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Application of hydrogen enriched natural gas in spark ignition IC engines: from
fundamental fuel properties to engine performances and emissions
Fuwu Yana,b, Lei Xua,b, Yu Wanga,b,*
a. Hubei Key Laboratory of Advanced Technology for Automotive Components, School of Automotive Engineering,
Wuhan University of Technology, Wuhan 430070, P.R. China
b. Hubei Collaborative Innovation Center for Automotive Components Technology, Wuhan 430070, P.R. China
*Corresponding author:
Yu Wang
School of Automotive Engineering
Wuhan University of Technology
Wuhan, 430070, P.R. China
Email: yu.wang@whut.edu.cn
Prepared for submission to Renewable and Sustainable Energy Reviews
1
Application of hydrogen enriched natural gas in spark ignition IC engines: from
fundamental fuel properties to engine performances and emissions
Fuwu Yana,b, Lei Xua,b, Yu Wanga,b,*
a. Hubei Key Laboratory of Advanced Technology for Automotive Components, School of Automotive Engineering,
Wuhan University of Technology, Wuhan 430070, P.R. China
b. Hubei Collaborative Innovation Center for Automotive Components Technology, Wuhan 430070, P.R. China
Abstract
This review offers a summary of the recent research progress on hydrogen enriched compressed natural gas (HCNG)
engines. Spark ignition IC engines fueled by HCNG have been shown to be advantageous over traditional gasoline
engines in terms of fuel efficiencies and pollutant emissions. Further taking into account the vast availability of natural
gas and the renewability of hydrogen, HCNG is considered to be a promising alternative fuel for large-scale applications
in spark ignition engines. The past decade has witnessed significant progress in the research and development of HCNG
engines. In this work, we intend to provide a comprehensive survey of recent studies on the power, efficiency and
combustion & emission characteristics of HCNG engines. Additionally, we also focus on correlating fuel properties with
experimentally observed HCNG engine performances, providing fundamental insights in the effects of hydrogen
addition on engine working processes. To reach this goal, the physicochemical properties of hydrogen and its mixture
with natural gas were firstly analyzed, followed by the detailed presentation and analysis of notable experimental results
on HCNG engine performances. Numerical models for HCNG engines, which serve as an efficient way for engine
parameter optimization, were also discussed. Apart from HCNG, the usage of other hydrogen-enriched fuels, such as
hydrogen-gasoline, hydrogen-diesel, biogas and hydrogen-ethanol/methanol in IC engines were also briefly discussed.
Keywords: Hydrogen addition; Fuel properties; HCNG engine; Combustion characteristics; Emissions;
2
Content
• 1. Introduction
• 1.1 The origin of HCNG engines
• 1.2 Motivation and outline of this review
• 2. Fundamental fuel properties of hydrogen / natural gas mixtures
• 2.1 Lower heating value
• 2.2 Adiabatic flame temperature and auto-ignition temperature
• 2.3 Flammability limits
• 2.4 Minimum ignition energy
• 2.5 Quenching distance
• 2.6 Ignition delay
• 2.7 Laminar flame speed
• 3. HCNG engine performances
• 3.1 Power performances
• 3.2 Fuel economy
• 3.3 Emission performances
• 3.2.1 NOx emissions
• 3.2.2 Total HCs emissions
• 3.2.3 CO emissions
• 3.2.4 Particulates and other unregulated emissions
• 3.4 Cycle-by-cycle variations and lean burn limit
• 4. Numerical studies on HCNG engine performance and emissions
• 5. IC engines fueled with other hydrogen-enriched fuels
• 5.1 Hydrogen-gasoline mixture as fuel
• 5.2 Hydrogen-diesel mixture as a fuel
• 5.3 Biogas and oxygenated fuel
• 6. Concluding remarks
3
1. Introduction
1.1 The origin of HCNG engines
Our society is heavily relying on internal combustion (IC) engines for energy conversion, especially in the
transportation sector [1]. While bringing us great convenience in facilitating mobility, traditional engines also consume
tremendous amounts of non-renewable, petroleum-based liquid fuels, mostly in the form of gasoline and diesel. In
addition, the extraction of energy from burning petroleum fuel in engines comes at a cost of severe emissions of
pollutants, which include but not limited to unburned hydrocarbons (UHCs), nitric oxide (NOx), carbon monoxide (CO),
particulate matter (PM) and greenhouse gases. Indeed, the dwindling crude-oil reserve and the accompanied
environmental pollutions are the two most pressing issues needed to be addressed today.
Nomenclature
Tad
λ
XH2
Tmax
T
P
SL
R
Abbreviations
BTDC
BMEP
CA
CCV
COV
COVIMEP
CR
DI
EGR
ETC
HCNG
ICEs
IMEP
ITE
LFL
LOL
MBT
MAP
MIE
NEDC
NG
S.I.
TDC
UFL
WOT wide open throttle
adiabatic flame temperature
excess air ratios
hydrogen mole (volumetric) fraction in
fuel blends
maximum in-cylinder gas temperature
temperature
pressure
ignition delay times
laminar flame speed
equivalence ratio
gas constant
before top dead center
brake mean effective pressure
crank angle
cycle-by-cycle variations
coefficient of variation
COV in terms of indicated mean
effective pressure
Pressure
compression ratio
direct injection
exhaust gas recirculation
European Transient Cycle
hydrogen-enriched natural gas
internal combustion engines
indicated mean effective pressure
indicated thermal efficiency
lower flammability limit
lean operation limit
maximum brake torque
manifold absolute pressure
minimum ignition energy
New European Driving Cycle
natural gas
spark ignition
top dead center
upper flammability limit
wide open throttle
4
Many alternatives to IC engines have been proposed in the past decades, targeting at sustainable and cleaner energy
utilization. In the transportation sector, these mainly include hydrogen fuel cell, electrical battery and external
combustion engines. However, none of these alternative power sources have yet succeeded in challenging the dominant
role of IC engines. This is especially true in the field of heavy-duty, long-distance cruising commercial vehicles where
the concept of using power battery or electric hybrid do not offer obvious benefits [1]. On the other hand, the relatively
low cost, wide range of power outputs and high durability are expected to keep IC engines from being replaced on a
large scale in the foreseeable future. Nevertheless, it is very clear that innovative technologies that can further improve
engine efficiency and emission characteristics have to be developed and implemented as soon as possible, in order to
meet the ever-increasing challenges of both crude oil and environment crisis.
IC engines, by definition, are machines that combust fuels to achieve energy conversion and output. Therefore, it is
natural to deduce that fuel type is an essential factor affecting engine performance. Indeed, many researchers are seeking
the possibility of improving engine performances by using cleaner fuels. At present, most engines are fueled by gasoline
and diesel that are refined from crude oil. This is due primarily to the fact that gasoline and diesel, being in liquid phase
at room temperature, have a very high energy density of more than 40 kJ per gram of fuel. In addition, historical reasons
such as established refining processes and fuel delivery infrastructures also play a noticeable role. However, to address
the above-mentioned crude oil and environmental issues, we now may need to sacrifice part of these benefits and turn
to more environmental-friendly fuels to power IC engines. Natural gas [2], methanol [3], ethanol [4], dimethyl ether
(DME) [5], and biodiesel [6] are among the most investigated alternative fuels. Out of these options, natural gas (NG),
with its vast availability, low emission (detailed below) and compatibility with both spark ignition (SI) and compression
ignition (CI) engines [2], has already earned itself widespread applications in the transportation sector, as demonstrated
by the worldwide annual growth of natural gas vehicles (NGVs) shown in Fig. 1.
5
Fig. 1 Worldwide demand growth of NGVs. Adapted from Khan et al. [2]
Being the major component of NG, methane has the highest hydrogen to carbon ratio among all hydrocarbons,
resulting in NG engines to have low specific CO2, UHCs and CO emissions. By switching gasoline to NG, the UHCs
and CO emission could be reduced by 30-35% and 20-30%, respectively [7]. CO2 emission was also reported to be
lower than gasoline engine [8]. In addition, the fact that NG has a better anti-knock property makes it possible for spark
ignition (SI) NG engines to have a higher compression ratio and thus higher thermal efficiencies, as compared to their
gasoline counterparts. However, if operated at stoichiometric condition, the high NG flame temperature could lead to
elevated engine-out NOx emissions such that a three-way catalyst is required for NOx emission reduction to reach the
regulation limits. An alternative way is to use lean burn technology which can inhibit in-cylinder NOx formation by
combusting fuel at lower temperatures. Indeed, the EURO-V emission standard can be met satisfactorily by lean-burn
NG engines without the use of exhaust gas after-treatment devices [9]. To meet the stricter NOx emission limit set by
the EURO-VI standard, even leaner NG-air mixtures need to be used for a further reduction of flame temperature, if the
expensive De-NOx devices are to be avoided [9]. However, due to NG’s relatively slow flame speed, it is typically
impossible for traditional SI NG engines to run at such lean conditions without significantly compromising engine
efficiencies. The enrichment of NG with a fast-burning fuel, i.e. hydrogen, which has a laminar burning velocity seven
6
times higher than NG, has been shown to be an effective method to extend the lean operation limit (LOL) of NG engines.
Hydrogen, characterized by its fast flame speed, wide flammability limit and low minimum ignition energy (MIE),
is an ideal fuel additive to conventional hydrocarbon fuels (including NG) to enhance their lean burn capabilities [10].
These features of hydrogen have been recognized to have important implications in improving both the emission
characteristics and thermal efficiencies of NG engines. In addition, hydrogen is characterized as a clean and renewable
energy carrier. Unlike hydrocarbon fuels, the only harmful product of hydrogen combustion is NOx. Although not
present in large quantities in nature, hydrogen can be produced by a variety of methods including fossil or renewable
fuel reforming, water electrolysis and etc. [11]. The application of neat hydrogen in engines has recently received some
interests [10, 12], however, its widespread use is still likely to be hindered by many practical difficulties which include
but are not limited to large-scale hydrogen production, storage, fueling infrastructures as well as engine abnormal
combustion. At the current stage, being an additive to hydrocarbon fuels seems to be a more reasonable approach to
promote the application of hydrogen as transport fuels. Further catalyzed by the facts that hydrogen can be produced
on-board by NG reforming [13], engines running on the mixtures of hydrogen and NG have attracted great attentions
from its birth to the present.
Early efforts of testing hydrogen-NG mixtures as an engine fuel can be dated back to early 1980s. Nagalingam et al.
[14] conducted engine experiments on a research engine fueled by various HCNG blends (hydrogen content changes
between 0%, 20%, 50% and 100%) and studied the engine performance and emission characteristics. The results showed
that the engine performance fueled with HCNG blends was between the neat NG and neat hydrogen fuel cases,
depending on the hydrogen fraction. Since then, a large number of studies were conducted for the investigation of HCNG
engine performances under different hydrogen fractions and engine operating parameters, as reviewed by Sierens et al.
[15] for studies before 2000. In addition, various testing campaigns for application of HCNG in light-duty vehicles were
also carried out [16]. The general results of these studies confirmed the benefits of hydrogen addition in reducing the
hazardous emissions of NG engines and improving their thermal efficiencies. We also learned that the effects of
7
hydrogen addition is not universal but largely depends on hydrogen fraction and engine operating conditions [17].
As will be elaborated in subsequent sections, although HCNG fuel can bring many benefits, simply blending
hydrogen with natural gas does not guarantee improvements in engine performances. The hydrogen fractions need to be
optimized and corresponding adjustments are required for various other engine operating and design parameters.
Hydrogen fraction has significant impacts on engine thermal efficiency and emission characteristics. In general,
increasing the level of hydrogen addition could result in an increase of in-cylinder temperature and a decrease of
combustion duration. These can contribute to a more complete combustion and therefore higher thermal efficiency,
lower UHCs and CO emissions. On the other hand, however, higher level of hydrogen could also lead to elevated NOx
emissions due to the high combustion temperature. In addition, heat loss tends to increase with the increase of hydrogen
fraction, having a negative impact on engine thermal efficiency. As a result, the balance between thermal efficiency and
pollutant emissions need to be determined. In fact, many researchers have investigated this issue and have reached
somewhat different conclusions. For instance, Moreno et al. [18] suggested the best balance could be achieved by using
the HCNG30 fuel, while HCNG20 fuel blend was recommended by Ma et al. [19], depending on the operating conditions.
When it comes to determining the optimal HCNG fuel compositions for applications in engines, it is essential to
understand the effects of hydrogen addition on the blend’s fundamental physicochemical properties. These properties,
including density, diffusivity, adiabatic flame temperature, laminar flame speed, ignition delay and etc., play important
roles in the combustion process and thus significantly affect engine power output, fuel economy as well as emission
performances. Furthermore, engine design parameters and operating conditions such as compression ratio, idling speed,
ignition timing and equivalence ratio also exert important impacts on engine performance and therefore should be
adjusted and optimized, considering their interplay with fuel properties. Many experimental and numerical studies
focusing on the influences of these parameters on HCNG engine performance have been conducted and, based on these
results, researchers try to understand the various effects of hydrogen addition and find out the optimum hydrogen fraction
and engine design/operating parameters.
8
1.2 Motivation and outline of this review
This paper aims to provide a comprehensive summary of recent research progress in both the fundamental
combustion properties of HCNG mixtures and the effects of its application on engine’s combustion and emission
performance. Our focus will be the application of HCNG in SI reciprocating engines, while compression ignition (CI)
engine [20] and Wankel rotary engine [21, 22] fueled with HCNG fuels are beyond the scope of the current work.
Indeed, there are existing reviews in the literature summarizing the application of natural gas as energy sources in
the transportation sector. For instance, Khan et al. [2] has systematically discussed the fuel properties of NG, its storage,
safety issues as well as the effects of its application on IC engine’s power, efficiency and emission performances. The
influence of fuel composition on the combustion and emission characteristics of natural gas engines were addressed by
Kakaee et al. [23]. A detailed comparative analysis between natural gas fueled spark-ignition and compressed ignition
engine performance and emissions has been presented by Korakianitis et al. [24]. These reviews were intended for NG
engines and did not cover the effects of hydrogen enrichment.
On the other hand, Tang et al. [25] presents a comprehensive review of the fundamental combustion properties of
hydrogen enriched hydrocarbons. The focus was given to the effects of hydrogen addition on flammability limits,
ignition delay and laminar flame speeds of various hydrocarbon fuels. Wang et al. [26] summarized mostly their own
research on the application of HCNG in direct injection (DI) engines and included also related optical studies on the
propagation and structures of turbulent premixed HCNG flames, with limited discussion on engine performances.
Akansu et al. [17] discussed the research progress and utilization of natural gas–hydrogen mixtures in IC engines before
2004 and the overall benefits of hydrogen enrichment on improving NG engine performance were discussed by
Nanthagopal et al. [27]. Sun et al. [28] reviewed the development of hydrogen-fueled engines in China and extended
their discussion to include hydrogen-enriched gasoline engine, hydrogen-enriched diesel engine, hydrogen-natural gas
fuel engine, and pure hydrogen-fueled engines, the latter of which is also the topic of more recent reviews [12, 29].
A closer look at the above-mentioned studies reveals that although hydrogen enriched natural gas is frequently a
9
review subject, in many cases it is not the primary topic and thus is only briefly mentioned. The current study intends
to discuss both the fundamental combustion properties of HCNG fuel and its relationship with engine combustion
behavior and performances. Special emphasis is given to the linkage between engine performance and fuel properties
such as ignition delay, laminar burning velocity, flammability limits and etc. These fuel properties could directly
determine the engine combustion process by influencing flame propagation, heat release rates and pollutant formation
inside the cylinder. The interactions between fuel properties, engine operating conditions and their effects on engine
performance are schematically shown in Fig. 2. Considering these interactions, the overall structure of this review is
organized as follows.
First, the physical and chemical properties of hydrogen and natural gas mixtures are introduced in Section 2. Special
attention was given to ignition delays and laminar burning velocities, considering their prominent impacts on the engine
combustion processes. Afterwards, the influence of hydrogen addition on NG engine power output, fuel economy and
emission performances are analyzed in detail in Section 3. Engine lean burn characteristics is given a special attention
since many benefits of HCNG engines, as compared to traditional NG engines, come from the fact that hydrogen
addition can significantly extend the lean burn capacity of NG engines. Furthermore, as engine design and development
can be greatly facilitated by numerical models nowadays, computational investigations of HCNG engines based on
different engine combustion models are presented in Section 4. Although this study focuses more on hydrogen enriched
natural gas, the effects of hydrogen addition on other practical transport fuels are also briefly discussed in Section 5, for
a more comprehensive overview of the application of hydrogen in IC engines. Lastly, concluding remarks are presented.
10
Fig. 2 Interplay among fuel property, operating conditions and engine combustion and emission performances
2. Fundamental physicochemical properties of HCNG
A thorough investigation of the fundamental physicochemical properties of the HCNG fuel is a prerequisite for the
understanding of the combustion and emission characteristics of HCNG engines. Engine overall performances are highly
dependent on the in-cylinder combustion processes such as ignition, initial flame development, subsequent flame
propagations and end gas auto-ignition. These combustion processes are in turn closely related to fundamental fuel
properties including, for example, minimum ignition energy, laminar flame speed, quenching distance, ignition delay
and etc. Therefore, comprehensive discussions of the physiochemical properties of HCNG fuel are firstly presented in
this section, with a focus on the effects of hydrogen fractions considering its significant influences on the fuel
combustion characteristics [30] and engine performances.
Table 1 summarizes several important fuel properties of methane and hydrogen, which are the major components
of HCNG. Properties of typical gasoline and diesel fuels are also shown for comparison purposes. Detailed discussions
on important individual properties are presented below.
11
Table 1 Typical physiochemical properties of relevant fuels [2, 10, 20, 26, 31]
Properties
Methane
Hydrogen
Gasoline
Diesel
Lower heating value (LHV) (MJ/kg)[26]
46.72
119.7
44.79
42.5 [20]
Volumetric LHV (MJ/m3) [26]
32.97
10.22
216.38
-
LHV of stoichiometric mixture
（MJ/m3）[26]
3.13
3.02
3.83
-
Density at NTP (kg/m3)[31]
0.67
0.08
720-775
833-881 [20]
Molar mass (kg/mol) [31]
16.04
2.02
100-105
204 [2]
Diffusion coefficient (cm2/s) [10]
0.189
0.61
-
-
Flammability limits in air
(vol%) (lower–upper) [10]
5.3-15.0
4.0-75.0
1.2-6.0
0.7-5 [20]
Laminar flame speed at NTP (m/s) [26]
0.38
2.65-3.25
0.37-0.43
-
Auto-ignition temperature in air (K) [10]
813
858
~500-750
~553 [20]
Adiabatic flame temperature in NTP air (K)
2224a
2379a
2470 [26]
2327 [2]
Minimum ignition energy (mJ) [26]
0.28
0.02
0.25
-
Quenching distance in NTP air (mm) [26]
2.03
0.64
2.0
-
*NTP means normal temperature (298.15K) and pressure (1 atm)
a denotes the authors’ calculations at NTP conditions with Chemkin software and GRI mech 3.0.
2.1 Lower heating value
Lower heating value (LHV) represent quantitatively the combustion heat released from a given amount of fuel
when complete combustion is achieved. As can be seen in Table 1, although methane and hydrogen have relatively
higher LHVs per unit mass than gasoline or diesel, their LHVs per unit volume are much lower. The volumetric energy
density of the stoichiometric fuel-air mixtures is also considerably lower for hydrogen and methane. This means that
the released combustion heat and work output will be reduced for S.I. engine fueled with HCNG mixtures. In addition,
due to their gaseous nature, hydrogen and methane also have lower stoichiometric mixture densities and therefore
occupy more space at the engine intake, which would lead to less intake of fresh air and thus lower engine volumetric
efficiency. The combined effects of low mixture density and reduced volumetric efficiency result in less work output
for engines fueled with HCNG or NG as compared to their gasoline counterparts. Relevant studies have demonstrated
that power output of a NG-fueled engine is typically 10-15% below that of a gasoline engine [24]. This drawback,
however, can be partially compensated by employing higher compression ratio (CR) and turbocharging technology. In
addition, direction injection (DI) strategy [32] could also be used to maintain high power output of gaseous fuel engines.
12
Fig. 3 Volumetric LHV of HCNG-air mixture at various hydrogen fractions and excess air ratios. Reproduced from Ma et
al. [33].
A comparison between methane and hydrogen shows that although the volumetric LHV of hydrogen is only around
1/3 of that of methane, hydrogen requires 50% less of air, on a per volume basis, for complete combustion. As such,
even the addition of hydrogen in methane will certainly reduce the LHV of the fuel per unit volume, its effect on the
volumetric heating values of the fuel/air mixture may exhibit more interesting trends. This is evident in Fig. 3 where the
volumetric LHVs of fuel-air mixtures are plotted as a function of hydrogen fraction (XH2) for various mixture excess air
ratios (λ). It is noticed that for stoichiometric mixture (λ = 1), the mixture LHV decreases as XH2 increases. However,
for relatively fuel-lean mixtures where 1.4 < λ < 1.6, the mixture LHV becomes insensitive to XH2. As λ further increases,
the trend get reversed in that the mixture LHV starts to increase as more hydrogen is added for λ > 1.6. These trends are
relevant and thus are important to note, as we will see in later sections, many benefits of HCNG engines come from
their operations with ultra-lean fuel/air mixtures, as opposed to the stoichiometric mixture strategy for conventional
gasoline engines. Under those ultra-lean operation, the power output of NG engines can actually be boosted by hydrogen
enrichment.
2.2 Adiabatic flame temperature and auto-ignition temperature
Adiabatic flame temperature (Tad) of the fuel/air mixture is an important parameter and has strong influences on
13
the chemical reaction rates inside the cylinder [34]. High combustion temperature is beneficial for a more complete
combustion and hence could reduce HC and CO emissions. However, NOx emission would increase with elevated
combustion temperature, mainly through the thermal NOx formation route. Theoretical calculation indicated that the Tad
of CH4 / H2 / air mixture flame tend to increase with increasing XH2 [35-37], although the increment is slight. For
stoichiometric fuel-air mixture, as reported by Boushaki et al. [35], Tad increases only 10 K with 20% hydrogen addition,
while this value was 33 K with 50% hydrogen addition.
Fig. 4 Variations of adiabatic flame temperature in air as a function of excess air ratio and hydrogen fraction. Calculated
with Chemkin software and GRI mech3.0 at temperature of 298 K and pressure of 1 atm.
Data shown in Fig. 4 confirms the increasing trend of adiabatic flame temperature with the increase of hydrogen
fractions and the trend is also seen to be similar at all excess air ratios (λ). Furthermore, the effects of λ on adiabatic
flame temperature are more pronounced than those of hydrogen ratios. Increased λ results in a notable decrease of
combustion temperature due primarily to the reduced fuel content and the dilution effects of the air [34]. Although
hydrogen addition can increase peak flame temperature of CH4 / H2 / air mixtures [38, 39] from a theoretical point of
view, in real engine applications, however, its effects of on maximum in-cylinder gas temperature (Tmax) may depend on
other factors such as charge efficiency, combustion phasing, heat loss and etc. As a result, the trend may change
depending on engine operating conditions. For example, Wang et al. [32] studied the in-cylinder combustion behavior
14
of a single cylinder DI SI engine at different load conditions. The results showed that Tmax increased with increasing
hydrogen fraction and the rate of increase become larger when XH2 exceeds 20% for all tested load conditions. The
authors attributed this result partially to the improved combustion speed after hydrogen addition. However, in another
study conducted at 70% wide open throttle (WOT) condition and an engine speed of 1200 rpm for a DI S.I. engine,
Zheng et al. [40] reported that when different HCNG blends with XH2 of 0-40% was injected at constant injection
pressure and duration, Tmax increased when XH2 was less than 10%, but when XH2 exceeds 10%, Tmax decreased with
increasing hydrogen fraction since the reduction of lower volumetric heating value of HCNG mixture become important.
Thus, when it comes to the effects of hydrogen addition on practical engine in-cylinder gas temperature, case by case
analysis considering all the relevant factors becomes necessary.
Auto-ignition temperature is a relevant parameter for the auto-ignition process of the end-gas in SI engines.
Hydrogen has the highest auto-ignition temperature among the four types of fuels listed in Table 1. This indicated that
the ignition of hydrogen-air mixture is relatively harder than the other three fuels, which is beneficial for the avoidance
of engine knock.
2.3 Flammability limits
Flammability limits refer to the flammable range of the fuel-air mixture in terms of molar fuel ratio. Both the lower
and upper limits can be defined. Due to various thermal and chemical reasons, fuel/air mixtures would not be able to
ignite and propagate in a self-sustained way when fuel/air ratio is higher than upper flammability limit (UFL) or lower
than lower flammability limit (LFL). This critical fuel property provide important information for explosion hazards
and the design of lean-combustion devices [41].
15
Fig. 5 Flammability limits of methane-hydrogen mixtures determined by various experimental methods. Data obtained from
Miao et al. [42]
Hydrogen has a much wider flammability range as compared to methane. Table 1 gives typical values of UFL and
LFL of hydrogen, methane, gasoline and diesel at atmospheric pressure and normal temperature condition. In fact,
accurate determination of UFL and LFL is somewhat an experimental challenge and the values of the flammability limits
may differ if different experimental methods are used [43]. Miao et al. [42] systematically reviewed the UFL and LFL
of CH4/H2/air mixtures obtained with various experimental approaches, which, as Fig. 5 shows, did have notable
influences on the absolute values of flammability limits. In addition, noting that natural gas has many other components
besides methane, each of which may have different flammability limits, Miao et al. continued to compare the UFL and
LFL values between CH4/H2 mixture and HCNG mixture. Their results indicated that the CH4/H2 mixture flammability
data is good enough to represent those of HCNG [42].
The initial pressure and temperature of the reactants also have significant effects on the flammability limits [44,
45]. Schoor et al. [44] measured the UFL of CH4/H2 mixture under different combinations of initial temperatures and
pressures in a constant volume combustion vessel. A 5% pressure rise was chosen as the criterion for flammability and
the results showed that the UFL increases linearly with both initial temperature and pressure.
Compared to UFL, LFL is more pertinent when it comes to HCNG mixtures. This is because in real applications,
HCNG is typically burned under fuel lean conditions. Hydrogen addition can extend the lean burn limit of NG engine,
16
which is believed to be a remarkable advantage of HCNG considering the possibility of improving engine thermal
efficiency and reducing pollutant emissions through lean combustion. LFL of the fuel is one of the most important
factors that determine the lean burn limit of HCNG engines. Shoshin et al. [41] experimentally measured the LFL of
CH4/H2 mixture with a XH2 range between 0 and 40%. The LFL was determined by visual observation of flame in tubes
of various diameters. The experimental results indicated that at a given tube diameter, the LFL in terms of equivalence
ratio was extended after hydrogen addition, which can be partially attributed to the preferential diffusion effects of
hydrogen [41].
Fig. 6 Effects of hydrogen fraction on lean counter-flow premixed CH4/H2/air flames. Adapted from Guo et al. [46].
Guo et al. [46] investigated the effects of hydrogen addition on the extinction strain rates of CH4/air mixtures in
counterflow premixed flames and the results are shown in Fig. 6. As can be seen, at a fixed equivalence ratio, hydrogen
addition can noticeably increase the extinction strain rate of the fuel/air mixture (upper part of the C curve) and the
equivalence ratio decreases with the increase of XH2 for a fixed extinction strain rate. The author attributed these findings
to the higher flame speed and lower LFL of hydrogen. Additionally, preferential diffusion of hydrogen results in a higher
combustion intensity and thus enhanced combustion stability for lean fuel mixture. As a result, flammability limit can
be extended to a lower equivalence ratio [46].
17
2.4 Minimum ignition energy
In spark ignition engines, the ignition of the fuel/air mixtures is initiated by an electric spark. The minimum energy
deposited by the spark into a combustible gas mixture for its successful ignition is defined as the minimum ignition
energy (MIE). The study of MIE is critical for our understanding of the flame kernel development, which had a major
effect on the following flame propagation process in spark ignition engines [47]. The knowledge of initial flame kernel
development is the prerequisite for designing engines with ultra-lean combustion.
Hydrogen has a much smaller MIE value (~0.02 mJ) as compared to methane (~0.28 mJ). Therefore, the MIE of
HCNG mixtures is expected to decrease with the increase of hydrogen fraction. This was experimentally confirmed by
Ma et al. [48], who measured the MIE for CH4 / H2 / air mixtures in a 20L constant volume combustion vessel at a
equivalence ratio of 1.5. A rapid reduction of MIE was observed as hydrogen fraction increased: MIE decreased from
118 mJ for CH4-air to 0.12mJ for H2-air. Ma et al. further reported that the MIE of CH4 / H2 / air mixtures exhibited an
approximately linear relation with hydrogen fraction, as demonstrated in Fig. 7. As a result, the MIE of the fuel blends
can be estimated from the following equation：
(1)
where and are the MIE and the mole fraction of component i, respectively. This linear relationship has
further implications in that it may be reasonable to assume that the controlling factors of MIE of HCNG mixture are
similar with those of neat H2 or CH4. This could be useful in quantifying the influence of various factors (i.e. electrode
gap distance, temperature and pressure conditions and etc.) on the MIE of HCNG mixture, even without direct
experimental data. Related parametric investigations on the MIE of H2/air and CH4/air are abundant in the literature [49-
51].
18
Fig. 7 MIE versus hydrogen content at equivalence ratio of 1.5. Measured at initial temperature of 293K and pressure of 1
atm in a 20L closed combustion vessel. Adapted from Ma et al. [48]
As an example, Han et al. [49] investigated the effects of equivalence ratio and electrode gap distance on the MIE of
CH4-air mixture. The results showed that the MIE reached a minimum when the equivalence ratio was 0.9, as evident
in Fig. 8. In theory, fuel mixture should be most easily ignited at equivalence ratio of 1 since the energy release rate is
expected to be the highest at the stoichiometric condition. However, because of the diffusion effect, the minimum values
of the MIE appeared on the lean side for the CH4-air mixture [49]. Regarding the electrode gap effects, there exists an
optimum gap distance where the required energy for ignition is minimized. As Cui et al. [50] explained, if the electrode
gap is too small, the heat loss to metal electrodes could easily exceed the heat production through chemical reaction
such that the reaction zone is quenched. On the other hand, if the electrode gap is too large, it may become increasingly
difficult to electrically break through the gaseous gap between the two electrodes, leading also to deteriorated ignition
quality.
19
Fig. 8 The effect of equivalence ratio on MIE. Measured at electrode gap distance L=1mm, and electrode radius R=12.7 mm.
Adapted from Han et al. [49]
The effect of spark duration on MIE was also investigated for CH4-air mixture [49] and H2-air mixture [51]. It was
observed that an optimum spark duration time exists for minimizing the MIE of CH4-air mixtures, as Han et al. [49]
explained that for a short spark duration, the temperature gradient and the vortex gas motion around the flame kernel
would adversely affect the formation of flame. While if the spark duration is too long, the energy dissipation through
thermal conduction would increase, also contributing negatively to ignition. The situation is somewhat different for H2-
air mixtures. Ono et al. [51] observed no change of MIE for when spark duration changes within 5 ns to 1ms in their
experiments.
Cui et al. [50] measured the MIE of CH4/air mixtures at different temperatures and pressures. The results indicate that
both pressure and temperature have a strong effect on the MIE. Further regression analysis shows that the MIE has a
linear relationship with 1/P2 and is inversely proportional to temperature. These results provide useful information for
the design and optimization of the HCNG engine ignition system.
2.5 Quenching distance
During the flame propagation in SI engines, when the flame front encounters relatively cold surfaces such as the
20
piston head or the cylinder wall, the flame may be quenched due to heat loss to the surfaces [52]. Such flame quenching
phenomena near the cylinder wall often result in incomplete combustion in the quenched zone, which become one of
the most important sources of unburned HC emissions in SI engines [53]. Since quenching distance determines the
thickness of the quenching layer [54] and hence has a significant effect on unburned HC emissions, it become a necessity
to understand the effects of hydrogen addition on the quenching distance of HCNG mixture.
Fig. 9 Experimental data of quenching distances of CH4 / H2 / air mixtures as a function of hydrogen fraction and excess air
ratio, measured at atmospheric pressure and a temperature of 15℃. Data available in Fukuda et al. [55].
Fukuda et al. [55] studied experimentally the influence of hydrogen fraction on the quenching distance of CH4/H2
mixture and the results are shown in Fig. 9. In their study, quenching distance is defined as the minimum distance
between two parallel plates below which flame propagation is no longer observable. As can be seen in Fig. 9, for a fixed
λ condition, the quenching distance decreased monotonously with the increase of hydrogen fraction (XH2). This trend is
as expected considering neat hydrogen has a much lower quenching distance (0.64 mm) as compare to that of methane
(2.03 mm), as shown in Table 1. Thermodynamically, the quenching distance represents the critical condition when the
heat loss to the cold wall is just strong enough such that the net release of heat energy from combustion cannot sustain
the chemical reactions in the flame. Hydrogen addition could enhance the combustion reaction and heat release rate,
and as a result, smaller flame quenching distance (i.e. larger heat loss) could be tolerated. It is also noticed that, for a
21
given XH2, flames with stoichiometric fuel/air mixture have the minimal quenching distances. This can be explained by
the fact that combustion reactions are most exothermic at stoichiometric conditions. Fukuda et al. further demonstrated
that the addition of small amount of hydrogen could lead to a disproportionate decrease of the quenching distance for
fuel-lean mixtures. As can be seen from Fig. 9, under the conditions of λ =1.4 and 1.6, there exists a rapid decrease as
XH2 is increased from 0 to 20%. This result showed that the effects of hydrogen addition on reducing flame quenching
are most prominent under fuel-lean conditions, suggesting the its potential contribution to improve engines’ lean burn
capabilities.
2.6 Ignition delay
Ignition delay represents the chemical reactivity of the fuel-oxidizer mixture and is generally used to quantify the
characteristic time of the auto-ignition process at a given temperature and pressure. The phenomenon of auto-ignition is
relevant to SI engine knocking, as it is now generally accepted that engine knock is initiated by the auto-ignition of
unburned end gas where the local temperature and pressure are raised by the compression of the burned gas and the heat
transfer from the flame front, to a point high enough to get ignited without any external energy assistance (e.g. electric
spark). The occurrence of engine knock may cause power loss, deteriorated emission performance and in some cases
even serious mechanical damages. It is therefore critically important for us to investigate the controlling factors for the
appearance of knock under different engine operating conditions [56, 57].
A knowledge of the auto-ignition characteristics of H2 / CH4 blends lays the foundation for our understanding of the
knock behavior of HCNG engines. A literature survey reveals the ignition delays of CH4-air or H2-air mixture have been
measured and studied over a wide range of temperature and pressures conditions. However, related studies are relatively
limited for the ignition delay of CH4 and H2 blends. We list notable measurements conducted in shock tubes and rapid
compression machine (RCM) in Table 2.
22
Table 2 Measurements of ignition delays for CH4/H2 mixture at different conditions
Source
Fuel Mixture
Experimental
conditions
Measurement device
Donohoe et al. [58]
XH2:30-80 %;
Φ: 0.3-1
T: 850-1800 K;
P: 1-30 atm
RCM and shock tube
Zhang et al. [59]
XH2: 0-100 % ;
Φ: 0.5
T: 1000-2000 K;
P: 0.5-2 MPa
Shock tube
Zhang et al. [60]
XH2: 0-100 %;
Φ: 0.5-2
T: 900-1750 K;
P: 1.8 MPa
Shock tube
Gersen et al. [61]
XH2: 0-100 %;
Φ: 0.5-1
T: 950-1060 K;
P: 15-70 bar
RCM
Chaumeix et al. [62]
XH2:20-100 %;
Φ: 0.4-1
T: 1250-2000 K;
P: 0.15-1.6 MPa
Shock tube
Gersen et al.[63]
XH2: 30 %;
Φ: 0.5-1
T: 900-1100 K;
P:20-80 bar
RCM
Huang et al.[64]
XH2: 15-35 %;
Φ: 1
T:1000-1300 K;
P:16-40 atm
Shock tube
The fact that hydrogen addition could promote the ignition chemistry of NG has been widely demonstrated [65-67].
For example, Herzler et al.[67] showed that the ignition process was continuously accelerated with the increase of H2
fraction for hydrogen-NG mixture over a temperature range of 900 K-1800 K and pressures of 1, 4 and 16 bar in shock
tubes. Ignition delays of CH4/H2 mixtures were also seen to decrease with increasing XH2 [59, 62], as shown in Fig. 10.
The addition of hydrogen enhanced the reaction rates of O + H2 H + OH and H + O2 OH +O, both of which are
chain-branching reactions and thus can lead to the production of large amounts of active radicals such as H, O and OH.
This increased radical pools in turn contribute to the reduction of ignition delays. Besides the chemical effects, the
preferential mass diffusion effects of hydrogen were also highlighted by Fotache et al. [68] and Dai et al. [69] for the
enhanced ignition of CH4/H2 blends in non-premixed counterflow ignition condition.
23
Fig. 10 Measured ignition delays for CH4/H2 blends under varying hydrogen fraction condition at equivalence ratio of 0.5
and pressure of 10 bar in shock tube. Data obtained from Zhang et al. [59]
The extent to which hydrogen addition can reduce the ignition delay exhibit non-linear relationships with hydrogen
fractions [59-61, 68, 70]. For instance, by measuring the ignition delays of 3.5% CH4 / 7% O2 mixture diluted in argon
in a shock tube, Lifshitz et al. [70] studied the effects of small amount of hydrogen addition (0.073% and 0.52%) on the
mixture ignition delays and found that, although hydrogen addition did shorten the ignition delays of methane, the
reduction was rather small and even less than the propane addition case. The authors concluded that adding small
amounts hydrogen had no chemical effects on methane ignition chemistry, while the observed reduction of ignition
delay time was caused purely by the added heat release of hydrogen. In another study, Gersen et al. [61] investigated
the ignition delay times of stoichiometric CH4/H2 mixtures in RCM at temperature and pressure ranges of 950-1060 K
and 15-70 bar, respectively. The results showed a slight reduction of ignition delays when XH2 ≤ 20%, while a significant
decrease in ignition delay was observed when XH2 exceeds 50%. Zhang et al. [59] systematically investigated the auto-
ignition characteristics of CH4/H2 mixtures with a wide range of hydrogen fractions at various temperature and pressure
conditions. Three different regimes were proposed, depending on the hydrogen fractions: (1) methane dominating
ignition chemistry when XH2 ≤ 40%, (2) combined chemistry of methane and hydrogen dominating ignition when XH2 =
60% and (3) hydrogen dominating ignition chemistry when XH2 ≥ 80%.
24
In addition, various investigations have shown that factors such as temperature, pressure and equivalence ratios also
affect the quantitative extent of hydrogen’s ignition-promoting effects [59, 60]. For instance, it was shown that the
presence of hydrogen was more efficient in reducing ignition delay times of CH4 / H2 mixtures at high temperature
conditions, and this is believed to be caused by the different fuel ignition chemistries dominating at high and middle-
low temperature range [61, 64]. At high temperature conditions, OH + H2 O + H2O and H + O2 O + OH are the
major reactions controlling ignition, and their rates can be enhanced by hydrogen addition. As a result, the high
temperature ignition delay is noticeably decreased with the increase of hydrogen contents. At low temperature, however,
the slower reaction of CH3O2 + H2 CH3O2H + H become important for the ignition. The effect of hydrogen addition
on such reaction and thus its influence on ignition delay is considerably weaker.
Fig. 11 Measured ignition delays times of CH4/H2 blends at different pressure and hydrogen fraction condition in RCM.
Compression temperature T=995±4K. Data obtained from Gersen et al. [61]
Huang et al. [64] and Gersen et al. [61] reported that the ignition-promoting effects of hydrogen are more pronounced
at lower pressure conditions. This can be seen in Fig. 11, where the ignition delay for various CH4/H2 mixtures at 995
K are plotted as a function of pressure. The reduction of ignition delays by hydrogen addition is noticeably smaller at
high pressure condition than that at lower pressures, especially when H2 fraction is high. This is explainable considering
the fact that ignition chemistry at the tested temperature is mainly dictated by the competition between the recombination
25
reaction H + O2 (+M) HO2 (+M) (ignition inhibiting) and the chain branching reaction H + O2 OH + O (ignition
promoting). Kinetic analysis indicated that at high pressure regime, the three-body recombination reaction H + O2 (+M)
HO2 (+M) become important. The increased consumption of the active radicals of H by this reaction leads to the
weaker effect of hydrogen addition on the decrease of ignition delay at high pressure.
It is also noticeable in Fig. 11 that the ignition delays at around 995 K decreased with the increase of pressure for all
tested hydrogen-methane blends [61], including pure hydrogen-air mixture. However, the experimental results of Zhang
et al. [59] suggested that the effect of pressure on ignition delays is actually more complicated than it may seem in Fig.
11, as the effects may vary as hydrogen fractions and temperature condition changes. When XH2 ≤40 % (methane-
dominating chemistry regime), the ignition delays tend to decrease with increasing pressure and this regime exhibit
typical dependence of ignition chemistry on pressure for hydrocarbons fuels such as neat methane [71]. On the other
hand, Zhang et al found that the ignition delays of pure hydrogen could even increase with the increase of pressure at
the temperature range of 1093-1170 K. In fact, the ignition delays of hydrogen at the pressure of 2.0 MPa is almost ten
times longer than that at the pressure of 0.5 MPa for T = 1093 K [59].
Equivalence ratio also plays a role in further complicating the effects of hydrogen addition on ignition delay of CH4/air
mixtures. Zhang et al. [60] reported that when XH2 ≤40 % (methane-dominating chemistry), ignition delays increased
with the increase of equivalence ratio for both high temperature and middle-low temperature conditions. When XH2=
60% (combined chemistry of hydrogen and methane), the ignition delays continue to increase with increasing
equivalence ratio at high temperature conditions, but it becomes rather insensitive to the change of equivalence ratios at
middle-low temperature conditions. For the cases where XH2 ≥80% (hydrogen-dominating chemistry), ignition delays
exhibit more complex dependence on equivalence ratio that the variations of ignition delays had no consistent behavior
and changed at different temperature range.
The above research results suggest a complex variation of ignition delay of CH4 / H2 / air mixtures with pressure,
temperature and hydrogen fractions. Nevertheless, empirical calculation formulas for ignition delay of the mixtures have
26
been proposed, based on statistics analyses of related experimental results. Many such equations correlate the ignition
delays of fuel blends with the those of individual fuel and the molar fractions of the each fuel component [72]. For
example, Cheng et al. [73] correlated the ignition delays of the CH4/H2/O2 mixtures through the following formula:
(2)
whereandare the ignition delay times of neat hydrogen and methane, respectively; is the mole fraction of
hydrogen in the fuel blends. Based on Eq. (2), Gersen et al. [61] linked the ignition delay time in a rapid compression
machine with the maximum pressure and the compression temperature, using an Arrhenius-like empirical relation:
(3)
where are fitting coefficients. Then the ignition delays of methane-hydrogen mixture can be derived from
neat hydrogen and methane fuels by introducing the weighting factor of hydrogen mole fraction, as follows:
(4)
Eq. (4) performs well in the prediction of the ignition delays of H2 / CH4 mixtures pressures of 15-70 bar, and
temperatures range of 950-1060 K and stoichiometric conditions measured in rapid compression machines [61]. These
correlations provide engineering tools for engine combustion modeling and thus are beneficial for HCNG engine design
and control optimizations.
2.7 Laminar flame speed
Laminar flame speed (SL) is a particularly important fuel property and has profound implications in the flame
propagation processes in S.I. engines. For S.I. engines, the rate of heat release, which influences the overall engine
performances, are dictated by the in-cylinder flame propagation speed. Although this flame propagation speed can be
significantly affected by the in-cylinder turbulence level, it still has close relationship with the fundamental laminar
flame speed.
Laminar flame speed (SL) is a quantitative characterization of the fuel-air mixture’s chemical reactivity, combustion
enthalpy and physical diffusivity [74]. In this section, the SL of HCNG mixtures and its relationship with various
27
variables, i.e. fuel composition, temperature and pressure, will be reviewed. It is worthwhile to mention that accurate
experimental determinations of SL are frequently counteracted by the presence of flame stretch [75] (e.g. in the constant
volume combustion bomb method and the stagnation flame method), flame instability [76] and/or non-adiabatic effects
(e.g. in the heat flux method). In this study we are mainly focusing on the influencing factors of SL and thus only derived
un-stretched, adiabatic flame speeds are considered without discussing the details of the measurement techniques.
Both numerical and experimental studies have showed that the SL of CH4/H2 mixture tend to increases with the
increase of hydrogen fractions [35]. Figure 12 presented the experimental SL for CH4/H2/air mixtures with a wide range
of hydrogen fractions and equivalence ratios [77]. As can been seen, SL monotonically increases with the increase of
hydrogen fraction, all the way to the pure hydrogen case. The trends are similar for all equivalence ratio conditions. Hu
et al. [77] pointed out that there exists strong correlation between SL and maximum radical concentrations of H and OH
radicals. Hydrogen addition contributes to the increased SL since it can increase the concentrations of H, O and OH
radicals and thus enhance the mixtures’ chemical reactivity [77-79]. In addition, the increased adiabatic flame
temperature brought by hydrogen enrichment also contributes to the increase in SL [78].
It can be further noticed from Fig. 12 that the rate of increase of SL is rather nonlinear with hydrogen fractions (XH2).
For example, at an equivalence ratio of 1.1, increasing XH2 from 0 to 20% results in an increase of SL from 0.37 m/s to
0.44 m/s while increasing XH2 from 80% to 100% lead a jump of nearly 1.0 m/s for SL. In general, two distinct regime
can be identified depending on XH2. In the case of XH2 ≤ 60%, a linear and relatively small increase in laminar burning
velocity can be seen with the increase of hydrogen fraction. In the case when 60% < XH2 ≤ 80%, SL increase exponentially
with XH2.
28
Fig. 12 Measured SL (combustion vessel method, initial temperature: 303 K, pressure: 0.1MPa) as a function of equivalence
ratio and hydrogen fractions. Adapted from Hu et al. [77]
Fig. 13 Measured SL (combustion vessel method, initial temperature: 353 K, pressure: 0.1MPa, stoichiometric mixture) as a
function hydrogen fractions. Adapted from Okafor et al. [80].
As a matter of fact, such non-linearity increase of SL with hydrogen fraction has been confirmed by many researchers
[58, 78, 80, 81]. For instance, as clearly depicted in Fig. 13, Okafor et al. [80] found that when XH2 ≤ 50%, the increase
of SL with XH2 was gradual, but when XH2 > 50%, SL increased rather rapidly with XH2 increase. In a numerical study,
Sarli et al. [81] divided the 0~100% range of XH2 into three regimes based on the extent of hydrogen addition effects on
SL: (I) methane-dominated combustion (0 < XH2 < 0.5) in which only a slight increase of SL with XH2 is observed, II)
29
transition regime (0.5< XH2 <0.9) and (III) methane-inhibited hydrogen combustion (0.9 < XH2 < 1), characterized by the
rapid increase of SL with increasing XH2. These results suggest that SL of binary CH4/H2 fuel blends cannot be obtained
simply from the linear combination of the SL of each individual fuel. This fact should be taken special care of if
phenomenological combustion models are to be applied for HCNG engine modelling.
Besides hydrogen fraction, initial temperature and pressure conditions also exert large influences on SL of CH4/H2
mixture. Hu et al. [82] measured the SL of CH4/H2 mixtures in a combustion vessel at various pressures and temperatures
and the results are shown in Fig. 14. It can be seen that regardless of the hydrogen fraction, SL of the H2/CH4 mixtures
always decreases with initial pressure (a) and increases with initial temperature (b).
Fig. 14 SL at different initial pressures and initial temperatures for CH4/H2 mixtures (Φ=0.8), measured at (a) initial
temperature of 373 K under different pressure condition. (b) initial pressure of 0.5MPa under different temperature conditions.
Adapted from Hu et al.[82].
Hu et al. [83] revealed that the effect of pressure on SL is through its influences on the following reactions: H + O2
O + OH being the main chain branching reactions and H + O2 + H2O HO2 + H2O being the main recombination
reactions. With the increase of pressure, the reaction rate of the pressure-insensitive reaction H + O2 O + OH is
almost unaffected, while the three-body recombination reaction H + O2 + H2O HO2 + H2O, which can effectively
30
remove the active radicals, gets progressively enhanced. As a result, the overall chemical reactivity and thus SL of the
fuel mixtures rate decreased with the increase of pressure. Elevating the initial temperature, however, will enhance the
temperature-sensitive chain branching reaction more significantly than the recombination reaction and thus help to
increase the concentration of OH, O and H radicals. Therefore, laminar flame speed tend to increase as initial
temperature increases. It is noticeable that the increment of SL becomes more pronounced at larger XH2 at any given
pressure condition, which is consistent with the results obtained at atmospheric pressure.
The effect of N2 dilution [84-86] and CO2 dilution [87] on SL of CH4/H2 mixture have also been investigated. The
results generally showed that SL are reduced with the increase of dilution ratio, as expected. This may be explained that
increased dilution of N2 or CO2 reduce the adiabatic flame temperature of mixture [86] and thus lower the chemical
reactivity and laminar flame speed.
Laminar flame speedis such an important parameter in developing SI engine combustion models that its accurate
prediction becomes a prerequisite in many model-based design projects. Although combustion reaction kinetics have
been recently advanced to a point where SL, especially for simple fuels such as CH4 and H2, can be reasonably accurately
computed using detailed chemical kinetic calculation, empirical equations for SL can still be useful in applications where
computational power are limited. As can be inferred from the above discussions, laminar flame speed can be generally
described empirically as a function of the equivalence ratio, initial mixture temperature and pressure. As for the
calculation of binary fuel mixtures of CH4 and H2, XH2 should also be paid special attention to considering the strong
nonlinear relationship of SL with XH2 as presented before.
Huang et al. [88] proposed a formula for calculating SL of hydrogen-natural gas-air mixture at normal temperature
and pressure conditions. The authors suggested that SL can be obtained for any given hydrogen fraction based on the SL
of natural gas-air mixture and hydrogen-air mixture through Eq. (5)-Eq. (7):
(5)
(6)
31
(7)
where SL,x denotes the laminar flame speed of HCNG fuel with a hydrogen fraction of x% and Φ is the equivalence ratio.
Similarly, Chen et al [89] correlated SL of CH4/H2 mixture with that of individual component as represented below:
(8)
(9)
where ρu is the density of unburned fuel mixture, m is the laminar flame speed in terms of mass flux and Y refers to the
mass fraction. respectively. The coefficient c is a free parameter and can be determined at one selected hydrogen
blending ratio. Hermanns et al. [86] summarized the experimental laminar flame speed data of CH4/H2 available in the
literature and proposed the following empiric correlation for SL when XH2≤40%, which can be used at elevated initial
temperatures :
(10)
where is dilution level, is temperature-dependent coefficient, is a fitting coefficient related to dilution level.
is initial mixture temperature and =298 K. This formula can predict the relationship of SL with hydrogen fraction,
equivalence ratio, dilution level and temperature. However, since the experiment data were obtained at atmospheric and
lower pressures, the calculation method of SL using Eq. (10) is not suitable for high initial pressure conditions.
3. HCNG engine performances
In the preceding section, the fuel properties of the HCNG mixture, especially those relevant to its application in S.I.
engines, were discussed. As we have mentioned before, the knowledge of HCNG fuel properties is a prerequisite for a
further understanding of HCNG engine combustion characteristics and its interactions with engine performance and
emissions. Over the past decade, many experimental studies on HCNG engine performance and emissions have been
conducted under a vast range of engine operating conditions and hydrogen fractions. The results of these studies have
significantly advanced the technological development of HCNG engines, in addition to the fundamental science of
32
hydrogen enriched combustion. We have summarized these studies in Table 3. Note we only listed relevant studies
conducted after 2007 and earlier investigations can be found in a previous study [17] .
In this section, we shall cover the power output, fuel economy, emission performance and cycle-by-cycle variations
of HCNG engines, while keeping in mind that many of the features of HCNG engine, as compared to NG engines, come
directly from the effect of hydrogen addition on the fundamental combustion characteristics of NG fuel.
Table 3. Notable experimental testing of spark ignition engine (vehicle) fueled with HCNG mixture in the last ten years (not
meant to be an exhaustive list).
Year
Apparatus and experimental
conditions
Hydrogen
fraction
Main topics
Reference
2007
Single cylinder, DI S.I. engine;
Engine speed:1200 and 1800 rpm
Engine load: 0.14-0.63 MPa (BMEP)
0-37%
Effect of hydrogen fraction and engine
load on thermal efficiency and
combustion behavior;
Ref [32]
2007
3-cylinder S.I. engine;
Engine speed:1300-2200 rpm;
Full load; MBT spark timing
29%
Effect of hydrogen addition on lean
limit, HC and NOx emissions
Ref [90]
2007
6-cylinder S.I. NG engine
Engine speed:1200 rpm
Engine load: 105 kPa (MAP)
0-50%
Effect of hydrogen addition on thermal
efficiency and emissions under
different spark timing and λ condition
Ref [91]
2007
Ford 4-cylinder S.I. engine;
Engine speed:2000 rpm;
Constant engine load;
Fixed ignition timing
0-30%
Effect of hydrogen content on emission
performance
Ref [92]
2008
6-cylinder turbocharged S.I. engine;
Engine speed:800-2800 rpm
Engine load:50-130 kPa (MAP)
0-45%;
Effect of hydrogen fraction, engine load
and speed, ignition timing and λ
combustion characteristics and CCV
Ref [93]
2008
6-cylinder turbocharged S.I. engine
Engine speed:1600 rpm
Engine load: 87 kPa,120 kPa (MAP)
0-50%
Effect of hydrogen fraction and ignition
timing on combustion characteristics,
emissions
Ref [94]
2008
6-cylinder S.I. engine;
Engine speed:1600 rpm;
Engine load: 0.55MPa (MAP)
20%
Effect of ignition timing and λ on CCV
Ref [95]
2008
6-cylinder turbocharged S.I. engine
Engine speed:800 rpm; No load
0-50%
Idling emission performance and CCV;
Ref [96]
2008
6-cylinder NG S.I. engine;
Engine speed:1200-2800rpm
Engine load:50-125 kPa (MAP)
0-50%;
Effect of hydrogen fraction, engine load
engine speed and ignition timing on
lean operation limit
Ref [97]
33
2008
3-cylinder S.I. engine;
Engine speed: 2000 and 3000 rpm;
Engine load:0.16;0.32MPa (BMEP):
MBT spark timing
0-40%;
Effect of hydrogen fraction on CCV
Ref [98]
2008
Vehicle test over NEDC cycle with a
Euro III Daily 2.8 NG engine
10-15%;
Effect of lean burn and stoichiometric
strategy on emissions
Ref [99]
2009
Vehicle test over NEDC with a Fiat
Punto 1.2 Natural Power engine
12%
Comparison of emissions for HCNG
fuel and NG fuel
Ref [100]
2009
4-cylinder S.I. engine
Engine speed:2000 rpm
Engine load: 65% load and full load
0-30%
Effect of hydrogen addition on engine
performance and emissions under
varying λ condition
Ref [101]
2009
4-cylinder S.I. gasoline engine;
Engine speed:1500-3000 rpm; WOT;
0-30%
Combustion behavior and emissions.
Ref [102]
2009
HH368Q S.I. gasoline engine;
Engine speed: 2000-3000 rpm; WOT
MBT spark timing
0-40%
Effect of EGR ratio and hydrogen
fraction on CCV
Ref [103]
2009
Vehicle test with a 4-cylinder, 1.3 L
S.I. engine
18%
Comparison of idling stability and
emissions with NG engine
Ref [104]
2009
HH368Q gasoline engine
Engine speed:2000 rpm and 3000 rpm
WOT condition; MBT spark timing
0-40%
Effect of hydrogen fraction and EGR
ratio on combustion behavior and cyclic
variation
Ref [105]
2009
6-cylinder turbocharged NG engine
Engine speed: 800-2400 rpm;
Engine load: 85kPa and 100 kPa
(MAP)
20-40%
Effect of hydrogen addition and spark
timing on engine performance and
emissions at lean burn condition
Ref [106]
2010
Six cylinder S.I. engine,
Engine speed:1200 rpm, WOT;
MBT spark timing
0-55%
Effect of hydrogen fraction on
Emissions and thermal efficiency under
different λ
Ref [33]
2010
6-cylinder turbocharged NG engine
Engine speed:1200rpm
Engine load: 20-100% load condition
55%
Effect of λ, spark timing and load
condition on engine performance and
emissions.
Ref [107]
2011
Single cylinder DI S.I. engine ;
Engine speed:1200rpm;
Engine load: 70% open throttle
0-40%
Combustion behavior, Thermal
efficiency and emissions under
various ignition timing
Ref [40]
2011
On-road bus test on specified driving
cycle, with 6-cylinder turbocharged
NG engine
5-25%
Effect of hydrogen fraction on fuel
consumption, CO2 and NOx emissions
Ref [108]
2011
Three-wheeler vehicle test over
Indian driving cycles
18%
Engine Power, fuel economy and
emissions
Ref [109]
2011
Four cylinder S.I. gasoline engine
Engine speed:1000rpm
Engine load: 0.31MPa (BMEP)
0-50%
hydrogen addition on combustion
characteristics under lean burn and
stoichiometric mixture condition
Ref [110]
2011
6-cylinder, turbocharged NG engine
Engine speed: 800 rpm; No load
0-75%
Effect of hydrogen fraction on engine
idling performance and emissions
Ref [111]
34
2011
11-L heavy duty 6-cylinder S.I.
engine,
Engine speed:1260rpm; half load
0-40%
Lean combustion performance and
NOx emissions
Ref [112]
2011
HH368Q three-cylinder,
Engine speed: 3000rpm ;WOT;
λ=1.4; MBT spark timing
0-40%
Effect of hydrogen addition on CCV
Ref [113]
2011
Single-cylinder S.I. engine,
Engine speed:1400 rpm
Engine load: 6 Nm
30%
Emission characteristics under different
λ and spark timing condition
Ref [114]
2011
HH368Q three-cylinder S.I. engine,
Engine speed : 2000 rpm; WOT
20%
Effect of EGR and ignition timing on
combustion and emission behavior
Ref [115]
2012
Vehicle test at full load
Engine speed 1000-3500 rpm
0-50%
Hydrogen content on combustion,
power and fuel economy and CO2
emissions
Ref [116]
2012
Heavy-duty 6-cylinder S.I. engine,
Engine speed:1260 rpm; WOT
30%
Effect of turbocharging and CR on lean
limit, thermal efficiency and power
performance
Ref [117]
2012
2-cylinder S.I. gasoline engine
60h long duration test;
Engine speed: 3000 rpm
Engine load: 3.1bar (MAP)
18%;
Comparison of emission characteristics
for HCNG and NG
Ref [118]
2012
6-cylinder turbocharged NG engine;
Engine speed: 1200 rpm;
Engine load: 50 kPa (MAP); λ: 1.6;
20%;
Effect of CR and ignition timing on
thermal efficiency and power
performance;
Ref [119]
2012
6-cylinder NG engine;
Engine speed: 700-800 rpm; No load
55%;
Combustion, Thermal efficiency and
Emissions at idling
Ref [120]
2012
4-cylinder S.I. engine
Engine speed:1500-3000 rpm
Engine load: 10-40% open throttle
0-40%
Effect of hydrogen fraction, engine load
and speed, ignition timing on Lean
operation limit
Ref [121]
2012
2-cylinder S.I. gasoline engine
Enginespeed:2500-4500rpm;
Full load; Fixed spark timing
0-50%
Effect of hydrogen fraction on
combustion behaviors
Ref [122]
2013
11 L, six-cylinder NG S.I. engine;
Engine speed:1260-2200 rpm
Engine load:475-1160Nm
30%
Effect of ignition timing on emissions
and thermal efficiency
Ref [7]
2013
6-cylinder turbocharged NG engine;
Engine speed:1200rpm
Engine load: 50 kPa (MAP)
Fixed ignition timing
55%
Effect of CR on HCNG engine
combustion and emissions under
varying λ condition
Ref [123]
2013
Vehicle test over different driving
cycles
0-30%
Effect of hydrogen addition on fuel
consumption and CO2 emission
Ref [124]
2013
6-cylinder NG S.I. engine;
Engine speed:1260 rpm
Engine load: half load
30%
Effect of CR on thermal efficiency and
NOx emissions
Ref [125]
2013
11-L heavy-duty HCNG engine;
Engine speed:1260 rpm; WOT;
30%
Effect of CR on knock limit and
emissions
Ref [126]
35
2013
1.4 L turbocharged NG engine;
Engine speed:2000-4000 rpm
Engine load: 2bar -14bar
15-25%
Hydrogen addition on fuel economy,
emissions and lean burn limit
Ref [127]
2013
Single cylinder S.I. engine;
Engine speed:1500 rpm
Engine load:0.5 bar (MAP)
0-100%
In-cylinder pressure diagnosis and
cycle-by-cycle variations
Ref [128]
2014
1.4 L 4-cylinder S.I. engine;
Engine speed:2000-5000 rpm;
Full load
80-95%
Combustion behavior and emission
performance
Ref [34]
2014
11-L 6-cylinder S.I. engine
Engine speed:2100 rpm; WOT
30%
Effect of CR and ignition timing on
NOx emission; thermal efficiency
Ref [129]
2014
4-cylinder turbocharged NG engine;
Engine speed: 2000-4600 rpm
Engine load: 2.0-7.9 bar (BMEP)
0-25%
Effect of hydrogen addition on
combustion process, CCV and cylinder-
to-cylinder variations
Ref [130]
2014
Single cylinder optical S.I. engine
Engine speed: 2000 and 3000rpm
Engine load: partial load and full load
0-40%
Effect of hydrogen addition on methane
combustion behavior and emissions
Ref [131]
2014
11-L 6-cylinder Heavy-duty engine;
Engine speed: 600 rpm; No load
30%
Effect of ignition timing and λ on idling
performance
Ref [132]
2014
11-L 6-cylinder Euro V NG engine
Engine speed:1260rpm; WOT
MBT spark timing
30%
Effect of valve overlap duration on
performance and emissions
Ref [133]
2014
6-cylinder Euro IV heavy duty NG
engine, endurance test over ETC
18 %
Effect of hydrogen addition on engine
power, emissions and fuel economy
Ref [134]
2015
2-cylinder S.I. engine
Engine speed:1500-3100 rpm
Full load condition
0-30%
Effect of hydrogen fraction on power
output, thermal efficiency and
emissions
Ref [135]
2015
Modified Isuzu 4BD1 S.I. engine
Engine speed:1500 rpm; Full load;
0-20%
Effect of CR and hydrogen fraction on
cylinder pressure, torque output, fuel
consumption and emissions
Ref [136]
2015
Single cylinder S.I. engine
Engine speed:1500 rpm;
Engine load :2.98-6.18 bar (BMEP)
0-30%
Effect of hydrogen fraction on thermal
efficiency and emissions
Ref [137]
2015
Single cylinder S.I. engine
engine speed:3500 rpm; WOT
0-25%
Effect of fuel injection strategies on
combustion behavior and emissions
Ref [138]
2015
Single-cylinder S.I. engine
Engine speed:1000-2500 rpm
equivalence ratio:0.7
0-100%
Cyclic variations;
Ref [139]
2015
Heavy duty 6-cylinder S.I. engine
over European Transient Cycle
18%
Comparison of emissions and fuel
consumption with NG engine
Ref [140]
2015
Optical single-cylinder DI S.I. engine
Engine speed: 2000 rpm, full load
0-40%
Combustion process and emissions
Ref [141]
2016
Single-cylinder S.I. engine;
Engine speed: 1500rpm
Engine load: 2.98-6.18 bar (BMEP)
0-30%
Particulate matter emissions
Ref [142]
36
2016
Heavy duty 6-cylinder NG engine;
Engine speed:1000-2400 rpm;
Full road
18%
Unregulated Emissions
Ref [143]
3.1 Power performances
HCNG engines generally have comparable power output level with NG engines, which may be explained by the
following competing effects of hydrogen addition on engine power. Firstly, hydrogen addition could increase the rate of
heat release inside the cylinder and therefore is beneficial to the engine thermal efficiency and power output. Secondly,
hydrogen addition may lead to increased heat loss due to reduced quenching distance and high cylinder temperature,
contributing negatively to power output. Thirdly, unless at ultra-lean conditions, hydrogen substitution will lower the
heating value of the inducted fuel-air mixture due to hydrogen’s low volumetric energy density, which in turn tend to
reduce the power output. As a result, the effects of hydrogen addition on NG engine power performance are rather minor
in most cases, and may change depending on different engine operating conditions [116, 135]. For example, Flekiewicz
et al.[116] measured the wheel-power output of vehicles equipped with HCNG engines on a chassis dynamometer at
full load conditions, and the results are shown in Fig. 15. As can be seen, although the tests showed an increase in power
output at certain XH2 between 10% and 15%, for other hydrogen fraction range, the effect of hydrogen fraction on power
performance is not universal and changed under different engine speed conditions.
Fig. 15 Power measured on wheels at full load conditions, as a function of XH2. Adapted from Flekiewicz et al. [116].
37
The effects of hydrogen addition on HCNG engine power performances also depend on excess air ratios (λ). Ma et
al. [107] reported that when λ was smaller than 1.2, as can be seen in Fig. 16, hydrogen enrichment resulted in a slight
decrease of power output. This is expected to be caused by the lower volume heating value of HCNG-air mixtures as
compared to that of NG-air mixtures (Fig. 3). As λ increased gradually from 1.0 to 1.6, a progressive drop of engine
power was observed for both NG and HCNG cases, and this is due to the reduction of heating value of the fuel-air
mixture, at increasingly leaner conditions. With λ further increased beyond 1.6 and towards the lean burn limit, a sudden
loss of power can be noticed. The much-decreased flame temperature, flame propagation speed and thus deteriorated
combustion quality at ultra-lean conditions provide an explanation. However, from Fig. 16, it is noticeable that HCNG
engines fueled with higher hydrogen content had the potential to delay this drastic drop of engine power toward leaner
mixture. In addition, the power out for HCNG engines are significantly higher than that of NG engines when (λ>1.6).
The explanation is that hydrogen addition could extend the lean burn capability of the engine due to the its effects on
increasing flame speed and enhancing combustion quality (by providing more reactive H, OH radicals), while engine
fueled by NG only may suffer from incomplete combustion and large-scale flame quenching at lean fuel/air conditions.
Similar conclusions were also obtained by Bysveen et al. [90] who concluded that the effects of hydrogen on promoting
engine performance is more significant at lean limit conditions.
Fig. 16 HCNG engine’s power performance versus excess air ratio and hydrogen fractions. Engine speed =1200 rpm, MAP
=105 kPa at MBT (maximum brake torque) spark timing. Adapted from Ma et al. [107]
38
Although the power output of HCNG engine is much higher than that of NG engine at ultra-lean conditions, it is still
very low as compared to the stoichiometry condition. As can be seen in Fig. 16, for the HCNG30 case, the engine power
loss was nearly 30% when λ increased from 1.0 to 1.8. To take advantage of the lean burn capability of HCNG engine,
turbo-charging technology and higher compression ratio (CR) are frequently used to maintain a high power output even
at extremely lean conditions. The fast flame speed and the high auto-ignition temperature of hydrogen indicated HCNG
engine could be operated without the risk of engine knock at higher CR as compared to conventional NG engine. Such
fact can be utilized to also compensate the power reduction of HCNG engine brought by the lower volumetric energy
density of hydrogen. By increasing the compression ratio, faster heat release rate, higher peak in-cylinder pressure and
temperature can be obtained, all of which have potentials to contribute positively to engine thermal efficiency as well
as power output. Nevertheless, issues such as engine mechanical stress, increased heat transfer loss and higher NOx
emission that accompany the utilization of high compression ratios should also be dealt with in practical applications.
In addition, the increase of compression ratio may also encounter issues at lean burn condition. As can be seen in Fig.
17, Park et al. [117] observed a significant reduction of torque at fuel lean condition (λ＞1.6) when using a higher CR
of 11.5 as compared to a CR of 10.5. This result indicated the potential difficulty of lean burn combustion with a high
CR for improving engine power performance. This may partially be attributed to the increased cyclic variations at higher
CR, as reported by Zhao et al. [123]. To address this problem, Park et al [117] suggested the strategy of using retarded
spark ignition timing to increase engine torque although this will lead to a slight drop of thermal efficiency.
39
Fig. 17 Engine torque performance at lean burn conditions with various CR. Engine speed =1260 rpm, at full load, MBT
spark timing. Reproduced from Park et al. [117].
3.2 Fuel economy
Many studies have shown that hydrogen enrichment has the potential to improve the fuel economy of NG engines
[108, 118, 124]. For example, Genovese et al. [108] conducted an on-road fuel economy tests of a HCNG fueled transit
bus over self-defined urban driving cycle. The fuel consumption rates of the bus were recorded and converted to NG
equivalent values. The hydrogen content of the HCNG fuel was varied from 0% to 25%. The results are listed in Table
4, where noticeable reductions in fuel consumption can be observed, especially when XH2 ≥ 20%.
Table 4 Energy consumption and equivalent CH4 consumption of a HCNG transit bus. Data taken from [108]
Composition
kWh/km
g/km CH4 equiv.
CH4
4.29
309.18
0
5% H2
4.12
296.31
4.2
10% H2
3.89
279.76
9.5
15% H2
3.72
267.58
13.5
20% H2
3.72
267.73
13.4
25% H2
3.65
262.73
15.0
Mariani et al. [124] also observed a 6% lower fuel consumption (in terms of MJ/km) for a passenger car fueled with
HCNG30 over the New European Driving Cycle (NEDC). The authors explained that hydrogen addition promoted the
40
heat release rate inside the cylinder, making the actual engine cycle thermodynamically more resembling the ideal Otto
cycle, and thus increased the engine thermal efficiency. It is worthwhile to mention that reduced fuel consumption also
contribute to lower specific CO2 emissions, and the potential of HCNG fuels in reducing greenhouse gas (CO2) emissions
is considered to be one of its advantages over conventional hydrocarbon fuels [144, 145]. Mariani et al. reported that
engines fueled with HCNG30 has 15% lower CO2 emissions as compared to NG, due to both the lower carbon content
and decreased fuel consumption. Genovese et al. [108] obtained similar conclusions, which can be clearly seen in Fig.
18. The actual reduction of CO2 emissions were larger than the reduction of carbon input due to hydrogen dilution effects
(carbon substitution in Fig. 19), which demonstrated the benefits of hydrogen addition in reducing specific fuel
consumption.
Fig. 18 Comparison of CO2 emissions reduction of HCNG engine fueled with different hydrogen fractions, in relation to CO2
emissions of pure CH4 fueled engine normalized to 100. Adapted from Genovese et al. [108].
The above results showed that hydrogen addition contribute positively to the fuel economy of NG engines. However,
there are also studies that demonstrated the opposite trends, indicating the effects of hydrogen enrichment on fuel
economy may not be universal. For example, in a NEDC test for a passenger car, Unich et al. [100] reported that using
HCNG12 fuel did not show any positive effects on engine fuel economy as compared to NG fuel, if the same ignition
timing map were used for both fuels. This fact was supported by Subramanian et al. [109], who conducted a fuel
41
economy test for a three wheeler vehicle over Indian Driving Cycle. Their results showed that the fuel consumption
even increased when using HCNG18 fuel, provided no ignition timing adjustment and operating parameter optimization
were done to the original NG engine.
Michikawauchi et al. [110] further reported that, even if the spark timing was adjusted to the MBT condition, a loss
of engine thermal efficiency could still be observed after hydrogen addition. As shown in Fig. 19, at stoichiometric
condition the brake thermal efficiency decreased from 26.9% for NG-fueled engine to 25.6% for HCNG50-fueled engine.
The authors explained that hydrogen addition increased the in-cylinder temperature, shorten the flame quenching
distance and as a result increased the heat transfer loss to cylinder wall, which then led to increased cooling loss and
reduced brake thermal efficiency. However, if the engine were operated at fuel-lean instead of stoichiometric conditions,
improved thermal efficiency with hydrogen addition can be observed [110].
Fig. 19 Heat balance map at stoichiometric condition. Engine speed=2000rpm, engine load=50Nm. Adapted from
Michikawauchi et al. [110]
As can be inferred from the above results, the effects of hydrogen addition on the vehicle or engine’s fuel economies
largely depend on engine operating parameters such as ignition timing and excess air ratio. To maximize the benefits of
hydrogen addition on improving engine thermal efficiency, it is essential to systematically investigate the effects of
hydrogen addition on HCNG engine thermal efficiency under different engine working conditions.
Ma et al. [91] measured the effects of hydrogen addition on indicated thermal efficiency (ITE) of a 6-cylinder S.I.
42
HCNG engine under different excess air ratio conditions, and the results are plotted in Fig. 20. Note the spark timing
was kept the same in this set of experiments, regardless of the hydrogen fraction. It can be noticed that hydrogen addition
was not obviously beneficial to engine ITE improvement when λ1.5. In fact, 50% addition of hydrogen actually worsen
the fuel economy of the original NG engine when operated with λ1.5.
Fig. 20 Indicated thermal efficiency versusand various hydrogen fractions. Engine speed = 1200 rpm; MAP = 105 kPa;
Ignition timing = 30° BTDC. Adapted from Ma et al. [91].
This phenomenon was explained by Ma et al [91] as follows: on one hand, the increase of the fuel burning rates with
hydrogen addition reduced the engine’s combustion duration (increased the extent of isochoric heat addition) and thus
was considered to be beneficial to engine thermal efficiency [32]; on the other hand, hydrogen addition would also (1)
decrease the quenching distance (2) increase the in-cylinder combustion temperature and (3) enhance the convective
heat transfer coefficient, all of which contributed to the enhancement of the heat transfer rate from the burned gases to
the engine cooling fluid. Such increased heat loss translated directly to a reduction in thermal efficiency, as also observed
also by De Simio et al. [146]. With higher hydrogen fraction, the negative effect of heat loss became dominant, resulting
in the observed decrease in thermal efficiency. A second possible reason is the constant spark timing used in the
experiments was not optimized for HCNG fuels, which could lead to unfavorable combustion phasing for the HCNG
cases. Anyhow, Ma et al. [91] further demonstrated that even the spark timing was adjusted to MBT spark timing for
43
each case, hydrogen addition still have only very limited positive effects on ITE when λ1.5 [33, 107].
However, when the engine was operated at fuel-lean conditions (λ>1.5), the benefits of hydrogen on engine thermal
efficiency become prominent. Generally speaking, as λ increases towards the engine’s lean operation limit (LOL), the
in-cylinder flame propagation speed decreases and consequently the combustion duration increases, leading to
incomplete combustion and large cycle-by-cycle variations and therefore deteriorated thermal efficiencies. Such
behavior can be readily seen in Fig. 20, where ITE for all cases decrease as λ increases toward the LOL. However, it is
important to note that the ITE can be maintained at increasingly wider λ range for the HCNG cases with higher hydrogen
content. When λ > 1.5, the in-cylinder temperature is relatively low due to the dilution effects and hence the heat loss
becomes less significant. Therefore, whether the flame can propagate across the combustion chamber in a timely manner
becomes the dominant factor affecting engine efficiency. Under such circumstances, hydrogen’s effects on the
enhancement of flame propagation speed and the extension of flammability regime outbalance other factors and
significantly increase the ITE.
Fig. 21 Variations of thermal efficiency versus excess air ratio for different fuel/CR. Engine speed = 1260rpm, at half load
(575 Nm, half of full load) and MBT spark timing condition. Adapted from Lim et al. [125].
The strategy of using combustion chamber with high compression ratio (CR) is frequently employed in HCNG
engines considering its benefits in improving engine thermal efficiencies. Many investigations [119, 123, 125, 126, 136]
44
have concluded that the increase of CR within a certain range can indeed improve the engine’s fuel efficiencies and
power performances. Lim et al. have shown, as can be seen in Fig. 21, that with the CR being increased from 10.5 to
11.5, a remarkable increase in the ITE for both HCNG 30 and NG engines can be observed. This can be attributed to the
increased peak pressure and expansion work, which can be easily shown through a simple thermodynamic analysis, for
the higher CR case. It is also noticed that the engine fueled with HCNG30 fuel has a higher thermal efficiency than that
with NG fuel at the same CR, although the difference is rather small (less than 1% when λ < 1.5). The positive effects
of high CR on engine efficiency are not without limits and a further increase of CR may even lead to reduction of thermal
efficiency. Tangöz et al. [136] performed a test on a 3.9L diesel engine with an engine speed of 1500 rpm and full load
conditions, and the results showed that the brake specific fuel consumption (BSFC) first went down as CR increased
from 9.6 to 12.5, and then started to rise as CR further increased from 12.5 to 15. This fact indicated the necessity of
CR parameter optimization for best engine thermal efficiency. In addition, the increased knock tendency and elevated
NOx emissions should also be taken into account when optimizing CR.
Fig. 22 Variations of indicated thermal efficiency versus ignition timing and hydrogen fractions. Engine speed=1600rpm,
MAP= 125 kPa and λ=1.3. Adapted from Ma et al. [93].
At a fixed engine speed and load, the engine ITE generally increases with the increase of spark advance until the
MBT spark timing [93, 94]. Figure 22 demonstrates such trend for a boosted engine fueled by HCNG with various
45
hydrogen fractions. As expected, there always exists an optimum spark timing, independent on the hydrogen fractions,
at which engine efficiency reaches the maximum value. Either retarding or advancing the ignition timing from this
optimum point will be detrimental to engine efficiency. In addition, we can also note from Fig. 22 that although
qualitatively similar, there are still quantitative differences among cases with different HCNG fuels in terms of the
effects of spark timing on ITE: the range of spark timing within which ITE can be maintained at a relatively high value
(i.e. > 38%) become wider as more hydrogen is added, especially toward the side of retarding spark timing. This can be
attributed again to hydrogen’s fast flame propagation speed, which could counteract with the retarded spark timing and
ensures a favorable combustion phasing for high engine efficiency.
In summary, the effects of hydrogen addition on fuel economy are not universal and may vary significantly depending
on the engine design parameters, operating conditions and possibly the matching between the engine and the vehicles.
We can observe an obvious increase of thermal efficiency with hydrogen addition at ultra-lean condition, while near
stoichiometric condition, hydrogen addition has slight and sometimes even negative effects on thermal efficiency.
Besides, ignition timing should be optimized to maximize thermal efficiency since hydrogen addition changes the
original combustion phasing of the NG engine. The design parameter of CR also influence the effects of hydrogen
addition on engine thermal efficiency.
3.3 Emission performances
The most remarkable advantage of HCNG engine compared to traditional NG, gasoline or diesel engine is the
potential lower level of pollutant emissions. In general, at a given engine condition, hydrogen addition could reduce
CO2, CO and HC emissions due to decreased carbon content in the fuel mixture and the enhanced combustion process.
NOx emissions in most cases increase with the increase of hydrogen content, mainly through the enhanced thermal NO
formation route. However, these general trend being stated, the influence of hydrogen enrichment on SI engine emissions
could still vary depending on engine design and operating parameters such as compression ratio, ignition timing, excess
air ratio, engine load and etc. As a result, the engine control logics, especially those designed with emission regulations
46
as primary constraints, needs to be modified if the original engine is to be operated with HCNG fuel. To accomplish
this, it is necessary to investigate the emission characteristics of HCNG engines and its relationship with hydrogen
fraction and other related engine design / operating parameters.
3.3.1 NOx emissions
Compared to engines running with NG only, almost all available experimental results listed in Table 3 showed that
NOx emissions levels increased after hydrogen enrichment at a given operating condition. In a 60-hour engine endurance
test under constant engine speed and load conditions, Mathai et al. [118] even observed a 59.1% increase of average
NOx emissions when a NG engine is fueled with HCNG (XH2 = 18%). Such increases in NOx emissions are believed to
be caused primarily by the elevated peak combustion temperature after hydrogen addition, which favors the formation
of NOx (mainly NO) via the thermal NO formation routes (the extended Zeldovich mechanism) [147]:
O+N2NO+N (R1);
N+O2 NO+O (R2);
N+OHNO+H (R3).
R1-R3 are the major reactions responsible for the thermal NO formation. The chain propagating reaction R1 has a
rather high activation energy of 319,050 kJ/kmol which means that high temperature condition is required for R1 to be
efficient. It can also be inferred from R1 that NO formation depends on oxygen mole fractions since a high O2 level is
beneficial for providing the necessary O radicals needed for the initialization of R1. Note here that it is considerably
easier for a O2 molecule to decompose to two O atoms than a N2 molecule to two N atoms, as N2 has a stronger
intermolecular triple bond as compared to O2’s double bond.
The dependence of NOx formation on combustion temperature and oxygen availability helps to explain the
experimental trend of engine-out NOx emissions with hydrogen fraction and λ. The presence of hydrogen in the fuel
would increase both the in-cylinder flame temperature and the radical concentrations of OH, H and O, promoting the
thermal NO formation [37]. As showed in Fig. 23, NOx emissions generally increases with hydrogen fraction for a given
47
λ, although the effects are not quite obvious under either the stoichiometric or ultra fuel-lean conditions (λ > 1.8)
Fig. 23 Variations of specific NOx emissions versus excess air ratio and hydrogen fraction. Engine speed=1200 rpm,
MAP=105 kPa and ignition timing =30°CA BTDC. Adapted from Ma et al. [91]
Regarding the effects of λ on NOx emission, it can be seen that as λ increased from the stoichiometric condition, NOx
emission first increase and then decrease, with its peak occurring at slightly lean conditions, due primarily to the
competing effects of temperature and oxygen availability. At fuel rich condition (small λ), the oxygen is relatively less
abundant and cannot effectively initiate R1, leading to the reduced emission of NOx; while at very lean conditions (large
λ), the combustion temperature will drop due to the dilution effects, which also decrease the rate of NOx formation. It
is noticeable that NOx emission can be maintained at very low levels in all cases if λ exceeds a certain critical value,
and hydrogen addition tend to slightly increase this critical value. This indicated the importance of lean combustion on
reducing NOx emissions and the advantage of improved lean burn limit after hydrogen addition.
48
Fig. 24 NOx emissions versus ignition timing for HCNG30 with each CR/ λ. Engine speed=2100rpm and wide open throttle
condition. Adapted from Lim et al. [129]
Fig. 25 NOx emission versus excess air ratio for each fuel and CR at MBT spark timings. Engine speed=2100rpm and wide
open throttle condition. Adapted from Lim et al. [129]
It is worthwhile to mention that the data presented in Fig. 23 was obtained at fixed ignition timing for all fuel blends.
In fact, as showed in other studies [33, 91, 95, 107, 130], when spark timing was adjusted to MBT condition, the NOx
emissions of HCNG engine became comparable and sometimes even lower than those of the NG engines. This is due
to the fact that the MBT spark timings of the HCNG engine are more retarded and such delayed ignition timing was
beneficial to the reduction of NOx emissions [94, 95, 129, 130]. Lim et al.[129] conducted a comprehensive
49
investigation on the effects of retarding ignition on NOx level for a HCNG S.I. engine with various compression ratios
and the results are shown in Fig. 24 and 25. As can be noticed from Fig. 24, NOx emission level decreased significantly
as the ignition timing was retarded. Quantitatively, for example, the NOx emission at the ignition timing of 18°CA
BTDC was 73.5% lower than that at 28°CA BTDC for a CR of 11.5 and λ of 1.7. The primary reasons for this lower
NOx with later ignition timing can be summarized as follows: (1) late ignition timing will lead to a large portion of the
combustion heat being released in the expansion stroke, where the increase in cylinder volume will counteract with heat
release in raising the temperature and results in a reduction of maximum combustion temperature; (2) late ignition also
shortens the time available for the formation of NOx, which is between ignition and exhaust valve opening.
Furthermore, it is interesting to note from Fig. 25 that 30% of H2 addition din not lead to an obvious increase of NOx
formation for the case with CR = 11.5, as opposed to the case when CR = 10.5. This may seem confusing at first glance
since higher CR is expected to result in more NOx emission due to higher flame temperature, as obtained by Zhao et al.
[123]. However, we should keep in mind that the results in Fig. 25 were obtained at MBT timings and the MBT timings
were more retarded for the higher CR case, which helps to reduce combustion temperature and thus avoid the increase
of NOx emissions. These results suggest the effect of the retarded MBT timings on the reduction of NOx could be
greater than the effect of CR on the increases in NOx.
The above results show that, through the optimization of spark timing and enhances lean combustion performance
after hydrogen addition, NOx emission can be maintained at low levels when traditional NG engines are to be fueled by
HCNG.
3.3.2 Total HCs emissions
One major source of HC emissions in S.I. engines is the unburned fuel trapped in the combustion chamber crevices
where flame cannot propagate into. Incomplete combustion caused by flame quenching near the cold combustion
chamber wall or local fuel/air mixture inhomogeneity is another source [53]. As discussed in Section 2, hydrogen
addition in NG can effectively reduce the flame quenching distance and extend the flammability regime, both of which
50
can help reduce unburned HC emissions from the above sources. Furthermore, hydrogen addition could typically
increase the combustion temperature and contribute positively to the post-flame oxidation of the intermediate HCs, thus
leading to a more complete combustion. Finally, the replacement of hydrocarbon fuel by hydrogen reduces the carbon
content of the fuel and this will undoubtedly reduce HC emissions. Due to these facts, hydrogen addition can effectively
reduce the engines’ HC emissions, as observed by many researchers [40, 91, 92, 101, 102, 130, 131, 135, 148].
Besides hydrogen fraction, fuel-air ratios also significantly affect HC emissions. Figure 26 shows the influence of λ
on HC emissions of HCNG engine fueled with different HCNG blends. We can readily notice the general variation trend
of HC emissions with λ, as also observed by other researchers [92, 101, 102, 130], that as λ increases from the
stoichiometric conditions, the HC emissions first decrease and then increases dramatically. The first reduction of HC
emission with λ when λ<1.2 is mainly because the increased availability of oxygen in the mixture, which is beneficial
for a more complete combustion. However, when fuel mixture further leaned out ( λ > 1.2), HC emissions increased
significantly because of the deteriorated combustion instability and even misfire at ultra-lean conditions.
Regarding the effects of hydrogen, it is clear from Fig. 26 that HC emission are consistently lower for HCNG fuels
as compared to that for NG fuel, across the whole range of λ. It can also be noticed that for each fuel there exists a
critical λ beyond which the HC emission starts to increase dramatically. Beyond this critical λ, the fuel-air mixture
becomes so lean that reliable combustion cannot be sustained and large-scale quenching can frequently occur. However,
the value of this critical λ can be significantly increased through hydrogen addition. This again means that hydrogen
addition can extent the engine’s lean burn limit by improving the combustion stability at lean conditions.
51
Fig. 26 Brake specific HC emissions versus λ and hydrogen fractions. Engine speed=1200rpm, MAP=105kPa, spark timing
= 30°CA BTDC. Adapted from Ma et al. [91].
Due to its strong greenhouse gas effects, CH4 is sometimes separately regulated in emission regulations, especially
for NG engines. Unburned CH4 emissions become particularly problematic at engine idling conditions. This is because
at idling conditions, the combustion temperature is relatively low and the residual gas fraction is high (due to the
throttling effects), both of which lead to slow flame propagation and increased possibility of incomplete combustion. In
some cases, hydrogen blending in NG has been demonstrated to be an effective method to reduce CH4 emissions. As an
example, Lee et al. [132] investigated the CH4 emissions for an 11-L, heavy-duty NG engine at idling condition. The
results showed that maximum CH4 fraction accounted for approximately 96% of the total HC emissions of the NG-
fueled engine. However, if HCNG30 was used as the fuel, this value decreased to 80% for the investigated λ and spark
timing range. The authors attributed such reduction to the decreased flame quenching distance after hydrogen addition.
There are also studies in the literature that come with a contradicting conclusion. He et al. [120] showed that, even with
a high XH2 (55%), considerable CH4 emissions still exits at idling condition. The authors explained that instead of
incomplete combustion [120], the CH4 emissions are mainly caused by scavenging due to the existence of valve overlap.
Ignition retarding was shown to be an effective way for reducing engine-out CH4 emissions at idling conditions. This
is because delayed ignition helps to increase the exhaust temperature such that further oxidation of unburned fuel in the
52
exhaust manifold becomes more efficient. In addition, as explained in Refs [96, 132], delayed ignition decreases the in-
cylinder pressure and this leads to less gas being trapped in the crevice volume and thus is beneficial to decrease CH4
emissions. Taking advantage of this and after experimental confirmation with a 11-L 6-cylinder CNG engine fueled with
HCNG30 fuel at full load and constant engine speed condition, Park et al. [133] proposed the strategy of reducing HC /
CH4 emissions by decreasing the duration of valve overlap period. Their results showed that, by decreasing valve
overlapping duration from original 32 crank angle degree (CAD) to 16 CAD, the average emissions of total HCs /CH4
were reduced by approximately 41% and 28% for HCNG30 fuel and CNG fuel respectively.
In many cases, the emissions of HCs and those of NOx exhibit opposite varying trends with engine operating
parameters. For instance, as the engine excess air ratio are increased toward the lean burn limit, HC emissions
dramatically increase while NOx emissions continue to decrease. These opposite trends suggest the well-known trade-
off phenomenon between HC and NOx emissions, as observed by many researchers [125, 129]. Figure 27 presents such
trade-off relationship for a S.I. engine fueled by NG and HCNG, with Euro V and VI emission limits for NOx and CH4
highlighted. As can be seen, simultaneous reduction of both NOx and HC emission need to be achieved for meeting
these emission regulation, which has been proven to be rather difficult if only in-cylinder approaches are used. In this
regard and considering the fact that the conversion efficiency of after-treatment at lean burn condition is relatively low
due to low exhaust temperature, Lim et al.[125] recommended the strategy of retarding ignition to reduce NOx emissions,
combined with a slightly fuel-rich combustion that helped to increase the conversion efficiency of after-treatment system
for HC and CH4 emissions.
53
Fig. 27 The trade-off relationship between NOx and CH4 emissions. Engine speed = 1260 rpm, at half load (575 Nm, half of
full load) and MBT spark timing. Adapted from Lim et al. [125].
It can also be seen from Fig. 27 that the increase of CR has a positive, although relatively small effect on reducing
both NOx and CH4 emissions. At the same CR, the engine fueled with HCNG30 has better emission performance
compared to the NG case, indicating the benefits of hydrogen addition on solving the trade-off problem between CH4
and NOx emissions.
3.3.3 CO emissions
Many experimental investigations have confirmed that hydrogen enrichment helps to reduce CO emissions and the
extent of the reduction increases
with the increase of hydrogen fraction [101, 111, 118, 137]. For example, the results of an experimental work by Park
et al. [112], conducted on an 11L heavy duty S.I. engine at constant engine speed and load with MBT spark timing
condition, showed that CO emissions were decreased by 15.7% and 16.2% compared with NG fueled engine when
fueled with HCNG30 and HCNG40, respectively [112]. The authors attributed the results to the reduced carbon content
and promoted combustion of the mixture with increasing hydrogen fractions. In another study, Mathai et al. [118]
conducted an engine test at constant speed of 3000 rpm, engine load of 3.1 bar (BMEP) and 24°BTDC ignition timing
and the results showed that the CO emission were reduced by 42.4% by 18% of H2 addition in NG. Hydrogen
54
enrichment can increase the combustion temperature and the concentrations of OH radicals (as well as H and O radicals)
[36, 37], both of which can promote the oxidation of CO into CO2 via CO + OH = CO2 + H.
Fig. 28 Brake specific CO emissions versus excess air ratio and hydrogen fractions. Engine speed=1200rpm; MAP=105kPa
and spark timing=30°CA BTDC. Adapted from Ma et al. [91]
As one can expect, CO emissions can be largely affected by oxygen availability. The effect of λ was highlighted here
since it is considered the most influencing factor on controlling CO emissions [149]. Figure 28 presents the experimental
results regarding the effects of λ on CO emissions for S.I. engines fueled by HCNG fuels with various hydrogen fractions,
where the engine speed, load and spark timing were held unchanged. As can be seen, as λ increases from 1.0 to 1.6, CO
emissions decrease noticeably for all fuel cases, due to the increasingly more supplies of oxygen. The formation of CO
is mainly through the incomplete oxidation of hydrocarbons following RH-R-RO2-RCHO-RCO-CO, where R is the
hydrocarbon radical [20, 150]. In HCNG engines, CO is an intermediate product of CH4 oxidation and forms almost
exclusively in the fuel-rich region of the flame, where oxygen is deficient. Adequate oxygen in the fuel-air mixture will
result in CO being converted further to CO2. However, when fuel mixture was further leaned out (λ>1.6), CO emissions
drastically increases. This is caused by the reduced combustion temperature and thus decreased oxidation rates and the
increased probability of large-scale flame quenching [111]. It is worthwhile to mention that increasing hydrogen content
increases the critical value of λ beyond which CO starts to increase dramatically. This result again confirmed the fact
55
that hydrogen addition can strengthen the lean combustion ability, which is beneficial to reduce emission level of HCNG
engine.
3.3.4 Particulates and other unregulated emissions
Generally, the level of particulate emissions is rather low for HCNG engines, as compare to traditional gasoline or
diesel engines. Limited studies conducted by Hora et al. [142] and Singh et al. [151] reported that particle emissions
from HCNG engine were mainly from lubricating oil, which enters the combustion chamber through the gaps between
piston rings and cylinder liner.
Hora et al. [142] reported that hydrogen enrichment in NG engine was beneficial for reducing total particle numbers,
especially at lower load condition. As shown in Fig. 29, at lower engine load (2.98 bar BMEP), HCNG10 and HCNG20
fuel has much lower total particle emissions compared to the NG case. At low engine
Fig. 29 Total particle number emissions of HCNG engine at different hydrogen fractions and engine load condition. Engine
speed=1500rpm, spark timing=20°BTDC. Adapted from Hora et al. [142].
load conditions, NG fuel is more likely to suffer from incomplete combustion which leads to large amounts of particle
emissions. Hydrogen addition in NG contributes to more complete combustion and thus lowers particulates numbers.
However, when adding 30% hydrogen, the contribution of lubricating oil to particulate formation become more obvious
as the heat release rate and pressure rise are significantly increased. At higher load conditions, the effects of hydrogen
56
addition does not result in notable change of particulate emissions. This is because particles are mainly produced by
lubricating oil combustion at higher loads, which is only minimally affected by hydrogen addition. In addition, by
analyzing the particle size distribution of engine-out particulate emissions, the authors reported that the number of nano-
sized particles (diameter DP＜50nm）are quite higher as compared to that of accumulation mode particles (100-300nm)
for all tested HCNG blends with XH2 of 0-30%, although their mass contribution was much smaller. This results was
important since the emissions of small particles pose even greater risks to human health. As recent emission regulations
are starting to limit both the mass and number of the emitted particles, it is now essential to pay more attention to the
particle number emissions characteristics of HCNG engines.
Singh et al. [143] presented an interesting comparative study about the unregulated emissions of a heavy duty SI
engine fueled with CNG and HCNG fuels. The test was conducted at wide open full throttle performance (FTP) cycle
from 1000 rpm to 2400 rpm for a heavy duty six-cylinder engine fueled with HCNG18 fuel and pure NG fuel respectively.
Results showed that compared to pure NG engine, HCNG18 fueled engine has significant advantage in reducing
unregulated emissions like Formic Acid, propane, ethylene, acetylene, benzene and etc. However, other unregulated
emissions such as methanol, ethanol, formaldehyde and acetaldehyde showed a slightly increase for HCNG18 fueled
case.
3.4 Cycle-by-cycle variations and lean burn limit
The preceding discussions of HCNG engine performances have demonstrated the potential benefits of fuel-lean
combustion in improving engines’ fuel economies and emissions. However, ultra-lean combustion may lead to misfire
or large cycle-by-cycle variations (CCV), which refers to the variations of the combustion processes among a statistical
number of engine working cycles. In general, the level of CCV reflects the combustion stability of the engine as well as
the drivability of the vehicle. Severe cyclic variations will adversely affect engine performance and lead to, for instance,
reduced power output, thermal efficiency and increased level of HC/CO emissions. Hydrogen enrichment has been
57
shown to reduce combustion cyclic variations of NG engines [128, 139]. It is relevant to investigate the CCV
characteristics of HCNG engines, with the objective being to extend the engine’s lean burn limit since the advantage of
hydrogen addition in NG engines come mostly from its effects on extending the engines’ lean burn capabilities [97].
In order to quantitatively assess CCV, the coefficient of variation (COV) in terms of IMEP is frequently used. COVIMEP
is defined as the ratio of standard deviation to the arithmetic mean of IMEP for a number of consecutive engine cycles
and can be calculated as follows:
(10)
where, is the arithmetic mean and is the standard deviation of IMEP over N engine cycles, which can be
expressed as:
(11)
where is the measured IMEP for a certain engine working cycle. In general, reliable determination of COV requires
experimental cylinder pressure data of a statistically enough number of engine cycles.
A number of researchers have confirmed the positive effects of hydrogen addition on reducing the COVIMEP of NG
engines [95, 98, 112, 113, 124] (e. g. as seen in Fig. 31). To understand the underlying promoting effect of hydrogen,
Wang et al. [152-154] studied the effects of hydrogen addition on the combustion and CCV characteristics of natural
gas in a constant volume vessel. They found that cyclic variations are initiated at the initial flame development stage,
and hydrogen addition can significantly enhance the stability of this early combustion process thus contributing
positively to the stability of the subsequent flame propagation and the overall combustion process. In addition, the
duration of both the initial combustion and the flame propagation stages can be decreased by hydrogen addition, again
helpful to inhibit cyclic variations. These explanations are consistent with those of Ma et al. [95], who conducted relevant
experiments in a 6-cylinder heavy duty HCNG engine and observed a decrease in the duration of the in-cylinder
combustion process by hydrogen addition. As shown in Fig. 30, 20% hydrogen addition obviously shortens both the 0-
10% burn duration (defined as the crank angle between spark ignition and 10% mass fraction burned, reflecting the
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initial flame development period) and 10-90% burn duration (reflecting the main flame propagation period). Moreover,
the effects of hydrogen addition on reducing the combustion duration were seen to be more pronounced at leaner fuel/air
mixture. These results were considered to be responsible for decreased CCVs.
Fig. 30 Effect of hydrogen addition on combustion duration at various excess air ratios. (a) 0-10% burn duration and (b)10-
90% burn duration, obtained at MBT spark timing, engine speed of 1600 rpm and 0.55MPa (BMEP) load condition. Data
from Ma et al. [95].
In another study [121], Wang et al summarized the underlying mechanisms of hydrogen’s effects on improving
combustion stability and decreasing CCV as follows: i) hydrogen addition enhances the initiation of the flame kernel
and therefore reduces the uncertainties of flame initiation; ii) the high laminar flame speed of hydrogen (approximately
8 times higher than that of methane) reduces the duration of the flame propagation process and this also contribute to
reduced COV ; iii) The shorter quenching distance of hydrogen makes flame propagate much close to cylinder wall,
result in more complete combustion; iv) wider flammability of hydrogen facilitates lean combustion limit. In addition,
better homogenization of the fuel due to the higher diffusivity of hydrogen as compared to methane also help to stabilize
the combustion and reduce CCV.
Besides hydrogen fraction, the cyclic variations of HCNG engines also depend considerably on engine working
conditions such as excess air ratios, ignition timing and engine speed. The effects of excess air ratios on the COVIMEP of
HCNG engines have been studied by Wang et al. [98] and the results are shown in Fig. 31. As can be seen, the addition
59
of hydrogen in the fuel helps to decrease the COVIMEP at a fixed λ and the effects become more pronounced at leaner
conditions. Similar results are also obtained by Ma et al. [33, 91, 107].
Fig. 31 COVIMEP versus excess air ratio and hydrogen fraction. Engine speed=2000 rpm, BMEP=0.16 MPa, ignition timing=
30°CA BTDC. Adapted from Wang et al. [98]
The fact that hydrogen addition reduces the cyclic variations more significantly at leaner condition [95, 154] makes
hydrogen a particularly attractive fuel additive for ultra-lean burn engines, as suggested by Ma et al.[95]. As a matter of
fact, the lean operational limit (LOL, in terms of excess air ratio) beyond which the engine cannot operate stably, are
closely related to the engine’s cycle-by-cycle variations. Many studies indeed have defined the excess air ratio at which
COV reached 10 % as the engine’s LOL.
The effects of engine operating parameters on the LOL of HCNG engines have been studied by Ma et al. and Wang
et al. [97, 121]. These results showed that LOL increased with the increase of engine load, which can be attributed to
the reduced residual gas fraction, less pumping loss and enhanced volumetric efficiency at higher engine load conditions.
With respect to the effects of spark timing, Wang et al. [121] concluded that there existed an optimum value for LOL
extension at a given engine speed/load condition, as evident in Fig. 32. This can be explained as follows: if the ignition
is too early, then due to inadequate compression, the temperature and pressure would be relatively low at the instant of
ignition, which has a negative effect on the initial flame development and thus is detrimental to the stability of the
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following flame propagation process, resulting in increased CCV; On the other hand, if the ignition is too late, the
delayed combustion would cause the fuel’s chemical energy to release late in the expansion stroke, which would result
in a prolonged combustion duration and thus higher CCV. In addition, retarded ignition timing would also increase the
heat transfer loss. This, in fact, indicated more energy is taken away by the cooling fluid, which contributes to the
weakening of the combustion intensity and the decrease in combustion stability [121]. As a result, neither over-retarded
nor over-advanced ignition timing is favorable for LOL extension.
Fig. 32 Effect of ignition timing on LOL at different engine load and engine speed conditions. Reproduced From Wang et al.
[121]
The influence of engine speed on the LOL of HCNG engines was more complicated and may change with
experimental conditions. For example, Moreno et al. [122] reported that at full load conditions, COVIMEP become smaller
as engine speed increased. On the other hand, Wang et al. [121] observed an opposite trend at low engine load conditions
(10% and 20% open throttle ). Ma et al. [97] concluded that the influence of engine speed on lean burn limit depends
on engine load, as evident in Fig. 33, where the variations of LOL with engine speed were shown for different engine
load conditions.
61
Fig. 33 Variation of lean burn limit versus engine speed and hydrogen fraction. Operated at MBT spark timing and
MAP=50KPa (a) and MAP=105KPa (b) respectively. Adapted from Ma et al. [97].
At low engine load, the in-cylinder turbulence intensity is low while the residual gas fraction is high. Therefore, under
such condition the increase in engine speed would help to enhance the in-cylinder flow motion and improve the gas
exchange efficiency, leading to a faster combustion process and extended LOL, as can be seen from Fig. 33a for the 0
and 10% H2 case. On the other hand, at high engine load, the flow motion and gas exchange efficiency are already high
and thus less dependent on engine speed (Fig. 33b) and we can even see a slight decrease of LOL with engine speed for
0 and 10% H2 addition cases at this load condition. Ma et al. believed this was caused by the strong turbulence at higher
engine speed, which may blow out the flame kernel and lead to misfire and combustion instability. It can also be noted
that the LOL did not change much with engine speed if more than 30% of hydrogen was added, which is due to the fact
that the enhanced combustion reactivity by hydrogen enrichment makes combustion stability less affected by turbulence
intensity and residual gas fractions.
The effects of a number of other engine operating parameters on LOL were also studied and the results showed that
increased temperature of coolant and lubricant oil [121], optimization of fuel injection [138] can also extend LOL at
certain operating condition.
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4. Numerical studies on HCNG engine performances and emissions
The experimental results presented in the last section provide rich knowledge about the operating characteristics of
HCNG engines, which contribute to provide important data based on which HCNG engine design parameter and control
strategy can be optimized for better thermal efficiencies and lower pollutant emissions. Numerical simulations, with
much reduced research costs, can provide effective complements for a deeper understanding of HCNG engine
performance and emission characteristics. Depending on the research objectives, two types of numerical engine models,
i.e. the thermodynamic model and the multi-dimensional CFD model, are generally used.
Thermodynamic models, including both the zero-dimensional single-zone model and multi-zone (mostly two-zone)
model, are generally used to conduct parameter optimization, with emphasis on the overall engine performance while
neglecting the details of in-cylinder combustion process [155]. Due to its computation flexibility and simplicity,
thermodynamic models have been widely adopted to predict overall engine combustion and emission performance. For
example, Djouadi et al. [156] developed a zero-dimensional single-zone model and investigated the combustion
behavior of HCNG15 fueled SI engine. The bulk cylinder pressure and temperature, the mass burned fraction and the
heat release rate as a function of crank angle were obtained. Chugh et al. [157] studied the effects of CO2 blending on
SI engine emissions fueled with CNG and HCNG18 by developing an Otto cycle model. One-step global reaction: Fuel
+ (O2 + 3.76 N2) → CO2 + H2O + N2 + O2 + CO + H2 + NO + CH4 was adopted to describe the combustion chemistry.
The results showed that by adding 6% CO2 to the HCNG18 fuel, both lower NO emissions and higher thermal
efficiencies can be achieved. Compared to above single-zone model [156, 157], two-zone models are more widely
adopted to achieve better simulation accuracy. These models generally have the following assumptions [158]:
(a) The combustion chamber is divided into two parts: burned products and unburned mixture.
(b) Each zone has uniform properties and the working fluid is recognized as ideal gas.
(c) The pressure at any time is uniform in the cylinder.
(d) Heat transfer takes place only through external surfaces of the combustion chamber. There is no heat transfer
63
between the burned and unburned region.
(e) Crevice effects, flame thickness and fluid velocity are ignored. The chemical composition of the unburned
charge is assumed to be frozen.
Then, by incorporating mass and energy conservation equations and the gas state equations (PV=nRT) for both burned
and unburned region, in-cylinder pressure, temperature and mixture component history can be derived through Eq. (12)-
(14):
(12)
(13)
(14)
In the above equations, subscripts u and b indicated unburned and burned region, P, V, and T refer to the in-cylinder
pressure, temperature and volume, respectively. Cp and Cv are the constant-pressure specific heat and constant-volume
specific heat, respectively. R is the gas constant and μ refers to the internal energy. Eqs. (12) - (14) constitute the basis
for both zero-dimensional and quasi-dimensional two-zone modeling studies. Heat transfer loss dQ/dθ and mass burning
rate dmb/dθ are important sub-models to close the above equations. These can be either correlated from empirical
formulas or obtained from experimental data.
The modeling procedures of two-zone thermodynamic model are similar among different studies and detailed
derivation could be found in Refs [158-160] with entrainment combustion model and Ref [161] with a fractal-based
turbulent combustion model, depending on the modelling approach of the turbulent mass burning speed. Additional
pollutant formation sub-models can be included for predicting engine emissions. The extended Zeldovich mechanism:
N + O2 NO + O; N + NO N2 + O; N + OH NO + H and reaction: CO + OH CO2 + H are mostly used to
describe NO and CO formation chemistry, respectively, as shown in Ref [162]. Following the above modeling steps,
many researchers have developed two-zone models, including both zero-dimensional [163-166] and quasi-dimensional
64
thermodynamic models [167-169], to conduct parameter study and predict HCNG engine performance, which will be
presented in the following section.
Fig. 34 Predicted fuel consumption and NOx emissions of HCNG engine with different hydrogen fractions and EGR ratio
over NEDC. Adapted from Mariani et al.[165].
For example, based on a zero-dimensional two-zone model combined with the extended Zeldovich mechanism,
Mariani et al.[165] studied the effects of hydrogen addition on HCNG fuel economy and NOx emissions. We can see
from Fig. 34a that, as compared to the natural gas fueled case, the engine fuel consumption reduced by 2.5%, 4.7% and
5.7% for cases of HCNG blends with 10%, 20% and 30% of hydrogen, respectively. Besides, with 10% EGR ratio, the
fuel consumption was further reduced which was mainly due to lower combustion temperature and thus reduced heat
loss to the cylinder walls. It can been seen from Fig. 34b that although NOx emissions increased when XH2 become
larger, 10% EGR ratio could result in 85% lower NOx emissions at each HCNG blends. These results thus showed that
EGR can both improve engine thermal efficiencies and reduce NOx emissions for HCNG engines.
Utilizing a zero-dimensional two-zone model, Ma et al. [164] studied the influence of combustion phasing and
combustion duration on HCNG engine thermal efficiency. The results show that there always exists an optimal
combustion phasing for best engine thermal efficiency. The results also showed although hydrogen addition could
65
decrease combustion duration which was beneficial to thermal efficiency because of the improvements in the degree of
constant volume combustion, hydrogen addition could also increase the engine heat loss and this may cancel out the
benefits of constant volume combustion. Morrone et al. [166] also investigated the characteristics of engine thermal
efficiency with different HCNG blends and engine load condition, based on a zero-dimensional two-zone model. They
reported that the engine brake efficiency showed no change when fueled with HCNG and CNG fuel, provided the same
ignition timing was used. However, when operating at MBT spark timing condition, as shown in Fig. 35, both HCNG10
and HCNG30 engine had higher engine thermal efficiency as compared to CNG cases. Furthermore, the increment was
more obviously at lower engine loads.
Fig. 35 Engine efficiency increment compared to base CNG engine at varying load conditions. Adapted from Morrone et al.
[166]
Tinaut et al. [167] developed a quasi-dimensional two-zone model to investigate the effects of hydrogen fraction
(varying between 0%, 3%, 6% and 15%) on engine combustion and NO, CO emissions. The extended Zeldovich
mechanism as well as CO + OH CO2 + H were included to describe NO and CO formation chemistry, respectively.
The results showed that NO emissions increased with hydrogen fraction due to elevated combustion temperature after
hydrogen addition. CO emissions showed no change with hydrogen fraction, which was due to the lean operating
conditions at which CO can be largely oxidized.
Afshari et al. [170] proved the capability of their quasi-dimensional two-zone model in predicting engine
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performances by validating with related experimental data. The NO emission characteristics of HCNG engine under
different hydrogen fractions and equivalence ratios were then studied and the results are shown in Fig. 36. As can be
seen in Fig. 36a, NO production rate increased with the increase of XH2 and the increment was more noticeable at the
peaks, as compared to the end of the expansion stage. From Fig. 36b, it was concluded that NO production exhibited
large dependence on equivalence ratios, with its maximum value being at lean condition (Φ = 0.9). It rapidly decreased
as the equivalent ratio increased to 1.0 and 1.1. Theoretically, NO production should peak at Φ = 1.0 since stoichiometric
flame has the highest heat release, while the authors explained that both extra oxygen and relatively high combustion
temperature at Φ = 0.9 lead to more NO production.
Fig. 36 Evolution of NO emissions with crank angle during an engine working cycle for different hydrogen fractions (Fig.
36a) and equivalence ratio (Fig. 36b). Reproduced from Afshari et al. [170].
The effects of partially stratified charge (PSC) strategy on SI engine fueled with different HCNG blends was simulated
with quasi-dimensional two-zone models [169]. PSC method, through creating richer mixture around spark plug, was
proposed to further improve combustion stability and thermal efficiency of SI engine at lean conditions.
67
Fig. 37 Indicated thermal efficiency versus hydrogen fraction for various implementations of the PSC strategies. Adapted
from Aliramezani et al. [169]
As can be seen in Fig. 37, engine thermal efficiency was improved by adopting PSC strategy for all fuel cases. This
was because PSC method could reduce the ignition delays and improve the combustion speed, especially for the early
flame period and thus contributed to enhance the flame stability and shorten the combustion duration. Regarding the
effect of hydrogen fraction, it can be seen that increasing XH2 also help to improve thermal efficiency which may
similarly be attributed to the faster flame speed and shorter combustion duration after hydrogen addition.
Facilitated by the rapid improvements of computational power, multi-dimensional CFD models have been
developed to study HCNG engine combustion process, performances and emissions [171-176]. Compared with
thermodynamic models, CFD models provide a detailed insight into the combustion process that the spatial distribution
of in-cylinder flow velocities, temperatures, pressures and chemical compositions can be obtained. In CFD models, the
combustion chamber geometry effect can be considered considered and detailed chemical kinetic models [177, 178] can
be included for a better understanding of the detailed combustion process and pollutant formation mechanisms. For
example, utilizing the commercial CFD package AVL FIRE with GRI-Mech 3.0 kinetic mechanism, Zaker et al. [172]
studied the effect of hydrogen fractions (0-50% in volume) on the in-cylinder combustion process as well as the engine
emissions and thermal efficiency characteristics. Fig. 38 shows the simulated velocity and temperature fields inside the
cylinder during the intake process. Such type of information contribute to our deep understanding of the in-cylinder
68
flow and combustion behaviors, which have major impacts on the overall engine performances.
Temperature contours
Velocity Vectors
(a) 10 degrees ATDC (b) 25 degree ATDC (c)70 degrees ATDC
Fig. 38 In-cylinder velocity vectors (top) and temperature contours (bottom) at different crank angles during intake process.
From Zaker et al. [172]
Wang et al. [174] developed a multi-dimensional model by integrating the KIVA-3V CFD code with the CHEMKIN
chemical kinetic package to study HCNG engine performances. The computed results well reproduced the experimental
data on NO and CO emissions of engine fueled with HCNG20 and CNG under the specified operating condition.
Similarly, by using the KIVA-3V code combined with detailed chemical kinetics, Gharehghani et al. [173] investigated
the effects of hydrogen enrichment on HCNG engine operating range and combustion characteristics. As shown in Fig.
39, hydrogen addition could extend the lean operation limit of CNG engine. The LOL extended by about 20% and 30%
by adding 30% and 50% hydrogen, respectively. This can be attributed to the enhanced flame propagation speed with
hydrogen addition. As demonstrated in Fig. 40, with higher XH2, the maximum heat release rate and in-cylinder pressure
approaches closer to the TDC and the heat release duration become smaller, which indicated increased combustion speed
after hydrogen addition. Besides, the author also reported that although hydrogen addition led to higher NOx emissions
at given equivalence ratio, NOx emission of HCNG engine can still be maintained at acceptable level by operating at
extended lean limit condition.
69
Fig. 39 Lean operation limit (defined as equivalence ratio when COVIMEP reaches upper limit of 10%) with different hydrogen
fractions. Adapted from Gharehghani et al. [173].
Fig. 40 In-cylinder pressure and heat release rate (HRR) for various hydrogen fraction at equivalence ratio Ф = 0.625. Adapted
from Gharehghani et al. [173].
The prediction of NOx emission for HCNG engines using CFD models received significant research interests. For
instance, using the KIVA-3V CFD code, Yoo et al. [175] showed that NOx emissions can be reduced through extended
lean burn limits after hydrogen addition. Their simulation results indicated that with 10% increase in λ, the NOx emission
could be reduced by 50%, suggesting the great benefits of lean burn technology on reducing NOx emissions. Wang et
al. [179] investigated the effects of EGR on engine NO emissions with the commercial CFD code AVL FIRE. The results
70
showed that NO emissions decreased obviously as EGR ratio increased, which largely decreased the in-cylinder gas
temperature.
Michikawauchi et al. [110] conducted a CFD simulation to investigated the effect of fuel injection timing (port
injection) on SI engine NOx emissions. The results showed that injection during the intake stroke (-300 degrees ATDC)
could decrease NOx emissions, compared to fuel injection in the expansion stroke (-720 degrees ATDC) at the specified
condition. The author explained that fuel injection during the intake stroke could create a leaner air-fuel mixture near
the spark plug, as clearly shown in Fig. 41. The air/fuel mixture near the spark plug is leaner than that around this area
for intake-stroke fuel injection, while the opposite holds for the expansion-stroke fuel injection case. With leaner mixture
near the spark plug, the combustion gas temperature was lower and therefore less NOx was formed.
(a) -300 degrees
ATDC (b) -720 degrees
ATDC
Schematic of center section in
combustion chamber
Fig. 41 Effect of fuel injection timing on excess air ratio distribution (right) in the depicted center section of combustion
chamber (left). Computed for pure methane at λ=1.55, engine speed of 1000 rpm and engine load of 50 Nm. From
Michikawauchi et al. [110].
In order to further investigate the NOx emission characteristics of HCNG engine, Kosmadakis et al. [147, 180]
assessed the contribution of different NO formation mechanisms to NOx emissions in HCNG engines. CFD code was
used to study the in-cylinder spatial NOx distribution and production rate. Four different NO formation mechanism, i.e.
thermal NO reaction mechanism, NNH route, N2O route and the prompt NO mechanism were considered. As suggested
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by the author, prompt NO formation pathways was negligible for all test case and can be omitted from the combustion
model. The N2O mechanism and NNH pathway become important at low combustion temperature and thus have
relatively large contribution to total NO emissions at lean condition (Φ=0.7 and Φ=0.8), as shown in Fig. 42. Thermal
NO mechanism is dominant for all test cases especially at high-temperature combustion conditions. The author
suggested that more NO formation routes should be included (except for prompt route) for more reliable prediction of
NO emissions.
Fig. 42 Contribution of four different NO mechanism in total NO emissions for HCNG30 fuel blends. Adapted from
Kosmadakis et al. [180].
Besides the preceding numerical studies which mainly dealt with engine performance and emissions, Bauer et al.
[181] conducted an interesting simulation of vehicle driving cycle for both highway and urban driving conditions. A
kinematic vehicle model was employed and together with the engine model, which was derived based on corresponding
engine experimental testing (see Bauer et al. [182] ) on power performances, fuel consumptions and emissions at varying
operating conditions. Various operating schemes shown in Table 5 were investigated. The simulation results validated
the effectiveness of hydrogen addition in reducing fuel consumption and pollutant emissions in vehicle driving cycles.
In addition, when determining the optimum XH2, Scheme 2 was recommended as the best method with XH2 in the range
of 11-31% for the highway cycle and 11-38 % for the urban cycle was, based on the criterion of maximizing the driving
range and minimizing total CO2 emissions for a given fuel storage capacity.
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Table 5. Operating conditions of different schemes for vehicle driving simulation
Testing schemes
Engine operating conditions
Scheme 1
Φ=1.0; XH2 = constant; Engine load: 20%-100%;
Scheme 2
Burn limit ≤ Φ ≤ 1.0; XH2 = constant; Engine load: 20%-100%;
Scheme 3
Burn limit ≤ Φ ≤ 1.0; XH2: 0-60%; Engine load: 20%-100%;
5. IC engines fueled with other hydrogen enriched fuels
As presented in the preceding sections, hydrogen has many attractive combustion properties which help to improve
engine thermal efficiency and emissions performance. These remarkable properties make hydrogen an ideal engine fuel
additive. Besides the blends of hydrogen and NG, recent years have also witnessed the development and application of
hydrogen enrichment in other fuels, e.g. gasoline, diesel, biogas, methanol, ethanol and etc. In this section, we intend to
provide a brief summary of the effects of hydrogen addition on the performance and emissions of engines fueled with
the above fuels.
5.1 Hydrogen-gasoline mixture as a fuel
The addition of hydrogen in conventional gasoline engines was extensively investigated in recent years. A number of
researchers have demonstrated the effectiveness of hydrogen addition in improving gasoline engine thermal efficiency
and reducing its pollutant emissions [183-188]. For example, Ji et al. [188] reported that the average brake thermal
efficiency increased from 25.12% of the original gasoline engine to 28.35% after the doping of 3% (in volume) hydrogen.
The author explained that hydrogen addition enhanced the combustion speed of fuel mixture and helped to achieve a
working cycle closer to the ideal constant-volume one. Besides, the high diffusion coefficient of hydrogen contributed
to the improved homogeneity of the gasoline vapor with air, which was beneficial for a more complete combustion and
thus a higher thermal efficiency [188].
73
Improved lean burn performance is another major benefit of hydrogen addition in gasoline engines. Due to the
relatively narrow flammability and slow combustion speed, gasoline engines generally suffer from high cyclic variations
at lean conditions. As hydrogen is featured by high burning velocities and wide flammability limits, its blending with
gasoline can improve the combustion performance at lean conditions. Related research results [189-192] have showed
decreased cyclic variations, improved thermal efficiency and reduced pollutant emissions of SI gasoline engine after
hydrogen addition at lean operating conditions. The effects of hydrogen addition on gasoline engine lean burn
performances at low load [193], varying spark timing [194] and idling condition [195, 196] were also investigated. The
results showed that engine indicated thermal efficiency improved [195, 196], HC, CO and even NOx emissions were
decreased after hydrogen enrichment at idle or lean conditions [196]. Considering gasoline engines have higher thermal
efficiencies and low emissions at WOT conditions, Wang et al. [197, 198] investigated the effect of hydrogen addition
on gasoline engine performance and emissions at lean and the WOT conditions. The experiment results showed that
HCs, CO and particulates were effectively reduced and thermal efficiency improved after the hydrogen blending
compared with the original gasoline engine. Moreover, load control of the engine can be realized at WOT condition by
adjusting excess air ratio due to the extended lean burn limit and combustion stability at lean condition with hydrogen
blending.
Besides lean combustion, oxygen-rich combustion is another effective method to realized high combustion efficiency
and low carbon emissions. As hydrogen and oxygen mixture can be produced simultaneously by electrolysis, there is a
possibility of removing hydrogen storage device on vehicles [199]. Motivated by its promising application, the effect of
hydrogen-oxygen mixture on gasoline engine performance and emissions was investigated [199-201]. The results
showed that, compared with the original gasoline engine, the brake power and thermal efficiency were increased and
total HCs and CO emissions were decreased while NOx also increased when fueling with hydrogen-oxygen mixture
[200, 202]. Moreover, at lean operating conditions, it had higher thermal efficiency and BMEP than both of the original
and hydrogen-blended gasoline engines [201].
74
5.2 Hydrogen-diesel mixture as a fuel
Compression ignition (CI) diesel engines are characterized by their high thermal efficiencies, but also high PM and
NOx emissions due to their non-premixed combustion mode [203]. Besides commercial after-treatment device such as
DPF or SCR, researchers also paid much attention to the optimization of the combustion processes in diesel engines,
aiming to minimize the in-cylinder pollutant formation. Blending diesel with gaseous fuels such as natural gas [204],
hydrogen or HCNG blends was demonstrated an effective approach to reduce pollutant emissions. For example,
Pichayapat et al. [205] showed that, compared with diesel-only operation, the average CO emission and HC emission
decreased by 12.97% and 15.84% respectively, while NOx emission decreased by 1.16% and PM emission decreased
by 9.14% with the diesel-HCNG dual fuel mode at an engine speed range of 800-4000 rpm [205].
Various other researchers have also studied the performance and emission characteristics of engines with hydrogen-
diesel dual fuel operation [206-215]. Research results generally showed that, after hydrogen enrichment, CO, CO2 and
smoke emissions were decreased obviously [212-215]. For example, Sandalcı et al. [214] found that, at a constant engine
speed of 1300 rpm and 5.1kw indicated power condition, the indicated specific smoke emission values with 16%, 36%
and 46% hydrogen energy content decreased by 57.8%, 70.4% and 75.2%, respectively, compared to that with diesel-
only operation. This was caused by the reduced carbon input and enhanced mixing and combustion process after
hydrogen enrichment. However, these studies also observed increased NOx emissions after hydrogen addition [212-
215]. To deal with this problem, EGR method coupled with hydrogen-diesel dual fuel combustion was recommended to
achieve both low NOx emissions and PM emissions [207, 208].
It is worthwhile to mention that the effect of hydrogen addition on diesel engine performance and emission
characteristics also depend on various operating conditions such as excess air ratio and engine load. The related details
has been extensively discussed in a recent review [216], with the main conclusion being that CO2, CO, particulate matter
emissions are generally reduced after hydrogen addition. It was also concluded that the brake thermal efficiency
increased with hydrogen fraction due to the increased heat release rate and cylinder pressure [216]. However, if the fuel
75
injection timing was not adjusted, the engine thermal efficiency may decrease after hydrogen addition, as shown in [213-
215].
5.3 Biogas and oxygenated fuel
As a promising renewable and sustainable fuel, biogas can be produced from rich biomass resource such as
biodegradable wastes and other cheap agricultural crops. Biogas has extensive utilization pathways as pointed out by
Budzianowski [217]. For example, biogas can be used as gas engine fuel for power generation and heat production [217,
218]. It can also be utilized as alternative transport fuel for automotive application. However, the large percentage of
CO2 content in biogas leads to lower burning speeds and narrower flammability limits as compared to NG [219]. This
suggests that the operation of ICEs fueled with biogas may suffer from large cyclic variations and thus combustion
instabilities. Blending hydrogen with biogas can significantly enhance the combustion stability of biogas flame [220]
and engine work stability [221]. As reported by Porpatham et al.[222], compared to the original biogas fueled engine,
HC emissions reduced obviously by hydrogen addition. Besides, the brake power and brake thermal efficiency increased
for XH2 of 5-15%. However, Park et al. [223] showed that the increased trend of thermal efficiency was limited to small
hydrogen fraction range. As a result, 5%-10% hydrogen fraction was recommended to achieve maximum thermal
efficiency considering the trade-off between enhanced combustion stability and increased cooling loss with increasing
hydrogen fraction.
Biogas-derived syngas, consisting of mainly CO and hydrogen, is another potential gaseous fuel for engines. Hagos
[225] showed that pure syngas was a good substitute for gaseous fossil fuels such as NG in direction injection SI engines.
Due to the increased hydrogen fraction and reduced CO2 content in syngas mixture, syngas-fueled SI engine have better
combustion behavior as compared to biogas [226, 227]. The effect of syngas component variation on fuel oxidation
kinetics [228] and the effect of hydrogen fraction on syngas fuel properties and combustion behavior as well as their
interactions with engine thermal efficiency were studied by Shivapuji et al. [229]. The results showed that the brake
thermal efficiency firstly increased with hydrogen fraction but then decreased for higher hydrogen levels due to
76
increasing engine cooling load.
Alcohol fuels such as methanol and ethanol are renewable and clean alternative-fuels for SI engines. A recent review
has systemically discussed the application of methanol in SI engine [3]. Experimental results showed that hydrogen
addition improve the thermal efficiency and emission performance of methanol engines [230, 231]. Hydrogen addition
enhanced the homogeneity of fuel-air mixture since the high latent heat of methanol make the vaporization and mixing
process more difficult, especially at low cylinder temperature condition such as cold start, low load and idling conditions
[232]. Research results showed that HC and CO emissions at cold start condition [233], idling condition [234] and part
load condition [232] were effectively reduced with hydrogen addition. Specifically, Zhang et al. [233] reported that HC
and CO emissions during cold start period (within 19 s) were reduced by 68.7% and 75.2% respectively with hydrogen
addition. The results in Refs [232, 234] reported reduced fuel consumption and increased thermal efficiency after
hydrogen addition, compared to pure methanol engine.
Similar conclusions were also obtained for ethanol-fueled SI engines. Wang et al. [235] reported that the indicated
thermal efficiency of ethanol engine was increased by 4.44% at idle and stoichiometric conditions, with 6.38% addition
of hydrogen. In general, hydrogen addition also contributes to reduce HC and CO emissions but may increase NOx
emissions [236]. Lean combustion strategy was suggested to reduce the NOx emission level. As reported by Greenwood
et al.[237], due to the enhanced lean burn performance with 15% and 30% hydrogen addition, the NOx emission of
ethanol-hydrogen blends fueled engine can be reduced by more than 95%, as compared to that in stoichiometric gasoline
operation.
6. Concluding remarks
Recent years have witnessed significant technological developments of HCNG engines towards their widespread
applications in the transportation sector. This review provides a comprehensive summary of recent research progress in
both the fundamental combustion properties of HCNG and the effects of its application on SI engine’s combustion and
77
emission performances. The important conclusions are summarized in the following:
(1) Hydrogen has lower MIE, shorter ignition delay and higher laminar flame speed, which contribute to the enhanced
initial flame kernel development and flame propagation process inside the cylinder. As a result, more complete
combustion and reduced cyclic variations can be achieved for HCNG engines. Hydrogen addition can extend the
lean operation limit of SI engines and this is considered to be one of the most important advantages for hydrogen
addition, as improved ultra-lean combustion performance can realize both high engine thermal efficiency and low
NOx emissions.
(2) The kinetic promoting effect of hydrogen addition on methane oxidation can be attributed to the increased
concentrations of OH, O and H radicals since hydrogen addition enhance the reaction rates of O + H2 H + OH
and H + O2 OH +O. This fact is responsible for the reduced ignition delays and improved laminar flame speed
of CH4/H2/air mixture after hydrogen addition.
(3) In general, compared to NG engine, hydrogen enrichment in NG could reduce engine HC, CO2 and CO emissions,
but may lead to higher NOx emissions. The reduction of carbon emissions can be attributed to both the lower carbon
input in HCNG mixture and the more complete combustion, since hydrogen has higher laminar flame speed and
lower quenching distance; while the increase of NOx emissions is mainly because of elevated combustion
temperature which increases the NOx formation rate through thermal NO routine. Delayed ignition timing and lean
combustion are recommended to regulate NOx emissions level.
(4) The effects of hydrogen addition on fuel economy depend on the balance between improved combustion efficiency
and increased cooling loss to cylinder wall. In general, compared to pure NG engine, hydrogen enrichment has
higher engine thermal efficiency at ultra-lean condition while the effects can be negligible near stoichiometric
condition. In order to maximize the thermal efficiency of HCNG engine, the operating parameter and design
parameter such as spark timing and CR should be optimized.
(5) Hydrogen content is the most important parameter that influence the HCNG combustion and emission performance.
78
HCNG engine fueled with a higher hydrogen fraction may cause abnormal combustion such as pre-ignition, knock
and backfire, while with a relatively lower hydrogen content, the benefits of hydrogen cannot be utilized fully. In
general, HCNG18, HCNG20 and HCNG30 are most widely adopted fuels in the open literature.
(6) Hydrogen blended with other transport fuels such as gasoline, diesel, biogas and alcohols was also beneficial for
improving engine in-cylinder combustion process and emissions performance.
Acknowledgements
This work was supported by the Fundamental Research Funds for the Central Universities with grant number
2016IVA038.
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