
The use of EGR has been proven as an effective method to extend DF operation. This is associated with 118
a reduction in combustion temperature as a result of the increased specific heat capacity and dilution level of 119
the in-cylinder charge [28, 29]. This delays the ignition time of the premixed fuel and hence allows to decrease 120
the levels of PRR and NOx emissions during dual-fuel operation [30]. Moreover, flame stability improves in the 121
presence of EGR at various air-fuel ratios [31, 32]. Nonetheless, Qian et al. [33] conducted a study on a 122
hydrogen-enriched diesel combustion and determined that increasing EGR levels reduced thermal efficiency at 123
all load engine settings. On the other hand, as the combustion temperature reduces as the air-fuel ratio 124
increases, combining hydrogen addition with higher air-fuel ratios, i.e. greater intake air pressures, can lead to 125
a decrease in NOx emissions. [26, 34]. 126
Abdelaal et al. [35] compared CDC and DF modes with and without EGR with 80% diesel replacement 127
(energy basis) at different engine loads in a single-cylinder light-duty natural gas diesel engine. When compared 128
to CDC, DF delivered a considerable reduction in CO
2
emissions at part loads, while thermal efficiency dropped 129
by roughly 13%. HC and CO levels, on the other hand, are higher in DF mode. With the inclusion of 20% EGR, 130
however, it was able to achieve similar thermal efficiency to diesel-only mode without significantly impacting 131
CO
2
levels. And, despite a decrease in HC and CO emissions, their values remained significantly higher than 132
the CDC. 133
The majority of previous works employing hythane fuel have mainly been focused on small- and light-duty 134
engines, with limited research on heavy-duty engines available. Moreover, studies with considerable high HEF 135
have indicated reasonable CO
2
reduction at the expense of a significant drop in thermal efficiency at part engine 136
loads. Therefore, the current study, which was conducted on a single-cylinder heavy-duty diesel engine with 137
port fuel injected hythane at an engine load of 0.6 MPa indicated mean effective pressure (IMEP), aims to 138
explore the CO
2
-ITE trade-off by using a HEF of up to 76%. Advanced engine and combustion control strategies, 139
such as late diesel injection, intake air pressure and EGR dilution were explored to identify the optimum 140
strategies for minimum GHG emissions of CO
2
and CH
4
without harming ITE and NOx emissions. The optimised 141
DHDF results were then compared to the conventional diesel only and a baseline diesel-hythane dual fuel 142
operations. 143
2. Experimental setup144
145
2.1 Engine setup and specifications 146
A schematic diagram of the single-cylinder compression ignition engine experimental setup is illustrated in 147
Figure 1. An eddy current dynamometer was used to absorb the power produced by the engine. An external 148
compressor supplied fresh intake air to the engine, which was controlled by a closed-loop system for boost 149
pressure. The intake manifold pressure was precisely controlled by a throttle valve positioned upstream of a 150
surge tank. A thermal mass flow metre was used to measure the air mass flow rate (m
air
). A water-cooled heat151
exchanger was used to regulate the temperature of the boosted air. To mitigate pressure oscillations, another 152
surge tank was installed in the exhaust manifold. The required exhaust manifold pressure was set using an 153
electrically controlled backpressure valve placed downstream of the exhaust surge tank. 154
155
Figure 1. Schematic diagram of the dual-fuel engine experimental setup. 156
Table 1 shows the HD engine hardware specifications. A 4-valve swirl-oriented cylinder head and a 157
stepped-lip piston bowl design constituted the combustion system. Separate electric motors controlled the 158
coolant and oil pumps. Throughout the experiments, the engine coolant and oil temperatures were set to 159
80°C, and the oil pressure was kept at 400 kPa. 160
Table 1. Single-cylinder HD engine specifications. 161
Parameter
Value
Bore/stroke
129/155 mm
Connecting rod length
256 mm
Displaced volume
2026 cm
3
Clearance volume
128 cm
3
Geometric compression ratio
16.8
Maximum in-cylinder pressure
18 MPa
Piston type
Stepped-lip bowl
Diesel injection system
Bosch common rail, injection pressure of 30-220
MPa, 8 holes, 150° spray
Hythane port fuel injection system
G-Volution controller and two Clean Air Power
injectors SP-010, injection pressure of 800kPa
162
Furthermore, the engine also included a prototype hydraulic lost-motion variable valve actuation (VVA) 163
system on the intake camshaft. This allows for the intake valve closing (IVC) to be adjusted, enabling for a 164
decrease in the effective compression ratio (ECR). This reduces compression pressures and temperatures, as 165
well as the mass trapped in the cylinder at a given boost pressure. 166
However, in order to simplify the experimental investigation, intake valve timings were kept constant at
167
baseline values throughout the experiments, with its intake valve opening (IVO) at -330 ± 1 crank angle degrees
168
(CAD) and IVC at -187 ± 1 CAD.
169
2.2 Fuel supply and proprieties
170
In this study, hythane gas, supplied by British Oxygen Company (BOC) Ltd, was employed as the premixed
171
fuel of the dual-fuel combustion and it is composed of 80% methane and 20% hydrogen gas mixture (molar).
172
Hythane gas was stored in a rack of six interconnected 20 MPa bottles outside of the engine test cell.
173
Specially developed hoses for the conveyance of CNG have been used, as they are constructed of a conductive
174
nylon core designed to dissipate static build-up. From there, Hythane was fed into a pair of pneumatically
175
controlled safety valves, a high-pressure filter and a high-pressure regulator that dropped the gas pressure to
176
1 MPa. The pressure regulator was kept constant by the hot engine coolant to counteract the reduction in
177
temperature experienced by the gas during expansion.
178
After flowing through the high-pressure regulator, hythane was fed into the test cell into an Endress +
179
Hauser Promass 80A Coriolis flow meter. After this mass flow meter, a low-pressure filter, a purge/pressure
180
regulator that adjusted the final hythane pressure to 0.8 MPa, and an emergency shut-off valve were connected,
181
before a flex hose connected the gas stream to the injector block. The injector block, designed for NG
182
application, was installed upstream of the intake surge tank to facilitate the mixing of the fuel gas with the intake
183
air. An injector driver controls the pulse width of the gas injectors and allowed the engine to run at different HEF
184
by altering the hythane mass flow rate (m
hythane
).
185
The high-pressure common rail diesel injection system, which can provide up to three injections per cycle,
186
was controlled by a dedicated engine control unit (ECU). The diesel mass flow rate (m
diesel
) was determined
187
using two Endress + Hauser Promass 83A Coriolis flow meters by measuring the total fuel supplied to and from
188
the diesel high-pressure pump and injector.
189
During the dual-fuel operation, the bulk fuel mass of port fuel injected hythane was ignited by direct injected
190
diesel. Table 2 lists the key properties of the diesel and hythane utilised in this experiment.
191
Table 2. Fuel proprieties of diesel and hythane.
192
Property
Unit
Diesel
Hythane
General proprieties
Lower heating value (LHV)
MJ/kg
42.9
52.1
Stoichiometric air-fuel ratio (AFR)
-
14.5
17.7
Gas density
kg/m
3
-
0.562
Cetane number
-
> 45
< 5
Liquid density (101.325 kPa, 20°C)
kg/dm
3
0.827
-
Normalised fuel’s molar mass
g/mol
13.9
16.5
Normalised molecular composition
-
CH
1.825
O
0.0014
CH
4.492
Gas composition (mole fraction)
Methane (CH
4
)
%
-
80.0
Hydrogen (H
2
)
%
-
20.0
Fuel contents (mass fraction)
Carbon (%C
fuel
)
%
86.6
72.6
Hydrogen (%H
fuel
)
%
14.2
27.4
Oxygen (%O
fuel
)
%
0.2
0.0
Calculated carbon intensity
Mass of CO
2
emissions per mole of fuel
gCO
2
/mol
44.0
44.0
Mass of CO
2
emissions per mass of fuel
gCO
2
/g
3.2
2.7
Assuming the complete conversion of hydrocarbon fuel into CO
2
gCO
2
/MJ
73.9
51.1
Maximum theoretical CO
2
reduction considering a constant ITE
%
-
30.9
Estimated CO
2
reduction with a HEF = 76%
%
-
23.5
193
An important parameter for the dual-fuel operation is the HEF, which is given by the ratio of the energy
194
content of the hythane injected to the total fuel energy supplied to the engine. As show in Table 2, using a HEF
195
of 76% can minimise exhaust CO
2
emissions by approximately 24% when hydrocarbon fuel is completely
196
converted into CO
2
.
197
(1)
where: m
diesel
and m
hythane
the mass flow rate of diesel and hythane, respectively; LHV
diesel
and LHV
hythane
the
198
lower heating value of diesel and hythane, respectively.
199
2.3 Exhaust emissions measurements and analysis
200
An AVL 415SE smoke metre was used to measure the smoke number downstream of the exhaust back
201
pressure valve. The measurement was taken in filter smoke number (FSN). Other exhaust emissions, such as
202
CO
2
, CO, CH
4
, HC, and NOx, were monitored using a heated line on a Horiba MEXA-7170 DEGR emission
203
analyser located in the exhaust pipe before the exhaust back pressure valve. The concentration of these
204
gaseous emissions in the exhaust stream was measured in parts per million (ppm). All the exhaust gas
205
components were then converted to net indicated specific gas emissions in g/kWh, according to Regulation No.
206
49 of UN/ECE [36]. The following is an example of the CO
2
conversion calculation:
207
(2)
208
where: u
CO
2
the raw exhaust gas constant; [CO
2
] the concentration of CO
2
in ppm; m
exh
the total exhaust mass
209
flow rate; P
ind
the engine net indicated power calculated from the measured IMEP
210
The aforementioned regulation also required that NOx and CO emissions be converted to a wet basis by
211
using a raw exhaust gas correction factor that is dependent on the in-cylinder fuel mixture composition. In
212
addition, the measurement of the HC was performed on a wet basis by a heated flame ionisation detector (FID),
213
while CO and CO
2
were measured through a non-dispersive infrared absorption (NDIR). A chemiluminescence
214
detector (CLD) was used to quantify NOx emissions. In this study, the EGR rate was defined as the ratio of the
215
measured CO
2
concentration in the intake surge tank to the CO
2
concentration in the exhaust manifold.
216
2.4 Data acquisition and analysis
217
Two National Instruments data acquisition (DAQ) cards linked to a computer were used to acquire the
218
signals from the measurement devices. The crank angle resolution data was sent to a USB-6251 high-speed
219
DAQ card, which was synchronised with an optical encoder with 0.25 CAD resolution. The low-frequency engine
220
operation conditions were recorded using a USB-6210 low-speed DAQ card. An in-house designed DAQ
221
software and combustion analyser displayed this data in real time.
222
Temperatures and pressures at relevant points were measured using K-type thermocouples and pressure
223
gauges, respectively. Intake and exhaust manifold pressures were measured by two Kistler 4049A water-cooled
224
piezoresistive absolute pressure sensors coupled to Kistler 4622A amplifiers. The in-cylinder pressure was
225
measured by a Kistler 6125C piezoelectric pressure sensor coupled with an AVL FI Piezo charge amplifier.
226
The crank angle-based in-cylinder pressure traces were averaged over 200 consecutives cycles for each
227
operating point and used to calculate the IMEP. It was also used to obtain the apparent net heat release rate
228
(HRR), following Heywood’s equation [37]
229
(3)
230
where: p the in-cylinder pressure; V the in-cylinder volume;
γ
the ratio of specific heats; θ the CAD.
231
Due to the fact that the absolute value of the heat released is less essential in this study than the bulk
232
shape of the curve to crank angle, a constant
γ
of 1.33 was assumed throughout the engine cycle.
233
The mass fraction burned (MFB) was estimated by the ratio of the integral of the HRR to the maximum
234
cumulative heat release. Combustion phasing was determined by the crank angle of 50% (CA50) MFB.
235
Combustion duration was represented by the period between the crank angle of 10% (CA10) and 90% (CA90)
236
cumulative heat release.
237
The ignition delay was defined as the period between the start of diesel main injection (SOI_2) into the
238
combustion chamber and the start of combustion (SOC), which was set to 2% MFB. The average in-cylinder
239
pressure and resulting HRR were smoothed using a Savitzky-Golay filter, after the combustion characteristics
240
and ignition delay were estimated.
241
The pressure rise rate (PRR) was calculated as the average of the maximum pressure variations over 200
242
cycles of in-cylinder pressure versus crank angle. The coefficient of variation of IMEP (COV
IMEP
) was determined
243
using the set of IMEP values from the 200 sampled cycles of the test engine.
244
(4)
245
where: σ
IMEP
the standard deviation of IMEP; IMEP the mean of IMEP.
246
The mean in-cylinder gas temperature at any crank angle position was computed using the ideal gas law
247
[37].
248
The electric current signal sent from the ECU to the diesel injector solenoid was measured using a current
249
probe. The signal was corrected by adding the energising time delay that had previously been measured in a
250
constant volume chamber. The resulting diesel injector current signal allowed the diesel injections be
251
determined.
252
The indicated thermal efficiency was classified as the ratio of work done to the rate of fuel energy supplied
253
to the engine, as shown below:
254
(5)
255
where: P
ind
the engine net indicated power calculated from the measured IMEP.
256
Combustion efficiency calculations were based on the emissions products not fully oxidised during the
257
combustion process except soot as:
258
(6)
259
where: LHV
CO
is equivalent to 10.1 MJ/kg [37].
260
Combustion losses associated with HC emissions were thought to be caused entirely by unburned hythane
261
fuel. This is a conservative approach since the LHV
hythane
is higher than the LHV
diesel
.
262
Finally, the relative air-fuel ratio (λ) was determined as follows:
263
(7)
264
where: AFR
hythane
and AFR
diesel
the stoichiometric air-fuel ratio of hythane and diesel, respectively.
265
2.5 Instrumentation specifications
266
Finally, before conducting the experiments, all of the instruments utilised are tested and calibrated under
267
the same operating conditions as the actual tests in order to ensure measurement accuracy. Table 3
268
summarises all of the measurement instruments used during the experiments, as well as the measurement
269
range values and accuracy.
270
Table 3. Test cell measurement devices
271
Variable
Manufacturer
Device
Measurement
range
Linearity/Accurac
y
Speed
Froude Hofmann
AG 150 dynamometer
0-8000 rpm
±1 rpm
Torque
Froude Hofmann
AG 150 dynamometer
0-500 Nm
±0.25% of FS
Clock Signal
Encoder
Technology
EB58
0-25000 rpm
0.25 CAD
Diesel flow rate
(supply)
Endress+Hauser
Proline Promass 83A02
0-20 kg/h
±0.10% of
reading
Diesel flow rate
(return)
Endress+Hauser
Proline Promass 83A01
0-100 kg/h
±0.10% of
reading
Hythane flow rate
Endress+Hauser
Proline Promass 80A02
0-20 kg/h
±0.15% of
reading
Intake air mass flow
rate
Endress+Hauser
Proline T-mass 65F
0-910 kg/h
±1.5% of reading
In-cylinder pressure
Kistler
Piezoelectric pressure
sensor Type 6125C
0-30 MPa
≤ ±0.4% of FS
Intake and exhaust
pressures
Kistler
Piezoresistive pressure
sensor Type 4049A
0-1 MPa
≤ ±0.5% of FS
Oil pressure
GE
Pressure transducer
UNIK 5000
0-1 MPa
< ±0.2% of FS
Temperature
RS
Thermocouple K Type
233-1473 K
≤ ±2.5 K
Fuel injector current
signal
LEM
Current probe PR30
0-20 A
±2 mA
Smoke number
AVL
415SE
0-10 FSN
-
CO
Horiba
MEXA-7170-DEGR
(Non-Dispersive
Infrared Detector)
0-12 vol%
≤ ±1.0% of FS or
±2.0% of
readings
CO
2
Horiba
MEXA-7170-DEGR
(Non-Dispersive
Infrared Detector)
0-20 vol%
≤ ±1.0% of FS or
±2.0% of
readings
HC
Horiba
MEXA-7170-DEGR
(Heated Flame
Ionization Detector)
0-500 ppm or 0-
50k ppm
≤ ±1.0% of FS or
±2.0% of
readings
CH
4
Horiba
MEXA-7170-DEGR
(Non-Methane Cutter +
Heated Flame
Ionization Detector)
0-0.25k ppm or
0-25k ppm
≤ ±1.0% of FS or
±2.0% of
readings
NO/NOx
Horiba
MEXA-7170-DEGR
(Heated
Chemiluminescence
Detector)
0-500 ppm or 0-
10k ppm
≤ ±1.0% of FS or
±2.0% of
readings
EGR
Horiba
MEXA-7170-DEGR
(Non-Dispersive
Infrared Detector)
0-20 vol%
≤ ±1.0% of FS or
±2.0% of
readings
272
3. Test methodology
273
The experimental testing was carried out at a constant engine speed of 1200 rpm and a fixed load of 0.6
274
MPa IMEP, which is equivalent to 25% of the full engine load and, represents a high residency area in a typical
275
HD vehicle drive cycle, such as WHSC, and indicated in Figure 2.
276
277
Figure 2. The selected test point over the experimental HD engine speed-load map.
278
Table 4 summarises the engine test conditions for the CDC, baseline DHDF and optimised DHDF operation
279
modes. The first part of the experiments comprised a comparison on engine emissions and performance
280
between the two aforementioned combustion modes by varying the HEF. This comparison was carried out using
281
a constant baseline late diesel injection. Both COV
IMEP
and PRR were used to define the HEF limit, which was
282
approximately 76%, resulting in an overall combustion mixture of 24% diesel, 61% methane, and 15% hydrogen.
283
Also, the intake and exhaust air pressure set-points from a Euro V compliant multi-cylinder HD diesel engine
284
were used in order to provide a sensible starting point.
285
Other experiments were carried out to obtain the engine calibration for optimised DHDF combustion mode
286
with the highest HEF. This optimisation included the sweep of several engine control parameters, namely diesel
287
injections timing, intake air pressure (P
int
), and EGR rate. As a result, an optimal point was reached that achieved
288
with the best trade-off between the GHG emissions (CO
2
and CH
4
) and the ITE whilst keeping the engine-out
289
NOx of less than 8.5 g/kWh. This NOx level was necessary in order to achieve a Euro VI emissions compliance
290
with a NOx conversion of approximately 95% in the SCR system.
291
Throughout the experiments, exhaust pressures were adjusted to provide a constant pressure differential
292
of 10 kPa above the intake air pressure to achieve a fair comparison with equivalent pumping work and to
293
realise the required EGR rate. Intake air temperature was maintained constant at 35°C during all the
294
experiments by using an air-to-water cooler and intake air heater. A diesel pre-injection (SOI_1) with an
295
estimated volume of 3 mm
3
and a constant delay time of 1.1ms (7.92 CAD at 1200 rpm) before SOI_2 was
296
employed to reduce the levels of PRR. Moreover, the diesel main injection timings were optimised to achieve
297
the highest ITE in DHDF combustion mode. However, it is worth noting that during this optimisation, the hythane
298
supply was maintained constant while the diesel was automatically adjusted by the ECU in order to achieve the
299
same IMEP, resulting in a slightly HEF variation (around 4%). The limits of the highest average in-cylinder
300
pressure (P
max
) and the maximum PRR were set to 18 MPa and 2.0 MPa/CAD, respectively. Finally, the COV
IMEP
301
of 3% limit was used to determine stable engine operation.
302
Table 4. Engine testing conditions.
303
Parameter
Unit
CDC
Baseline DHDF
Optimised DHDF
Engine load (IMEP)
MPa
0.6
0.6
0.6
Engine speed
rpm
1200
1200
1200
Diesel injection strategy
-
Pre- and main
injection
Pre- and main
injection
Pre- and main
injection
Diesel SOI_2
CAD ATDC
-5
-5
Sweep
Diesel injection pressure
MPa
100
100
100
Intake air pressure (P
int
)
kPa
125
125
Sweep
Exhaust air pressure
kPa
135
135
Sweep
Intake air temperature
°C
35 ± 1
35 ± 1
35 ± 1
ECR
-
16.8
16.8
16.8
HEF
%
0
Sweep
~76
EGR
%
0
0
Sweep
304
Regarding the control of GHG and pollutant emissions from DF combustion engines, Regulation No. 49 of
305
the United Nations Economic Commission for Europe (UN/ECE) [36] enhances the Euro VI emissions standards
306
for on-road HD vehicles by establishing five different types of dual-fuel engines. For the sake of clarity, this
307
study will focus on the evaluation of Type 2B heavy-duty dual-fuel (HDDF) engines. These operate in the hot
308
section of the World Harmonised Transient Driving Cycle (WHTC), with an average gas energy fraction
309
(GEF
WHTC
) ranging from 10% to 90%, while still enabling for diesel-only engine operation.
310
The Euro VI emissions standards for Type 2B HDDF engines are shown in Table 5 for both the stationary
311
(WHSC) and transient (WHTC) test cycles. It is worth noting that, with the exception of the HEF experiment, all
312
optimised DHDF experiments used the highest HEF with the goal of maximising hythane utilisation, which
313
contributed to achieve a GEF
WHTC
of more than 68%.
314
315
Table 5. Euro VI emissions limits for Type 2B heavy-duty dual-fuel engines
316
Emission
Unit
WHSC
WHTC (GEF%
WHTC
> 68%)
Nitrogen oxides (NOx)
g/kWh
0.40
0.46
Carbon monoxide (CO)
g/kWh
1.50
4.00
Particulate matter (PM)
g/kWh
0.01
0.01
Total unburned hydrocarbon (HC)
g/kWh
0.13
-
Methane (CH
4
)
g/kWh
-
0.50
4. Results and discussion
317
The results and discussion section examines the impact of hythane addition at a baseline DHDF for various
318
substitution ratios, as well as the optimisation of the DHDF mode for the highest diesel percentage replacement,
319
which includes diesel injection timing, intake air pressure, and EGR rate sweeps. A comparison of CDC,
320
baseline, and optimised DHDF operations is discussed at the end of this section.
321
4.1 The impact of HEF
322
In this study, a baseline diesel main injection at -5 CAD ATDC (after top dead centre) with a small diesel
323
pre-injection to attenuate COV
IMEP
and PRR were employed for different HEF, varying from 0% (diesel-only) to
324
a maximum value of 76%. Because of the exponential growth of PRR, which caused strong knocking, unstable
325
combustion (high COV
IMEP
) was observed for HEF higher than 76%. Additionally, this experiment was performed
326
without EGR and with a constant intake air pressure of 125 kPa.
327
Table 6 shows the engine performance, combustion characteristics and indicated specific exhaust
328
emissions whereas Figure 3 depicts the in-cylinder pressure, mean in-cylinder gas temperature, HRR and MFB
329
traces, for CDC and DHDF operations. As seen in Table 6, increasing the HEF resulted in a 15% reduction in
330
CO
2
emissions for a HEF of 76%. This was expected of the addition of hydrogen into the combustion, because
331
the low reactivity port injected fuel has a lower carbon composition (lower C to H ratio) than diesel, as shown in
332
Table 2. Nonetheless, methane slip rose dramatically as HEF increased. This was mainly attributed to the two
333
following reasons. First, hythane is mainly composed by methane, resulting in increased unburned CH
4
levels
334
in the exhaust pipe from the crevices. Second, the inclusion of hythane resulted in a longer ignition delay,
335
in other words, a later SOC, due to the fact that the premixed charge has a lower cetane number comparing to
336
CDC. This aspect, combined with the slower flame propagation speed of methane that results in incomplete
337
and longer combustion duration (CA10-CA90) [7], and a lower and longer HRR peak (Figure 3), resulting in
338
an increase in unburned CH
4
and HC, and as a consequence, a reduction of combustion efficiency [15]. The
339
slower combustion rate can be seen in the MFB trace, which is also shown in Figure 3, with a clear delay of
340
CA50. This lower combustion efficiency had a direct impact on the loss in ITE of roughly 5 percentage points at
341
76% HEF. In addition, the increase of CO formation for higher rates of diesel replacement is explained by the
342
unburned fuel generated from incomplete combustion, which led to lower mean in-cylinder gas temperature.
343
Moreover, a minor increase in NOx was seen with increasing HEF percentage. This is explained in part by
344
the presence of hydrogen, which has a higher flame temperature, resulting in a higher peak in-cylinder gas
345
temperature, as shown in Figure 3. As the result, DHDF produced higher exhaust temperature. Specifically, the
346
DHDF operation with 76% HEF yielded a higher exhaust gas temperature (EGT) by about 32°C higher than that
347
measured for CDC. This level of temperature is more favourable for the methane oxidation catalyst (MOC) used
348
in DF engines, since the device typically requires an EGT of more than 400°C for high CH
4
conversion efficiency,
349
and hence a reduction in methane slip [38, 39]. Furthermore, at the maximum HEF, soot emissions were
350
slightly reduced, as shown in Table 6. This is likely because diesel fuel contributed for only 24% of total energy
351
supplied to the engine, resulting in lower local fuel-air equivalence ratios [9].
352
In terms of the combustion process, Figure 3 indicates that increasing the HEF resulted in a decrease in
353
the in-cylinder pressure. This can be explained by the slower propagation speed of methane [7], the major
354
compound in the mixture. However, it was observed in Figure 3 that the peak of HRR in DHDF was earlier than
355
that in CDC. And on this event, the addition of hydrogen can possibly increase the reactivity of the fuel mixture,
356
leading to earlier peak of the heat release rate. In addition, it can be seen that there was a small heat release
357
of the pre-injected diesel (SOI_1) before SOI_2, which was visible only in the DF combustion mode. This can
358
be further explained by the increased reactivity of the fuel mixture by adding hydrogen.
359
Table 6. The impact of HEF on low engine load operation.
360
Parameter
Unit
HEF = 0%
HEF = ~38%
HEF = ~76%
SOI_2
CAD ATDC
-5
-5
-5
COV
IMEP
%
2.07
2.37
2.54
PRR
MPa/CAD
0.55
0.56
0.44
P
max
MPa
7.54
7.38
6.85
EGT
°C
359
385
391
SOC-SOI_2
CAD
6.4
6.8
7.2
SOC
CAD ATDC
0.9
1.3
1.7
CA50
CAD ATDC
9.1
9.2
11.4
CA10-CA90
CAD
21.1
24.6
25.2
λ
-
2.60
2.39
2.22
ITE
%
44.2
41.0
39.7
Combustion Efficiency
%
99.5
95.4
92.9
ISCO
2
g/kWh
666
621
566
ISCH
4
g/kWh
0.0
7.6
12.0
ISNOx
g/kWh
7.7
8.6
8.9
ISsoot
g/kWh
0.0169
0.0193
0.0152
ISCO
g/kWh
1.2
7.7
9.0
ISHC
g/kWh
0.7
7.0
11.2
361
362
Figure 3. In-cylinder pressure, mean in-cylinder gas temperature, HHR and MFB for low engine load
363
operation with various HEF.
364
4.2 The effect of SOI_2
365
In this study, diesel injection timing was investigated in order to analyse its influence on exhaust emissions
366
and engine performance with 76% HEF. Diesel pre- and main injections were used in a DHDF engine. The
367
experiment was performed without EGR and with a constant intake air pressure of 125 kPa.
368
Figure 4 show indicated specific exhaust emissions, engine performance and combustion characteristics
369
for different HEF respectively, while the in-cylinder pressure, mean in-cylinder gas temperature, HRR and MFB
370
traces of 3 different SOI_2 at approximately 76% HEF were depicted in Figure 5.
371
Although CO
2
emissions decreased with more advanced SOI 2, which can be explained in part by a shorter
372
combustion period near top dead centre (TDC), the main reason was the lower diesel consumption. This smaller
373
ISFC
diesel
, as seen in Figure 4, can be explained by the ECU's automatic diesel amount adjustment to maintain
374
IMEP constant, since the hythane supply was held constant during the diesel injection sweep, resulting in a
375
slight HEF variation. This increase in diesel amount at late injection timings, on the other hand, contributed to
376
higher combustion efficiency by enhancing the combustion process. Besides, more advanced timings improved
377
the homogeneity of the in-cylinder charge, leading in lower CO and soot levels [12]. By using more advanced
378
SOI_2, both pressure and temperature were significantly increased as shown in Figure 5, which increased NOx
379
emissions but also improved reduced unburned fuel (HC and CH
4
) at the end of combustion, and hence
380
improving combustion efficiency.
381
Delaying the diesel injection, on the other hand, retarded the combustion phasing, resulting in a longer
382
CA10-CA90. As a result, both the ITE and the in-cylinder pressure decreased. However, it is noted that the
383
peak thermal efficiency was obtained at intermediate injection timing, due to optimised combustion phasing as
384
indicated by the values of CA50. As a conclusion, more advanced SOI_2 demonstrated lower carbon emissions
385
and higher engine performance, being -11 CAD ATDC the best timing to optimal trade-off between indicated
386
thermal efficiency and carbon emissions. It allowed for a reduction in CO
2
of 44.6 g/kWh, corresponding to an
387
8% drop, and a reduction in CH
4
of 0.3 g/kWh, equivalent to a 3% reduction. The ITE was also increased by
388
roughly 2 percentage points. Likewise, at this SOI_2 timing, soot emissions were reduced by about 55%,
389
maintaining them below Euro VI limits. Despite this, EGT dropped as SOI_2 advanced, moving away from the
390
optimal temperature of the MOC in order to achieve high CH
4
conversion efficiency.
391
392
(a)
393
394
(b)
395
396
(c)
397
Figure 4. Effect of diesel SOI_2 at low engine load DHDF operation on: (a) engine performance, (b) net
398
indicated specific exhaust emissions and (c) combustion characteristics.
399
400
Figure 5. In-cylinder pressure, mean in-cylinder gas temperature, HHR and MFB for low engine load DHDF
401
operation with various diesel SOI_2 at 76% HEF.
402
4.3 The effect of intake air pressure
403
Following the studies of DHDF with different injection timings, intake air pressure was swept for 3 different
404
pressures at 76% HEF: 125 kPa, 135 kPa and 145 kPa. EGR was not used in this experiment and diesel
405
injection timing was kept constant at -11 CAD ATDC, which corresponded to the optimised timing achieved in
406
the previous experiment.
407
The combustion characteristics, performance and exhaust emissions results for the intake pressure sweep
408
are summarised in Table 7, whereas Figure 6 depicts the in-cylinder pressure, mean in-cylinder gas
409
temperature, HRR and MFB traces of this experiment.
410
Table 7. The effect of P
int
on low engine load DHDF operation.
411
Parameter
Unit
P
int
= 125 kPa
P
int
= 135 kPa
P
int
= 145 kPa
HEF
%
76
76
76
SOI_2
CAD ATDC
-11
-11
-11
COV
IMEP
%
2.33
3.12
2.35
PRR
MPa/CAD
0.73
0.78
0.62
P
max
MPa
8.39
8.80
9.10
EGT
°C
363
341
326
SOC-SOI_2
CAD
6.5
6.3
6.1
SOC
CAD ATDC
-5.0
-5.2
-5.4
CA50
CAD ATDC
4.8
4.8
5.0
CA10-CA90
CAD
19.2
20.6
21.4
λ
-
2.29
2.50
2.68
ITE
%
41.0
40.5
39.7
Combustion Efficiency
%
93.2
91.8
90.9
ISFC
diesel
g/kWh
52.5
54.6
57.8
ISFC
hythane
g/kWh
127.5
127.9
128.8
ISCO
2
g/kWh
517
519
530
ISCH
4
g/kWh
11.3
14.0
15.6
ISNOx
g/kWh
14.9
14.6
14.4
ISsoot
g/kWh
0.0071
0.0118
0.0086
ISCO
g/kWh
5.9
7.4
9.0
ISHC
g/kWh
11.0
13.5
15.1
Higher intake air pressures allowed for more air dilution of the charge in the combustion chamber, resulting
412
in a leaner and lower reactivity mixture (higher λ). This, however, resulted in poor ignition and more incomplete
413
combustion, leading to a longer CA10-CA90 and thus more unburned fuel (HC and CH
4
). This resulted in a drop
414
in combustion efficiency as well as a 1.3 percentage point loss in ITE for the highest P
int
, as shown in Table 7.
415
Albeit the decreased amount of burned fuel led in a slightly decrease in CO
2
ppm, ISCO
2
increased when P
int
416
was increased due to lower ITE. On the other hand, CO also suffered an increase with higher P
int
. One possible
417
reason is that incomplete combustion (longer CA10-CA90) generates more CO because CO does not have
418
enough time to oxidise and form CO
2
[40]. Another reason can be the lower in-cylinder combustion temperatures
419
noticed for higher P
int
due to higher relative air-fuel ratios, since CO formation is also function of mixture
420
temperatures [35, 6]. However, the higher air dilution of the charge for higher intake air pressures increased the
421
heat capacity ratio, allowing the peak in-cylinder gas temperature to be reduced, as shown in Figure 6, resulting
422
in lower NOx formation [26, 34].
423
Additionally, the longer combustion process is believed to be responsible for the ISFC
diesel
increase of
424
around 4% and 10% for P
int
of 135 kPa and 145 kPa, respectively. It is noted that the intake pressure of 125
425
kPa provided the best compromised between performance and carbon emissions.
426
427
Figure 6. In-cylinder pressure, mean in-cylinder gas temperature, HHR and MFB for low engine load DHDF
428
operation with various Pint.
429
4.4 The effect of EGR
430
The last approach used in this study to optimise DHDF for the highest HEF operation was the sweep of
431
EGR rate up to 30%, as shown in Table 8. SOI_2 and P
int
were kept constant at -11 CAD ATDC and 125 kPa,
432
respectively, which corresponded to the optimised values achieved in the previous experiments. The
433
combustion characteristics, performance and exhaust emissions results for EGR rate sweep are summarised
434
in Table 8, while Figure 7 depicts the in-cylinder pressure, mean in-cylinder gas temperature, HRR and MFB
435
traces of this experiment.
436
The increase in EGR rate produced lower oxygen concentration and higher heat capacity in the in-cylinder
437
charge, resulting in a slightly longer ignition delay. The longer ignition delay, on the other hand, resulted in a
438
more homogeneous in-cylinder charge, resulting in a higher first HRR peak, as shown in Fig. 9. In addition, the
439
utilisation of EGR extended the combustion duration. As a result, CA50 was delayed, indicating that there was
440
room to optimise SOI_2 for more advanced timing when EGR was employed [41].
441
The increased in-cylinder temperature, as shown in Figure 7, contributed to a little reduction in CO and HC
442
emissions as well as methane slip, resulting in more complete combustion, in other words, higher combustion
443
efficiency. This is because with EGR, a portion of the unburned fuel (HC and CH
4
) is recirculated and reburned
444
in the mixture, due to the presence of a sufficient amount of oxygen in the combustion chamber [35]. As the
445
result, diesel and hythane ISFC will be lower, leading to an improvement in the ITE and minor CO
2
reduction.
446
However, at 30% EGR rate, a reverse effect was found, resulting in an increase in CO, HC, and CH
4
, while soot
447
emissions exceeded the Euro VI limit. This can be due to a lack of oxygen, resulting in poor combustion and
448
more unburned fuel. Therefore, the effectiveness of EGR to reduce HC, CO, and CH
4
emissions by reburning
449
some of the unburned fuel is dependent on the availability of oxygen in the combustion chamber [35].
450
The NOx emissions were dramatically reduced from 14.9 to 3.1 g/kWh with 30% EGR while the soot
451
emissions were slightly increased due to the reduction in the in-cylinder air-fuel ratio.
452
As a conclusion, it can be stated with a degree of confidence that EGR of 25% provided the best trade-off
453
between exhaust emissions and efficiency.
454
Table 8. The effect of EGR on low engine load DHDF operation.
455
Parameter
Unit
EGR =
0%
EGR =
10%
EGR =
20%
EGR =
25 %
EGR =
30%
HEF
%
76
76
76
76
76
SOI_2
CAD ATDC
-11
-11
-11
-11
-11
COV
IMEP
%
1.76
1.52
1.61
1.56
1.76
PRR
MPa/CAD
0.75
0.70
0.62
0.58
0.61
P
max
MPa
8.61
8.46
8.44
8.35
8.32
EGT
°C
361
363
367
368
369
SOC-SOI_2
CAD
6.4
6.5
7.2
7.4
7.5
SOC
CAD ATDC
-5.1
-5.0
-4.3
-4.1
-4.0
CA50
CAD ATDC
4.5
4.8
5.1
5.3
5.4
CA10-CA90
CAD
19.1
19.2
19.3
19.6
19.7
λ
-
2.42
2.12
1.95
1.87
1.78
ITE
%
41.1
41.5
42.4
42.7
42.8
Combustion Efficiency
%
93.2
94.0
94.1
94.4
94.3
ISFC
diesel
g/kWh
50.7
47.6
45.0
43.8
43.8
ISFC
hythane
g/kWh
121.2
121.1
121.0
120.4
120.1
ISCO
2
g/kWh
517.1
518.4
513.9
513.1
513.8
ISCH
4
g/kWh
10.9
9.9
9.0
8.4
8.6
ISNOx
g/kWh
14.9
10.4
6.4
4.3
3.1
ISsoot
g/kWh
0.0071
0.0081
0.0093
0.0098
0.0128
ISCO
g/kWh
6.0
5.6
4.9
4.8
4.9
ISHC
g/kWh
10.5
9.5
8.8
8.2
8.4
456
457
Figure 7. In-cylinder pressure, mean in-cylinder gas temperature, HHR and MFB for low engine load DHDF
458
operation with various EGR.
459
460
4.5 Comparison of different engine combustion modes
461
This section compares the three different combustion modes employed in this study to demonstrate the
462
impact of baseline DHDF and optimised DHDF on engine performance and exhaust emissions at low engine
463
load. Table 9 shows that 76% HEF in a baseline DHDF lowered CO
2
emissions by 15%. The addition of hythane,
464
on the other hand, reduced ITE while elevating methane slip, CO, and HC, which led to a 7% reduction in
465
combustion efficiency. Despite this, optimising DHDF combustion using advanced engine control strategies,
466
such as low booster pressure, diesel injection optimisation, and EGR dilution might mitigate the aforementioned
467
negative effects.
468
With this optimisation of DHDF, the CO
2
was reduced by 23% when compared to CDC, which is consistent
469
with the estimated CO
2
reduction provided in Table 2, as well as the CO
2
reduction achieved by the literature
470
review, while thermal efficiency was compromised by only 1.5 percentage points (approximately 3%) when
471
compared to conventional diesel combustion. This is an effective result for low engine load conditions when
472
compared to some literature review presented in the Introduction Section. Likewise, NOx emission and soot
473
emissions were reduced by 44% and 42%, respectively. On the other hand, since CH
4
emissions have increased
474
significantly, and taking into account that 1 g of methane in the exhaust gas is equivalent to 27 g of CO
2
over
475
100 years (IPCC Sixth Assessment Report) [16], methane has offset the carbon reduction provided by the
476
optimised DHDF, yielding 11% more equivalent CO
2
(overall GHG emissions) than the CDC mode. Despite the
477
overall GHG levels increase, it still represents an improvement over what De Simio et al. [6] reported. As a
478
result, methane slip control is essential to keep DHDF mode as a viable solution to reduce real (equivalent) CO
2
479
emissions from ICEs. MOC are commonly employed with DF engines to oxidise unburned CH
4
, although it may
480
be difficult to obtain high methane conversion efficiency at part engine loads due to its light-off temperature
481
(about 400°C), which is still roughly 30°C higher than the EGT achieved by the optimised DHDF regime. Hence,
482
additional optimisation, such as the LIVC strategy [14], is needed to meet the MOC temperature requirement.
483
The use of LIVC may also help to improve the flammability of the in-cylinder charge [14], which may result in
484
higher combustion temperatures and reduced HC and CO. Moreover, CO levels can be greatly reduced by
485
applying a simple oxidation catalyst in the exhaust line [6].
486
In summary, DHDF optimisation indicated an increase in combustion efficiency when compared to its
487
baseline DF, resulting in a more complete combustion. Despite the fact that CO, HC, and CH4 levels remain
488
high, this optimisation indicates a positive trend of reducing undesired engine-out emissions and shows that
489
there is still room for improvement, making this DF operation a possible viable solution for short-term
490
applications.
491
Table 9. Comparison of engine efficiencies and emission for three combustion modes
492
Parameter
Unit
CDC
Baseline DHDF
Optimised DHDF
ITE
%
44.2
39.7 (-10%)
42.7 (-3%)
Combustion Efficiency
%
99.5
92.9 (-7%)
94.4 (-5%)
ISCO
2
equivalent
g/kWh
666
890 (+34%)
740 (+11%)
ISCO
2
g/kWh
666
566 (-15%)
513 (-23%)
ISCH
4
g/kWh
0.0
12.0 (+1614%)
8.4 (+1100%)
ISNOx
g/kWh
7.7
8.9 (+16%)
4.3 (-44%)
ISsoot
g/kWh
0.0169
0.0152 (-10%)
0.0098 (-42%)
ISCO
g/kWh
1.2
9.0 (+650%)
4.8 (+300%)
ISHC
g/kWh
0.7
11.2 (+1500%)
8.2 (+1071%)
Conclusions
493
In this study, engine experiments were conducted to demonstrate the capability of advanced engine
494
combustion control strategy to enable significant increase in the replacement of diesel fuel with hythane at a
495
relatively low engine load in order to improve the CO
2
-thermal efficiency trade-off in heavy-duty engines. Testing
496
was carried out with port fuel injection of hythane, containing 20% hydrogen and 80% methane molar basis, on
497
a single-cylinder heavy-duty diesel engine operating at a constant engine speed of 1200 rpm and 0.6 MPa
498
IMEP, a typical part-load operating condition of 25% of total engine load. The hythane energy fraction (HEF)
499
was held at 76% ± 1% while dual-fuel combustion mode was optimised for the best trade-off between the lowest
500
CO
2
/CH
4
and the highest ITE possible, whilst keeping the NOx emission low. Engine control strategies, such as
501
intake air boosting, diesel injection strategy and EGR addition were explored to identify and achieve an
502
optimised diesel-hythane dual-fuel (DHDF) combustion operation. The main findings can be summarised as
503
follows:
504
1. The baseline DHDF combustion mode using 76% hythane energy fraction demonstrated a further reduction
505
in CO
2
emissions by 15% when compared to the CDC under the same combustion operating conditions.
506
This was due to the lower C to H ratio of hythane than diesel fuel, which was influenced by the mixture's
507
hydrogen content. However, this was accompanied with a 10% drop (5 percentage points) in the ITE as
508
well as an increase in CO and unburned HC and CH
4
due to incomplete combustion. Soot emissions, on
509
the other hand, were lowered by around 10% to remain within the Euro VI standard due to lower local fuel-
510
air equivalence ratios caused by the replacement of diesel fuel in the in-cylinder mixture.
511
2. More advanced diesel injection timings resulted in a considerable reduction in CO
2
emissions as well as
512
lower CO and soot levels due to a shorter combustion duration around TDC, which improved in-cylinder
513
mixture reactivity by promoting the fast burning rate of hydrogen. SOI_2 at -11 CAD ATDC provided the
514
best balance of ITE and carbon emissions. As a result, CO
2
emission was decreased by 44.6 g/kWh,
515
reflecting an 8% drop, and a reduction in methane slip of 0.3 g/kWh, equivalent to a 3% reduction.
516
3. Increase in the intake air pressure led to lower reactivity of the in-cylinder charge, causing poor ignition
517
and incomplete combustion, resulting in slightly higher CO and CO
2
levels and a substantial increase of
518
unburned HC and methane slip (from 11.3 to 15.6 g/kWh). Consequently, both combustion and indicated
519
thermal efficiencies fell by about 2.3 and 1.3 percentage points, respectively.
520
4. The introduction of 25% EGR significantly reduced the NOx emissions from 14.9 to 4.3 g/kWh due to a
521
reduction in combustion temperature. Also, EGR dilution enabled more complete combustion by reburning
522
unburned fuel, resulting in lower levels of CH
4
, HC, and CO, as well as an improvement in ITE.
523
5. The optimised DHDF operation at HEF of 76%, by appropriate diesel injection , lower intake air pressure,
524
and EGR addition, resulted in a CO
2
reduction of 23% when compared to CDC, though ITE was lowered
525
by 1.5 percentage points, corresponding to a 3% reduction. Overall GHG emissions (equivalent CO
2
)
526
increased by 11% due to the increase in methane slip.
527
Overall, this experimental study provides a better understanding of the impact of high HEF on performance
528
and all engine-out emissions of a diesel-hythane dual-fuel combustion at low engine load. It is shown that diesel-
529
hythane engine has the potential to contribute to a noticeable CO
2
reduction in the transportation sector if clean
530
energy is employed to produce the hydrogen content of hythane.
531
Additional studies on different engine speeds and loads are being carried out in order to verify the potential
532
impact of hythane at different engine operating conditions, and the RCCI mode and LIVC will also be
533
investigated to lower exhaust emissions.
534
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535
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536
Appendix
537
Notation
538
ATDC
After Top Dead Centre
BTE
Brake Thermal Efficiency
CA10
Crank Angle of 10% Cumulative Heat Release
CA50
Crank angle of 50% Cumulative Heat Release
CA90
Crank angle of 90% Cumulative Heat Release
CA10-CA90
Combustion Duration
CAD
Crank Angle Degrees
CDC
Conventional Diesel Combustion
CH
4
Methane
CI
Compression Ignition
CNG
Compressed Natural Gas
CO
Carbon Monoxide
CO
2
Carbon Dioxide
COV
IMEP
Coefficient of Variation of IMEP
DAQ
Data Acquisition
DF
Dual-Fuel
DHDF
Diesel-Hythane Dual-Fuel operation
ECR
Effective Compression Ratio
ECU
Engine Control Unit
EGR
Exhaust Gas Recirculation
EGT
Exhaust Gas Temperature
EU
European Union
GHG
Greenhouse Gases
GWP
Global Warming Potential
HC
Hydrocarbons
HD
Heavy-Duty
HEF
Hythane Energy Fraction
HRR
Heat Release Rate
ICE
Internal Combustion Engine
IMEP
Indicated Mean Effective Pressure
IPCC
Intergovernmental Panel on Climate Change
ISFC
Net Indicated Specific Fuel Consumption
ISCH
4
Net Indicated Specific Emissions of Methane
ISCO
Net Indicated Specific Emissions of Carbon monoxide
ISCO
2
Net Indicated Specific Emissions of Carbon dioxide
ISHC
Net Indicated Specific Emissions of unburned Hydrocarbon
ISNOx
Net Indicated Specific Emissions of Nitrogen Oxides
ISsoot
Net Indicated Specific Emissions of soot
ITE
Indicated Thermal Efficiency
IVC
Intake Valve Closing
IVO
Intake Valve Opening
LHV
Lower Heating Value
LIVC
Late Intake Valve Closing
m
Mass Flow Rate
MBF
Mass Fraction Burned
MOC
Methane Oxidation Catalyst
NOx
Nitrogen Oxides
PFI
Port Fuel Injection
P
int
Intake Air Pressure
P
max
Maximum Average In-cylinder Pressure
PM
Particulate Matter
PRR
Pressure Rise Rate
RCCI
Reactivity-Controlled Compression Ignition
SCR
Selective Catalyst Reduction
SOC
Start of Combustion
SOC-SOI_2
Ignition Delay
SOI_1
Start of Diesel pre-injection
SOI_2
Start of Diesel main injection
TDC
Top Dead Centre
WHSC
World Harmonised Stationary Cycle
WHTC
World Harmonised Transient Cycle
λ
Relative Air-Fuel Ratio
γ
Ratio of Specific Heats
539
