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Korakianitis2011-NaturalGasFueledSparkIgnition.pd+
Progress in Energy and Combustion Science 37 (2011) 89e112
Contents lists available at ScienceDirect
Progress in Energy and Combustion Science
journal homepage: www.elsevier.com/locate/pecs
Review
Natural-gas fueled spark-ignition (SI) and compression-ignition (CI) engine
performance and emissions
T. Korakianitis*,1, A.M. Namasivayam, R.J. Crookes
School of Engineering and Materials Science, Queen Mary University of London, Mile End Road, London E1 4NS, United Kingdom
a r t i c l e i n f o
a b s t r a c t
Article history:
Received 21 August 2009
Accepted 22 April 2010
Available online 8 June 2010
Natural gas is a fossil fuel that has been used and investigated extensively for use in spark-ignition (SI) and
compression-ignition (CI) engines. Compared with conventional gasoline engines, SI engines using natural
gas can run at higher compression ratios, thus producing higher thermal efficiencies but also increased
nitrogen oxide (NOx) emissions, while producing lower emissions of carbon dioxide (CO2), unburned
hydrocarbons (HC) and carbon monoxide (CO). These engines also produce relatively less power than gasoline-fueled engines because of the convergence of one or more of three factors: a reduction in volumetric
efficiency due to natural-gas injection in the intake manifold; the lower stoichiometric fuel/air ratio of natural
gas compared to gasoline; and the lower equivalence ratio at which these engines may be run in order to
reduce NOx emissions. High NOx emissions, especially at high loads, reduce with exhaust gas recirculation
(EGR). However, EGR rates above a maximum value result in misfire and erratic engine operation. Hydrogen
gas addition increases this EGR threshold significantly. In addition, hydrogen increases the flame speed of the
natural gasehydrogen mixture. Power levels can be increased with supercharging or turbocharging and
intercooling. Natural gas is used to power CI engines via the dual-fuel mode, where a high-cetane fuel is
injected along with the natural gas in order to provide a source of ignition for the charge. Thermal efficiency
levels compared with normal diesel-fueled CI-engine operation are generally maintained with dual-fuel
operation, and smoke levels are reduced significantly. At the same time, lower NOx and CO2 emissions, as well
as higher HC and CO emissions compared with normal CI-engine operation at low and intermediate loads are
recorded. These trends are caused by the low charge temperature and increased ignition delay, resulting in low
combustion temperatures. Another factor is insufficient penetration and distribution of the pilot fuel in the
charge, resulting in a lack of ignition centers. EGR admission at low and intermediate loads increases
combustion temperatures, lowering unburned HC and CO emissions. Larger pilot fuel quantities at these load
levels and hydrogen gas addition can also help increase combustion efficiency. Power output is lower at
certain conditions than diesel-fueled engines, for reasons similar to those affecting power output of SI engines.
In both cases the power output can be maintained with direct injection. Overall, natural gas can be used in both
engine types; however further refinement and optimization of engines and fuel-injection systems is needed.
Ó 2010 Elsevier Ltd. All rights reserved.
Keywords:
Natural gas
Biogas
Methane
Diesel
Biodiesel
Alternative
Fuel
Emulsion
EGR
Hydrogen
Engine
Performance
Emissions
Spark
Ignition
Compression
Contents
1.
2.
3.
Document layout . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 90
Introduction . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 90
2.1.
Natural gas and biogas as part of the general fuel supply . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 90
2.2.
Fuel and engine performance parameters . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 93
2.2.1.
Analysis of cylinder-pressure data . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 95
Natural gas in spark-ignition engines . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .97
3.1.
Performance . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 97
3.2.
Exhaust emissions . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 99
3.3.
Natural gas direct injection in SI engines . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 100
* Corresponding author. Tel./fax: þ44 207 882 5301.
E-mail address: [email protected] (T. Korakianitis).
1
Email forward for life [email protected]
0360-1285/$ e see front matter Ó 2010 Elsevier Ltd. All rights reserved.
doi:10.1016/j.pecs.2010.04.002
90
4.
5.
6.
T. Korakianitis et al. / Progress in Energy and Combustion Science 37 (2011) 89e112
Natural gas in compression-ignition engines . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .101
4.1.
Performance . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 102
4.2.
Exhaust emissions . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 104
4.3.
Dual-fuel CI operation with natural gas and alternative pilot fuels . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 106
4.4.
Natural gas direct injection in CI engines . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 108
Summary and conclusions . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .109
5.1.
Suggested future work . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 110
Nomenclature . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 110
References . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 111
1. Document layout
A detailed review of previous work investigating natural gas as
a fuel in reciprocating piston engines, both spark-ignited (SI) and
compression-ignited (CI), is presented. Previous reviews (such as
reference [1]) only discussed SI engines, while this paper includes
CI engines in dual-fueling. In addition, in reference [1] there is little
discussion of exhaust emissions, while it is very specific about other
topics, such as lean burn, spark timings and intake flow passages.
We first discuss the need for alternative fuels in terms of the
current state of fossil fuel reserves, as well as issues regarding
pollution. General performance parameters used in engine testing
and analysis are then introduced in order to guide the discussion in
later sections. How natural gas is implemented in SI and CI engines
is then discussed. Engine performance, combustion characteristics,
and emission levels are assessed and quantified. Natural gas is the
main fuel discussed in this work; other fuels (such as hydrogen
addition, high-cetane pilot fuels etc) serve as secondary fuels to
facilitate or improve natural gas engine operation, and have been
presented as such throughout the manuscript. Multi-fuel operation
can be considered in both SI and CI engines. The term dual-fuel is
used in the CI section exclusively, to indicate when a second fuel,
such as diesel, is injected in the cylinder in order to act as a source
of ignition of the natural gas in CI engines. Finally, the overall
conclusions and recommendations are presented.
that came into force in September 2009 [5] requires both gasoline
and diesel engines to reduce their emissions of nitrogen oxides
(NOx) by about 30%. At the same time, consumers expect improved
engine performance and fuel consumption. This is reflected in the
literature, where there is an increasing need to extract more power
from smaller powerplants (increasing power output per unit
engine mass; otherwise known as specific power) [6e12]. In order
to accomplish these goals, a new generation of energy-conversion
powerplants would need to be produced. This new generation
would have improved thermal efficiency, increased specific power,
and reduced harmful exhaust emissions [6,9,13]. Various options to
increase efficiency and power while reducing emissions of automotive powerplants already exist, such as hybrid systems and highpressure direct fuel injection. While these innovations have worked
very well in approaching these targets, alternative fuels will need to
be implemented in order to surpass them as well as reduce crude
oil dependence on a larger, more significant scale.
Table 1
Typical natural gas composition by volume (from [14]).
Species
Content
Methane
Ethane
Propane
Butane
Pentane
Carbon Dioxide
Nitrogen
92%
3%
0.7%
0.02%
0.1%
0.6%
3%
2. Introduction
2.1. Natural gas and biogas as part of the general fuel supply
Fossil fuel consumption is steadily rising as a result of population growth in addition to improvements in the standard of
living. It can be seen from Fig. 1 that world population has been
increasing steadily over the last 5 decades, and this trend is
expected to continue [2]. As a result, total energy consumption has
grown by about 36% over the last 15 years [3]. Energy consumption
is expected to increase further in the future, as world population is
expected to grow by 2 billion people in the next 30 years [2]. These
energy trends can be seen in Fig. 2. Increased energy demand
requires increased fuel production, draining current fossil fuel
reserve levels at a faster rate. In addition, about 60% of the world’s
current oil reserves are in regions that are in frequent political
turmoil [3]. This has resulted in fluctuating oil prices and supply
disruptions. For example, Fig. 3 shows that oil prices doubled from
June 2007 to May 2008, only to halve in May 2009 [4].
Crude oil is a finite resource dependent on availability and
stability of fossil fuel supplies. As internal combustion (IC) engines
are expected to continue service well into the next century, the
road transport sector in particular needs more secure and
sustainable future fuel sources. Legislated reductions of certain
exhaust gas emissions are another issue facing the current generation of IC engines. For example, the EURO 5 emission regulation
One of the more established alternative fuels is natural gas.
Natural gas is a gaseous fossil fuel, consisting of various gas species.
A detailed typical composition is shown in Table 1 [14]. Fossil
natural gas is found either together with other fossil fuels (such as
crude oil in oil fields, as well as with coal in coal beds) or on its own.
The properties of natural gas are very similar to those of methane,
which is its primary constituent. One of the reasons why natural gas
is the focus of this work is its significantly larger proven reserves
compared with crude oil (the current known reserves-to-production (R/P) ratio for crude oil is about 40 years, while for natural gas
it is about 60 years) [3]. The variation of these R/P ratios over the
past quarter century is also shown in Fig. 4. The R/P levels of crude
oil and natural gas have been relatively steady over the last two
decades. This is because, in addition to new reserves being
discovered, previous supplies of natural gas that were previously
inaccessible can now be obtained as a result of new technology
allowing practical and economical recovery.
It will be seen in the following that SI and CI natural-gas fueled
engines would provide better performance with direct injection of
the gas in the engine cylinder. This requires gas pressures up to
30 MPa. Compressing gas directly from atmospheric pressure and
temperature to 30 MPa in isentropic compressors requires work
input equal to about 3.6% of the energy content (enthalpy of
combustion) of the natural gas, and more in terms of exergy content.
T. Korakianitis et al. / Progress in Energy and Combustion Science 37 (2011) 89e112
Fig. 1. World population 1950e2050 (December 2008 update) (from [2]).
Fig. 2. Fossil fuel consumption from 1983 to 2008 (from [3]) with approximate current
reserves-to-production ratios in remaining years.
Fig. 3. Brent Europe crude oil prices June 2007eMay 2009 (from [4]).
Over such high-pressure ratios industrial compressors would normally be intercooled. Isothermal compression (which is the appropriate ideal case for intercooled compressors) of natural gas from
atmospheric pressure and temperature to 30 MPa and atmospheric
temperature requires work input equal to about 1.7% of the energy
content (enthalpy of combustion) of the natural gas, and more in
terms of exergy content. On the other hand natural gas is frequently
stored in vessels at 20 MPa. On-board compression of natural gas
from 20 MPa and atmospheric temperature to 30 MPa requires work
input equal to about 0.1% of the enthalpy of combustion of the
natural gas, and more in terms of exergy content. These work
91
(exergy) expenditures for natural gas compression should be taken
into account in the well-to-wheel assessment of the fuel.
There already exists a fairly extensive and developed general
supply infrastructure for natural gas as a result of its use in electrical power generation. For example in the UK, use of natural gas
has been steadily increasing in market share, replacing coal as the
main fuel source for electricity generation [15]. In 2008, natural gas
provided 43% of total electrical energy, while coal provided about
34%. This gradual substitution is being implemented mainly to
reduce carbon dioxide (CO2) emissions. From stoichiometry,
compete combustion of methane produces less CO2 by mass
compared to complete combustion of coal (assuming both
combustion processes taking place in air). This is because about
one-third of natural gas (by molar mass) is hydrogen, while coal is
primarily pure carbon. Natural gas also has a significantly higher
combustion enthalpy per unit mass (also known as the lower
heating value) than coal (45 MJ/kg compared with 34 MJ/kg) [16].
Where automotive powerplants are concerned, natural gas has
already been used with reasonable success [17]. This is most
apparent in South America, where in Brazil and Argentina there are
nearly 3.4 million natural gas vehicles (NGVs) in operation. This is
especially large compared with just 221 NGVs running in England
[17]. South American countries encourage alternative fuel use
(Brazil also uses bio-ethanol as an engine fuel extensively) to
reduce their dependence on crude oil and establish energy security; and as a result the crude oil demand of Brazil and Argentina
have been relatively constant for the past decade [3]. South America
also has the largest number of natural gas refueling stations, with
both Brazil and Argentina having nearly 2000 stations respectively,
compared with 31 in England [17].
Natural gas has a higher combustion enthalpy per unit mass
than gasoline or diesel. This can be seen in Fig. 5 which compares
the energy densities per unit mass and per unit volume. The left
side of the figure indicates the total mass, which includes fuel mass
and storage container mass, of various fuels in various storage
mediums. The grey areas shown in the figure indicate the penalty of
the weight of the container. The actual values indicated by the
smaller columns in the bar chart are lower that the bars themselves
indicate (in some cases the smaller bars have been exaggerated to
make them visible in the chart). As a result of the gaseous state of
natural gas, conventional liquid fueling systems will not work. This
concerns the on-board fuel storage and fuel-injection/carburetion
systems in particular. High-pressure compressed natural gas tanks
are required to replace the liquid-fuel tank of a conventional
vehicle, in addition to high-pressure fuel lines. Modified fuel
injectors or fuel induction systems are required, as a higher mass
and volume flow rate are needed to overcome the low density of
natural gas. Despite these modifications, NGVs cannot operate over
the same distances as conventional vehicles. This is reflected in
Fig. 5, which shows the energy densities per unit volume of various
fuels in various storage media. It is clear that an NGV carrying
a typical pressurized natural gas tank does not carry the same
amount of fuel energy (in MJ) as the same-volume tank of gasoline
or diesel. Hydrogen has similar limitations, where liquid hydrogen
must be stored at 260 C (about 1 bar) to 245 C (about 20e100
bar), imposing serious insulation and boil-off prevention problems.
While general supply and handling of natural gas is widespread
(powerplant electricity generation, building heating), there is a lack
of a dedicated natural gas refueling infrastructure for NGVs. Unlike
in South America, the natural gas refueling infrastructure elsewhere in the world does not exist on the same scale as the gasoline
infrastructure. Therefore it is difficult to regard a gaseous fuel as
a complete replacement of gasoline and diesel-fueled vehicles for
the near term. However, NGV use in cities and suburban areas
remains attractive. This is because travel distances are relatively
92
T. Korakianitis et al. / Progress in Energy and Combustion Science 37 (2011) 89e112
Fig. 4. Reserves-to-production ratios for crude oil (left side) and natural gas (right side) from 1984 to 2008 (from [3]).
120.00 (STP)
DME
Hydrogen
Methane
Energy Carriers
10
0
DME
Hydrogen
Methanol, 15.80
RME, 34.00
0.75 (gas,100 bar)
1.5 (gas,200 bar)
4.10 (liquified)
11.65 (hydrides)
0.26 (nanotubes)
0.04 (STP)
0.25 (gas,10 bar)
2.49 (gas,100 bar)
4.98 (gas,200 bar)
21.00 (liquified)
20
18.30 (liquified)
Ethanol, 21.20
Gasoline, 32.60
30
Diesel, 36.60
40
0.01 (STP)
0.08 (gas,10 bar)
0
50
0.05 (STP)
0.41 (gas,10 bar)
4.05 (gas,100 bar)
8.10 (gas,200 bar)
20
Enthalpy of combustion
(lower heating values) / MJ/litre
40
60
0.80 (gas,100 bar)
1.63 (gas,200 bar)
17.40 (liquified)
7.33 (hydrides)
13.08 (nanotubes)
50.00 (STP)
0.27 (gas,10 bar)
2.60 (gas,100 bar)
4.94 (gas,200 bar)
24.60 (liquified)
Methanol, 20.00
RME, 38.60
60
0.08 (gas,10 bar)
80
Ethanol, 26.90
Gasoline, 44.00
100
Diesel, 42.50
28.40 (STP)
0.44 (gas,10 bar)
3.87 (gas,100 bar)
6.80 (gas,200 bar)
16.10 (liquified)
Enthalpy of combustion
(lower heating values) / MJ/kg
120
Methane
Energy Carriers
Fig. 5. Fuel energy densities: left, enthalpy of combustion per unit mass; and right, enthalpy of combustion per unit volume (the smaller bars have been exaggerated to make them
visible in the two bar graphs).
short and a large refueling network is not necessary. NGVs produce
lower levels of carbon monoxide (CO) and non-methane unburned
hydrocarbons (HC), where the former is toxic to humans and the
latter is a known carcinogen [16,18]. An additional feature of natural
gas engines is that their main HC emission is methane. In the USA
HC emissions regulations are specified in terms of reactivity in the
photochemical smog cycle, and methane has negligible reactivity.
On the other hand, the global warming effect of methane on
greenhouse gases is 30 times the effect of CO2 over 100 years [19].
Photochemical smog is also a problem in densely populated areas,
where NOx reacts with volatile organic compounds (VOCs) in
sunlight to produce particulates and ground-level ozone. As natural
gas-powered engines produce low smoke and particulate levels,
their contribution to smog formation is minimal compared with
gasoline and diesel-powered engines.
Thus NGVs become desirable in densely populated cities, where
local buses and trucks are retrofitted to run on natural gas
[18,20,21]. Public transport systems have central refueling stations,
which make the refueling infrastructure for natural gas easier to
implement. In addition, natural gas is cheaper to buy compared
with conventional fuels. In 2008 natural gas cost about 1.2 US
dollars per 100 MJ while crude oil cost about 1.6 US dollars per
100 MJ [3]. Therefore, while an initial investment would be
required to convert a conventional vehicle to an NGV, the lower
price of natural gas compared with other fuels will return the
investment expense over time [22].
As natural gas is a fossil fuel like gasoline and diesel, it is not
renewable. However, methane (which is the main constituent of
natural gas) can be produced in renewable manner [23,24]. The
entire process of collecting, purifying and using methane gas
emissions from landfill and biomass decomposition is fairly
straightforward, especially when compared with the FischerTropsch process used in gas-to-liquid and biomass-to-liquid
processes. Commonly known as biogas or landfill gas, this gas is
a by-product of anaerobic biological decomposition. It is usually
composed of at least 50% methane and up to 50% CO2, with trace
amounts of hydrogen, nitrogen and hydrogen sulfide [14]. When
directly used in engines [23,24] lower power output is obtained
(compared with operation on pure natural gas). This is a result of
the significant amount of contaminant gas species contained in
biogas, which slows the flame speed of methane. Increased
frequency of basic engine maintenance is also required. Tests with
simulated biogas (pure natural gas with varying volume substitutions of CO2) show reduced emissions of NOx as well as reduced
power and thermal efficiency [14,25,26]. In order to achieve similar
performance as pure natural gas engines, biogas has to be purified
prior to use, which may increase production complexity and cost.
Well-to-wheels (WTW) life cycle analyses have been performed
T. Korakianitis et al. / Progress in Energy and Combustion Science 37 (2011) 89e112
[21,27,28], concluding that fossil natural gas consumes similar
amounts of energy per unit mass to gasoline and diesel (of the
order of 2 MJ/km) over its life cycle. Fossil natural gas produces
similar amounts of CO2 throughout its life cycle as gasoline and
diesel (of the order of 150 g/km). Methane produced from renewable sources such as biomass and landfill gas has significantly lower
WTW CO2 than natural gas (about 250 g/km). This means a net
reduction of atmospheric CO2 can occur. However, as a result of the
added processing and purification required, the energy requirements rise to about 3.5 MJ/km [27], where these numbers correlate
to the energy needs of a state-of-the-art European mid-size vehicle.
Before discussing how natural gas performs as a fuel in different
types of engines, it is necessary to illustrate the different parameters that can influence general engine performance as well as
exhaust emissions. This will later guide the discussion of trends and
results presented in this review.
Firstly, most hydrocarbon fuels (of various compositions of
carbon and hydrogen, CaHb) completely burn in air according to
a standard stoichiometric equation to produce CO2, water (H2O)
and nitrogen (N2). Modeling air as a 21% oxygen (O2) and 79% N2
mixture by volume, the stoichiometric combustion of any hydrocarbon fuel is represented by
b
b
ðO2 þ 3:773N2 Þ/aCO2 þ H2 O
Ca H b þ a þ
4
2
b
N2
þ 3:773 a þ
4
(1)
And using molecular masses of 12.011, 2.016, 31.998, 28.157 g/mole
for C, H2, O2, N2 respectively (where the molecular weight of
nitrogen is adjusted from its value of 28.012 g/mole to include the
effects of other gases in the composition of air [16]), the fueleair
mass ratio of the stoichiometric oxidation reaction is computed as
F
12:011 þ 1:008y
¼
A st
34:559ð4 þ yÞ
(2)
Here y ¼ b/a. Comparing the actual fueleair ratio (F/A)ac to the
stoichiometric fueleair ratio (F/A)st gives rise to the definitions of
equivalence ratio f and its inverse l:
fh
ðF=AÞac
1
¼
l
ðF=AÞst
(3)
In the real engine combustion process the chemical constituents
dissociate at high combustion temperatures and pressures, so that
combustion is not stoichiometric. Hundreds of species and reaction
equations affect the process, and as a result new species are formed
in the exhaust products of combustion such as nitrogen oxide (NO)
and nitrogen dioxide (NO2) (commonly referred to as NOx), CO and
CO2 as well as unburned HC and excess O2 in both excess-fuel and
excess-air combustion processes. The emissions characteristics of
engine-fuel combinations, measured in terms of specific emissions
(mass flow rate of emission divided by engine power produced, (kg/
s)/(W) or g/MJ), or parts per million (ppm) of emission in the engine
exhaust, are an integral part of overall engine-performance studies.
Engine compression ratio (rc) is defined as the maximum
volume (Vmx) over the minimum volume (Vmn) permitted by the
piston assembly and cylinder geometry, defined by
rc h
Vmx
Vmn
Maximum possible theoretical (as well as non-ideal) engine
thermal efficiency and engine power are functions of engine
compression ratio, and each is maximized at different compression
ratios [6,8]. Thermal efficiency can also be increased by having the
expansion ratio larger than the compression ratio [31,29,30].
Steady-state engine performance is governed by a number of
equations derived from fundamentals. The thermal efficiency (hth)
_ Þ divided by the
is defined as the (brake) shaft power output ðW
b
_
rate of energy expended ðEin Þ, where the rate of energy expended is
_ f Þ times the lower
equal to the mass flow rate of fuel consumed ðm
heating value of the fuel (LHVf) (the energy released per mass of
fuel oxidized)
_
W
b
2.2. Fuel and engine performance parameters
(4)
93
hth h _
Ein
¼
_
W
b
_ f LHVf
m
(5)
where the LHVf is used instead of the higher heating value (HHVf) as
the water formed from oxidation of the hydrogen in a typical
hydrocarbon fuel is usually in vapor form at the operating conditions of engine exhaust.
Maximum power is usually produced when the maximum
amount of fuel is used, at stoichiometric conditions, so that at
maximum power
_f
m
F
z
_a
m
A st
(6)
_ a and m
_ f are the mass flow rates of air and fuel through the
where m
engine.
The volumetric efficiency of the engine hv is defined as the ratio
_ a divided by the
of actual mass flow rate of air through the engine m
maximum possible mass flow rate of air that the engine geometry
would allow
hv h
_a
m
ra;in Vd kNs
(7)
where ra,in is the density of air at engine inlet, Vd is the engine
displacement, and k is the number of power strokes per engine
revolution, so that k ¼ 1 for engines operating on two-stroke cycles
and k ¼ 1/2 for engines operating on four-stroke cycles. Combining
the above equations we obtain expressions for the maximum shaft
power at near-stoichiometric design-point operating conditions,
and for the actual power at off-design part-load conditions.
F
_
_
h
W
¼
m
LHVf
a th
b;mx
A st
(8)
F
_
LHVf
W
b;mx ¼ khth hv ra;in Vd Ns
A st
(9)
F
_
LHVf
W
b;ac ¼ khth hv ra;in Vd Ns
A ac
(10)
F
_
h
h
r
f
¼
k
V
N
LHVf
W
b;ac
th v a;in d s
A st
(11)
These simple but powerful equations justify the impetus for
most avenues of engine research. Power output is directly
_ a , and to equivalence
proportional to the mass flow rate of air, m
ratio, f. Inlet and outlet manifold polishing and variable manifold
geometry are used to optimize hv, and for the same reason we
prefer to inject fuel inside the engine cylinders rather than in the
intake manifold (the volume of vaporized fuel at the intake
reduces the volume flow rate of air, and thus reduces the
94
T. Korakianitis et al. / Progress in Energy and Combustion Science 37 (2011) 89e112
_ a through the engine). Maximizing ra,in leads to turmaximum m
bocharging, supercharging and intercooling. Power output is
proportional to Vd and Ns. Higher values of Vd for automotive
engines are taxed in Europe and Japan, and Ns is limited by mean
piston speed (lubrication, vibration and stress considerations), gasdynamic choking of the airflow in the intake ports, and valve-train
dynamics. Alternative fuels affect the influence of LHVf and (F/A)st
in the equations, and later sections show issues with the value of f
that can be used with natural gas. NASA has recently funded
research in developing engines with geometric variations permitting k ¼ 2 for use in unmanned aerial vehicles [8,9]. Earlier engines,
e.g. of the early 1900s, were less efficient than modern engines of
the same power range. To quantify this, there has been an
improvement of about 24% in the average fuel economy of
passenger cars over the last two decades [32], and some of this
improvement is due to smaller improvements in automotiveengine thermal efficiencies over the same time. The largest available slow speed marine diesels have increased their efficiency from
the region of 45% in the 1960s to approaching 55% in the 2000s.
This increase has been achieved with technological advances in
turbocharging, intercooling, super-long strokes decreasing the
surface to volume ratio and resultant heating losses, etc. Similarly
contemporary lawn mower engines are less efficient than modern
automotive diesels which are less efficient than modern slow
speed marine diesels due to decreasing surface to volume ratios
with increasing engine size, and resultant heating losses. Thus
engine thermal efficiency is a function of engine technology and
power output (physical size).
Other important engine parameters are the brake mean effective pressure BMEP and the indicated mean effective pressure IMEP,
_ , frictional power
derived from considerations of brake power W
b
_ , and mechanical efficiency hme, defined
_ and indicated power W
W
i
fr
by
[16]. A higher octane number indicates a higher resistance to autoignition, which in turn allows higher compression ratios to be used.
Higher compression ratios result in higher thermal efficiency.
Methane has about: 2.2% higher values of LHVf and 17.2% lower
values of (F/A)st than gasoline; and 5.9% higher values of LHVf and
19.0% lower values of (F/A)st than diesel. Furthermore, natural gas
has much lower density than gasoline or diesel, so that when it is
injected and expanded in the intake manifold it reduces engine
volumetric efficiency.
Fig. 6. Typical gasoline-engine fuel map, adapted from [33].
_ ¼ W
_ þW
_
W
i
b
fr
(12)
_
F
W
b
BMEPh
¼ hth hv ra;in f
LHVf
A st
kVd Ns
(13)
_
W
i
IMEPh
kVd Ns
(14)
_
W
BMEP
hme h _ b ¼
IMEP
Wi
(15)
where both BMEP and IMEP are measured in units of pressure
(MPa). The IMEP has the physical meaning of the average pressure
in the cylinder, given its name by the historical recordings of
pressure-indicator (pressureevolume) diagrams recording the
pressure measured inside the cylinder plotted as a function of
volume, and taking the average value of the pressure over the
_ , obtained
operating cycle. The corresponding indicated power W
i
by integrating the pressureevolume trace, is higher than the brake
_ .
_ by the frictional power W
(shaft) output power W
b
fr
Any alternative engine fuel should have a beneficial effect on these
equations (increased LHVf and increased (F/A)st, especially when
compared with gasoline or diesel). As an automotive fuel, natural gas
has qualities which make it very suitable for use in reciprocating
piston engines [1]. A property comparison between methane (the
main constituent of natural gas, see Table 1), gasoline and diesel fuel is
shown in Table 2. Methane (and natural gas) have a higher octane
number compared with gasoline. Here, octane number is a measure
a fuel’s resistance to autoignition (commonly known as knocking)
Most CI and SI engines exhibit a power versus speed characteristic similar to that shown in Fig. 6. The region of maximum thermal
efficiency hmx is attained just below 100% of engine rated power and
speed. The value of hmx is a function of engine technology and size
(power rating). Small lawn mower SI engines have maximum
thermal efficiencies about 5%, large lawn mower SI engines attain
efficiencies of about 8%, small automotive SI engines attain efficiencies about 25%, small truck CI engines attain efficiencies about
40%, and large slow speed marine diesel engines can attain
maximum efficiencies in the region of 50e55%. Around the
maximum thermal efficiency regime lie contours of lower thermal
efficiencies as illustrated in Fig. 6. The figure also illustrates the idling
and maximum engine speed, the stalling torque (max BMEP) and the
turbocharger matching and bearing load limit lines. Typical 1st, 3rd
and 5th gear for automotive engines are also shown, where the
gearing would need to be carefully designed to ensure that the load
line in the top forward gear passes through the maximum efficiency
region at engine r/min and power corresponding to highway speeds.
In other applications, for instance in marine propulsion, the
propeller must be carefully matched with the engine at full load so
that the full load line, corresponding to line 5, passes through the
maximum efficiency region at ship design speed and power, while
the light load condition (empty ship) corresponding to load line 1
should be above the manufacturer’s suggested minimum load line
(heavy fuel operation in slow speed marine diesel engines below the
minimum load line causes excessive engine fouling). The fuel map of
a particular engine model will be slightly different with different
fuels, and therefore the design of the gearing for a particular application, e.g. automotive, must be optimized for the combination of
T. Korakianitis et al. / Progress in Energy and Combustion Science 37 (2011) 89e112
Table 2
Chemical properties and physical characteristics of common fuels at standard
atmospheric conditions.
Fuels
Diesel
Gasoline
Methane
Hydrogen
Chemical formula
Density/kg/m3
Lower heating value
(LHVf)/MJ/kg
Octane number
Cetane number
Stoichiometric fueleair
ratio ((F/A)st)
Autoignition temperature/ C
CnH1.8n
827e840
42.5
CnH1.87n
750
44.0
CH4
0.725
45.0
H2
0.09
120.0
e
52
0.069
95
e
0.068
120
e
0.058
120
e
0.029
250
280
650
585
particular engine, particular vehicle, vehicle operating profile, and
the fuel. Gearing studies of natural gas vehicles are reported in [34],
where the appropriate gearing ratios for a modified SI engine (for
use with natural gas) were found.
Fig. 7. Typical NO, HC and CO trends with equivalence ratio in a SI engine, adapted
from [16].
All these performance parameters have a direct relationship with
the exhaust emissions produced, often with contradictory effects.
For instance, while higher compression ratios are favored in order to
increase thermal efficiency, they also result in higher NOx emissions
because of the resultant higher combustion chamber temperatures.
This is also the case when running stoichiometric fueleair mixtures,
as seen in Fig. 7 (which is applicable to gasoline engines, but the
general trends are similar for natural gas engines as well). In addition, while combustion of lean fueleair mixtures (f < 1) result in low
NOx emissions (as seen from Fig. 7) this can also result in lower
power output (as seen by the effect of f in equation (11)). However,
running an engine on fuel-rich mixtures (f > 1) is also undesirable
(the implied benefits from equation (11) are not realized because
there is not sufficient oxygen to oxidize the fuel), and this results in
high unburnt HC and CO emissions. Knock limits are also a factor
when deciding ideal operating parameters. For instance if an engine
is running too high a compression ratio, resistance to knock is lowered. This would require the need for spark retardation with respect
95
to combustion TDC (which can affect thermal efficiency and therefore power output as well as exhaust emissions). Another alternative
to reduce knock is to introduce diluents such as recirculated exhaust
gas (EGR) in the intake charge. However, this reduces the amount of
air through the engine (c.f. equation (8)) and reduces volumetric
efficiency (c.f. equations (7) and (11)). Thus EGR can produce higher
HC and CO emissions while reducing NOx. These tradeoff effects
need to be kept in mind when using alternative fuels such as natural
gas, especially on standard engines modified to run on natural gas.
2.2.1. Analysis of cylinder-pressure data
In addition to the parameters shown in the previous subsection,
another important parameter is the diagram of energy change of the
working fluid in the chamber (frequently referred to as the “heatrelease” diagram). When plotted against crank angle, these plots are
especially useful in engine-performance analyses. This is especially
true for CI-engine analyses, where the ignition point depends on
other factors such as fueleair charge composition, instantaneous
pressure and temperature (while in SI engines the ignition point is
effectively the spark timing). In CI engines these diagrams pinpoint
the ignition point, in addition to quantifying combustion progress
and duration. Most natural gas engines are SI engines, as natural gas
does not readily ignite spontaneously under compression. However,
methane has a high-octane number (Table 2) and therefore
consideration of natural gas use in CI engines is warranted as CI
engines employ an inherently high compression ratio. It is therefore
logical to investigate how natural gas will work in these engines as
well [35]. When high-octane fuels are used in CI engines (provided
ignition is supplied by a small amount of high-cetane pilot fuel), this
mode of CI engine operation is called “dual-fueling”, a term not to be
confused with bi-fueling in which vehicles can run on two different
fuels but not simultaneously. Dual-fueling further complicates CIengine combustion, as one fuel is used to ignite another within a very
short time. This ignition phenomenon can be observed by using
diagrams of rate of energy change of the working fluid. The term
“rate of energy change of the working fluid” (dEn/dt, derived in the
following equations) is preferred here instead of the frequently used
term “heat-release rate”, or the term “thermal energy change”. The
term dEn/dt derived in the equations below includes two terms:
dQto/dt as rate of heat transfer between the walls and the charge; and
dEch/dt as rate of chemical energy released by the oxidation of the
fuel. The sum of the two terms results in the rate of energy change of
the charge in the chamber. This sum corresponds to the net rate
energy change of the working fluid in the chamber, and it is neither
a rate of heat exchange, nor a rate of heat release.
A simplified single-zone model based on the principle of energy
conservation for a closed system is used to calculate the rate of
energy change of the working fluid data from cylinder-pressure
data [36]. The entire system is assumed to have average properties
and no distinction is made between burnt zones, unburned zones
or fuel zones. The charge mixture is treated as a single fluid with
uniform pressure and temperature, while charge flow into crevices
such as in-between the piston and cylinder walls is neglected.
Enthalpy of reaction is also treated by means of a separate energy
supply to the system. With these assumptions, the following
equation is obtained:
Qto þ Wto ¼ DU
(16)
where Qto is the heat transfer to the system (usually negative as
heat leaves the combustion chambers through the walls and into
the thermodynamic systems surrounding the charge), Wto is the
work done to the system and DU is the increase in internal energy
of the system. This internal energy increase can be further divided,
in mass specific terms, into fuel enthalpy (subscript f), internal
96
T. Korakianitis et al. / Progress in Energy and Combustion Science 37 (2011) 89e112
energy of reactants (subscript r), and internal energy of products
(subscript p), as shown below:
DU ¼ mp up mr ur mf hf
(17)
A mass balance between the products and reactants gives
mp ¼ mr þ mf and is substituted in equation (16) to give:
DU ¼ mr þ mf up mr ur mf hf
(18)
which is rearranged and rewritten (from the definition of enthalpy
h h u þ Pv) to give:
DU ¼ mr up ur þ mf up hf
h
i
¼ mr up ur þ mf up uf Pf vf
(19)
The term (up uf) can be expanded to give:
up uf
up u0p þ u0p u0f þ u0f uf
¼
(20)
where the terms with the superscript 0 are at standard-state
reference conditions.
As a result, (u0p u0f ) is the specific internal energy of combustion of the fuel at standard-state conditions. Previous work [36] has
concluded that the values of (up u0p), (u0f uf) and Pfvf are small
(only 1.15% of the specific internal energy of the fuel) and therefore
negligible. Equation (19) then becomes:
DU ¼ mr up ur þ mf u0p u0f
evaluated as a function of crank angle, enabling evaluation of the
“rate-of-reaction” or “rate of energy change of the working fluid”
term dEn/dt in equation (28), which is then plotted against crank
angle. An example of a conventional and dual-fuel CI-engine
diagram of rate of energy change of the working fluid can be seen in
Fig. 8 (from [37]). The start of combustion, otherwise known as the
point of ignition, is defined as the crank angle at which the rate of
energy change of the working fluid in the plot suddenly rises above
the zero datum. Ignition delay is defined as the time period
between the start of fuel injection and start of combustion, or onset
of ignition [16]. Fig. 9 also shows a typical rate of energy change of
the working fluid for multiple cycles of an SI engine (from [16]). In
SI engines usually there is no second peak as most of the fuel is
burned fairly quickly. This is a result of the premixed nature of the
fueleair mixture (provided fuel is injected into the intake air prior
to entering the combustion chamber). Fig. 9 indicates that cycle-tocycle variations are significant in SI engines. These variations are
caused by different ratios of fueleair mixture being inducted for
each cycle. This parameter is especially important, as it can
significantly affect engine stability during operation. Natural gas
can significantly affect cycle-to-cycle variations when used in SI
engines, as detailed in Section 2.
(21)
This is substituted in equation (16) and rearranged to give:
mf u0p u0f ¼ Qto Wto þ mr up ur ¼ Ech
(22)
where Ech is the gross chemical energy resulting from the combustion process and Qto is the total heat transfer from the cylinder walls
to the working fluid. Equation (22) can then be rewritten for small
incremental changes, which gives it in the same form as in [16]:
dEch þ dQto ¼ P dV þ mr du
(23)
where the dWto term has been replaced by the definition e P dV.
Also by definition, du h cvdT so that
dEch þ dQto ¼ P dV þ mr cv dT
(24)
where cv is the instantaneous isochoric specific heat capacity of the
system and dT is the instantaneous temperature change. dT can be
written using the ideal gas equation P ¼ rRT or PV ¼ mRT to give:
dEch þ dQto ¼ P dV þ mr cv
V dP þ P dV
mr R
Fig. 8. Comparison of net rate of energy change of the working fluid dEn/dt plots for
dual-fuelling operation and normal engine operation (from [37]).
Natural gas mixed with air generates different flame propagation speeds than gasolineeair mixtures. Comparing stoichiometric
combustion of methane (representing natural gas) and isooctane
(25)
Finally, equation (25) can be rearranged to give:
En hEch þ Qto
dEn ¼
1þ
c cv v
PðdVÞ þ
VðdPÞ
R
R
(26)
(27)
where dEn is the net energy change of the working fluid in the
combustion chamber, which is the sum of dEch and dQto. Dividing
through by time increment dt equation (27) becomes:
dEn
E_ n h
¼
dt
cv dV
cv
dP
þ
P
V
1þ
dt
dt
R
R
(28)
Pressure transducers are used to record the pressure inside the
cylinder as a function of crank angle, and cylinder volume is also
Fig. 9. Gross rate of chemical energy change dEch/dq diagrams taken over ten cycles for
a SI engine (from [16]).
T. Korakianitis et al. / Progress in Energy and Combustion Science 37 (2011) 89e112
(representing gasoline) in air at atmospheric conditions, methane
has a laminar flame speed of about 38 cm/s while isooctane has
a flame speed of about 32 cm/s [16]. However, these speeds change
according to the charge conditions (i.e. temperature and pressure).
When typical cylinder charge conditions just before ignition are
taken into account (typical temperatures about 700 K and pressures about 3 MPa), the laminar speeds of methane and isooctane
(again at stoichiometric conditions) are about 70 cm/s and 90 cm/s
respectively. This results in slower flame propagation with
methane or natural gas across the combustion chamber, slowing
the rate at which energy is converted from fuel chemical energy
into mechanical work on the piston. This characteristic must be
taken into account during the design stages of dedicated and
converted natural-gas engines.
3. Natural gas in spark-ignition engines
The UK Department of Transport [38] reports that about 66% of
new passenger cars sold in 2008, and the majority of personal
vehicles in the USA, are fueled by gasoline, i.e. using SI engines. As
a result, any alternative fuel should possess qualities that allow its
use in these current engines. In other words, consumers should still
be able to use their vehicles if and when a new alternative is phased
in to replace gasoline. Natural gas is usually inducted or injected in
the intake manifold, resulting in lower volumetric efficiencies, and
reduced power (c.f. equation (11)). On the other hand the higher
octane rating of natural gas allows for increased compression ratios.
In combination with turbocharging or supercharging and intercooling this can result in 50% increase in power output and thermal
efficiency. Additional induction of small amounts of hydrogen in
the fuel mixture increases flame speed and combustion stability.
Properly tuned natural gas SI engines result in reduced NOx, CO, CO2
and non-methane HC but increased emissions of CH4 compared to
gasoline. The following sections review the experimental data
justifying the above statements. The modifications required for
engine optimization are also presented.
3.1. Performance
Previous work [1,14,39] reports that engines fueled by natural
gas can produce thermal efficiency of the order of 5% higher than
gasoline-fueled engines. This can be due to several factors. The first
is the high-octane number of natural gas, which allows comparatively higher compression ratios than gasoline engines to be used.
Previous work [14] found that compression ratio can be increased
from 8:1 to 13:1 with an engine running on natural gas, while
compression ratios higher than that caused knocking. Increasing
compression ratio in SI engines can significantly increase thermal
efficiency. For example, doubling the compression ratio can
increase thermal efficiency by about 13% [16]. The 2.2% higher
combustion enthalpy (LHVf) of natural gas compared to gasoline
erroneously indicates natural gas engines may produce more
power than gasoline engines (equation (11)). This expectation is not
fully met, and several factors act in favor or against increased efficiency and power for reasons explained below, so that careful
changes must be planned in natural gas engines in order to derive
their full potential. The stoichiometric fueleair ratio of methane is
about 17.2% lower than that of gasoline, and (with respect to fuel
effects) it is the product of f (F/A)st LHVf that affects power in
equation (11). Natural gas SI engines generally operate with higher
inlet-manifold pressures (roughly 10% higher) than gasoline
engines, resulting in lower pumping work losses [40]. These lower
losses are caused by wider throttle settings which reduce the
restriction in the intake manifold, with a beneficial effect to volumetric efficiency. Despite the above beneficial effects, natural gas
97
engines produce lower power levels across the operating range
compared to gasoline-fueled engines (at the same compression
ratio and comparatively advanced spark timing when fueled with
natural gas) [41]. This is primarily because of the lower density of
natural gas, which requires the induction of a large volume of
natural gas in order to produce similar power levels as gasoline
engines. This volume of gaseous fuel displaces a similar volume of
air from entering the engine, therefore reducing volumetric efficiency hv (by up to 10e15% depending on engine configuration [1]),
with a corresponding reduction in power as shown in equation (11).
This limits the amount of air that can be induced and fuel that can
be burnt, significantly reducing power output as indicated by
equations (8)e(11). Airflow into the engine can also be further
restricted by flow mixers and venturis installed for natural gas
induction in retrofitted natural gas engines [41]. It will be shown
later in the CI engines section that the location of natural gas
injection in the intake manifold has a significant effect on the
resultant volumetric efficiency.
Fig. 10 shows that combination of these beneficial and adverse
factors reduce power output of a gasoline engine converted to run
Fig. 10. Power comparison for a typical bi-fuel natural gas (vehicular natural gas,
VNG)/gasoline vehicle (from [41]).
on natural gas by about 15%. These tests were taken on a rolling road,
with the vehicles accelerating from first to top gear (Fig. 10 shows
data taken when in top gear only). A secondary factor is the
advanced spark timing these natural gas engines employ. The spark
timing (advanced relative to combustion TDC) is required due to the
slower flame speed of natural gas. Previous work [42] showed that,
depending on engine speed, the spark advance can be up to 10
crank angle earlier compared to gasoline operation. As ignition is
occurring earlier in the cycle, increases in charge pressure from
natural gas combustion is working against the piston during the
compression stroke, increasing the work transfer from the piston to
the charge and reducing the network the engine is able to produce
during the full cycle. A 10 advance or retard beyond maximum
brake torque spark timing (for a gasoline engine at wide open
throttle) can reduce torque output by about 2% [16]. Steps should be
taken to modify gasoline engines converted to run on natural gas in
order to compensate for these sources of power loss, and locate the
high efficiency islands in beneficial areas related to the gearing,
where the effect of the location of gearing lines is indicated in Fig. 6.
This effect should be accounted for in the original engine-gearing
design of vehicles planned to operate on natural gas.
The power loss in naturally aspirated gasoline engines converted to run on natural gas can be compensated by increasing the
compression ratio [14], or by forced induction (supercharging or
turbocharging) [43,44]. Natural gas operates under high compression ratios without knocking as a result of its high-octane number.
Forced induction increases the ra,in term in equation 11 from that of
naturally aspirated engines, directly increasing power output. In
98
T. Korakianitis et al. / Progress in Energy and Combustion Science 37 (2011) 89e112
[43] a 1.0 l SI engine was operated on both gasoline and natural gas,
with and without supercharging. The results showed that natural
gas (with natural aspiration) produced 10e15% lower torque output
than gasoline. On the other side, increasing the compression ratio
combined with supercharging (at 0.6 bar boost pressure) increased
torque by about 50% from the baseline naturally aspirated gasolinefueled configuration.
Fig. 12. Laminar flame speed variation of a stoichiometric natural gaseair mixture
with hydrogen addition by volume (from [45]).
Fig. 11. Coefficient of variation in IMEP versus equivalence ratio (from [1]).
Natural gas engines can run leaner fueleair mixtures than
gasoline engines. Cycle-to-cycle variations remain of the same order
as in gasoline engines (4% coefficient of variation in indicated mean
effective pressure, COV of IMEP, equation (14)) [1]. This can be seen in
Fig. 11. The figure also shows that the lean operating limit of natural
gas (inducted into the engine’s intake manifold using a mixer) is
extended by about 28% compared to gasoline. However, significantly
lower cycle-to-cycle variations of IMEP (about 1.5% COV of IMEP)
were produced when natural gas was injected into the intake
manifold [1]. Further extension of the lean operating limit was also
obtained. This is because of more precise control and distribution of
the fueleair ratio, which helps maintain a relatively constant
fueleair mixture from cycle-to-cycle. It is desirable to decrease this
fuel-lean operation limit further, in the hope of obtaining decreased
fuel consumption and lower emissions of regulated pollutants.
However, under these very fuel-lean conditions, cycle-to-cycle
variations increase significantly [1]. The slow burning velocity of
natural gas, combined with excessive air (which make some local
charge regions too fuel-lean) produce excessively low combustion
temperatures. This results in low thermal efficiencies. Misfire can
also occur (as a result of some cycles failing to ignite at all), and
significant engine damage is a possibility in extreme cases.
One way to overcome these high cycle-to-cycle variations at
very fuel-lean operating conditions is to add another fuel with
a higher flame speed to the existing natural gaseair mixture.
Adding hydrogen gas to the natural gaseair mixture in the inlet
tract has promising results [27,40,45e49]. The higher flame speed
of hydrogen speeds up combustion, resulting in more fuel being
burnt in the time available. The laminar flame speed of stoichiometric natural gasehydrogeneair mixture increases exponentially
with hydrogen concentration [45], as shown in Fig. 12. Previous
work [48,50] found that hydrogen addition has a greater effect on
the beginning stages of combustion than in later stages of
combustion. This is because hydrogen burns faster than natural gas
in the early stages of combustion, and also because the flame
development stage, which occurs just after ignition, is less turbulent than the middle and latter stages of flame propagation. In the
cases near the limits (i.e. in-between the laminar and turbulent
phases), the laminar flame speed of hydrogen is about 7 times
higher than the laminar flame speed of natural gas, and is of
comparable value to its own turbulent flame speed. However, the
turbulent flame speed of natural gas is about 10 times higher than
its laminar flame speed [48], so that the turbulent flame speed of
natural gas is higher than that of hydrogen. As a secondary but
contributing effect, hydrogen gas addition introduces hydrogen and
hydroxyl (OH) radicals, both of which increase combustion reactivity during the flame development period [50].
As a result of the shortened combustion duration, spark timing can
also be set closer to TDC, starting combustion later in the cycle and
reducing the work done on the charge during compression. As
hydrogen has a wide flammability range (0.1 f 7.1 compared with
0.7 f 4 for gasoline [51]), a relatively small amount of hydrogen
can be used. Higher peak combustion chamber pressures are reached
with increasing hydrogen gas addition (up to 1 MPa higher) [45].
Efficiency is increased, going up to 21% for a 30% hydrogen substitution by volume, compared with 12% for operation with pure natural
gas [45]. Fig. 13 indicates that increasing the percentage of hydrogen
gas addition by volume in natural gas significantly reduces the COV of
IMEP, as a result of the more volatile fuel mixture.
Hydrogen addition also increases tolerance of exhaust gas
recirculation (EGR) in natural-gas fueled engines [27,47,48].
Previous work done with natural gas engines running with EGR has
shown that EGR has a detrimental effect on combustion progress at
all volume substitutions [27]. As pure natural gas combustion is
relatively slow to begin with, EGR lowers the concentration of
oxygen in the inducted charge, slowing flame propagation even
further. This produces comparatively lower combustion temperatures. The higher specific heat capacity of the total inducted charge
also contributes to this effect [16,52]. Fig. 14 indicates that peak
thermal efficiency occurs at a spark timing of 29 before TDC when
fueled with pure natural gas, while a similar peak thermal efficiency with 10% EGR occurs at a spark timing of 44 before TDC for
T. Korakianitis et al. / Progress in Energy and Combustion Science 37 (2011) 89e112
99
of hydrogen was introduced for the same EGR level (10%) and spark
timing (44 BTDC), thermal efficiency increased about 0.5% [47]. For
an EGR substitution of 20% used with hydrogen and natural gas,
there was a 20 retard (relative to combustion TDC) in the optimal
spark timing compared with the same condition without hydrogen
addition [48]. At EGR substitutions less than 20%, hydrogen addition had a very positive effect on the mentioned parameters.
Increasing hydrogen addition in tandem with increasing EGR rates
help maintain thermal efficiency levels and peak combustion
pressures [27]. Without hydrogen, increasing EGR rates results in
significantly lower combustion quality. EGR substitutions of more
than 20% produce significantly low peak combustion pressure
values (close to motored pressure levels), in addition to significantly low rates of combustion regardless of the amount of
hydrogen substitution for the conditions of the tests in [47].
3.2. Exhaust emissions
Fig. 13. COV of IMEP with increasing hydrogen addition by volume in natural gas (from
[49]).
the conditions of the tests in [47]. This highly advanced timing
results in combustion occurring significantly earlier, increasing the
work done on the charge by the piston during compression, in
addition to increased heat transfer to the cylinder walls. Hydrogen
addition alters this behavior [27]. When a 10% volume substitution
Natural gas engines generally produce lower emissions of CO and
non-methane hydrocarbons compared to normal gasoline engines
[22,41]. CO2 emissions are slightly reduced [20,22,41] due to basic
stoichiometry: 1 g methane produces 2.8 g CO2 while 1 g of gasoline
produces 3 g of CO2. The lower equivalence ratio in natural gas
engines results in CO emission reductions between 50% and 90%
compared to gasoline engines (though CO and HC emissions
increase significantly at extremely low equivalence ratios as a result
of deteriorating combustion quality [16,22,41]). Unburned HC
emissions are reduced by up to 55% at the same time [41]. Both these
trends correlate with Fig. 7. Other work reports exactly the opposite,
where total unburned HC emissions increase [22] (by about 160%).
However, 90% of these HC emissions were unburned methane, and
non-methane HC emissions were lowered (by about 70% [22]). As
mentioned in Section 2, the contribution of methane to smog
formation is negligible compared to non-methane unburned HC
emissions; but methane is a powerful greenhouse gas (it possesses
a global warming potential 30 times more than CO2 over 100 years)
[21]. The increase of methane emissions is typical of converted
gasoline engines to run on natural gas when the engines are not
optimally modified [41]. These cases are often end-user conversions,
where the engine tuning (such as spark timing, injection timing etc)
for the use of natural gas is not optimized in the conversion. In these
engines a significant amount of natural gas can escape combustion
through flame quenching, adsorption in crevice volumes, and
adsorption in the lubrication oil film on the cylinder walls.
Fig. 14. Fuel conversion efficiency comparison with spark timing for various natural gasehydrogen mixtures and EGR rates (from [47]).
100
T. Korakianitis et al. / Progress in Energy and Combustion Science 37 (2011) 89e112
NGVs with aftermarket natural gas engine conversions (conversions not explicitly endorsed by the vehicle manufacturer) produce
up to 170% higher levels of NOx emissions compared with typical
gasoline levels [22]. The higher NOx levels in comparison to
conventional SI engines operating on gasoline result from higher
combustion chamber pressures and temperatures, because of the
advanced spark timing and higher compression ratios. This causes
NOx formation rates to accelerate (compared to normal gasoline
operation) during combustion as the compression stroke progresses,
because of increasing combustion temperatures and pressures. The
higher oxygen concentration in fuel-lean mixtures encourage higher
NOx formation rates and emissions. These trends correlate to natural
gas engines running higher compression ratios compared to gasoline
engines. When compression ratios between the two engines are the
same, it is likely that lean fueleair mixtures in natural gas engines
and in gasoline engines will produce lower NOx emissions as a result
of lower combustion temperatures compared with stoichiometric
operation, as seen from Fig. 7. The higher oxygen concentration at
fuel-lean conditions only offsets the lower temperatures up to
a certain level (about f z 0.9); lower temperatures become the
greater factor and reduce NOx emissions at leaner mixtures [16].
Exhaust gas catalysts can be used to reduce these high NOx
emissions, sometimes by more than 90% [18]. Three way catalysts
(catalysts that reduce CO, HC and NOx emissions) are common in
gasoline engines, however they operate best with a stoichiometric
mixture ratio. As natural gas engines can run with relatively fuellean mixtures compared with gasoline engines, only selective
reduction catalysts can be used under these conditions. As these
catalysts are complex by design, expensive, and difficult to maintain [50], it is preferable to reduce the emissions in the combustion
process itself. One method to achieve this is to introduce a diluent
in the natural gaseair charge. This diluent can be a variety of gases,
EGR for example. This reduces the oxygen concentration in the
charge in addition increasing its specific heat capacity of the charge,
and reducing flame speeds. These factors reduce peak combustion
temperatures, thus suppressing NOx formation [16,47,52]. An
example of the level of reduction in NOx emissions is shown in
Fig. 15, which shows the effect of volumetric substitution of
hydrogen (an effect to be discussed later) and mass substitution of
EGR in the intake charge on NOx emissions.
Other gases which are inert can be used as an alternative to EGR
[14,25], such as CO2 and nitrogen (N2). Gas species have different
specific heat capacities (CO2 has a higher specific heat capacity than
N2 for example), and therefore will alter the specific heat capacity of
the charge mixture. However, any inert diluent introduced into the
intake charge can have detrimental effects on combustion progress,
especially with the inherently slow combustion of natural gas. This is
shown by the increased emissions of unburned HC, as seen in Fig. 16.
Previous work has shown that adding gases with higher flame
velocities, such as hydrogen gas, to the natural gaseair mixture can
speed up combustion [27,40,45e49]. Hydrogen addition to natural
gas engines without EGR reduces emissions of unburned HC and
CO, while at the same time increasing NOx emissions. These trends
can be seen in Figs. 15 and 16, where results for increasing hydrogen
fraction are plotted with increased EGR rates. These emission
trends result from shortened combustion duration and increased
combustion temperature. It has also been suggested that simultaneous addition of hydrogen and excess air can lead to a decrease in
all recorded emissions [46]. However, this only occurs at certain
operating conditions [46]. Previous work [27] indicates that, for an
engine running with a 10% hydrogene90% natural gas fuel mixture
by volume, NOx emissions decrease with increasing EGR mass
substitutions. Unburned HC emissions decreased with increasing
EGR substitution up to about 12%, after which levels increase. This
can be attributed to more complete combustion as a result of the
Fig. 15. Reduction of NOx in a natural gas engine with varying hydrogen and EGR levels
for a fixed spark timing (from [27]).
Fig. 16. Increase of unburnt HC in a natural gas engine with varying hydrogen and EGR
levels (from [27]).
accelerated combustion, which overcomes the effects of flame
speed reduction with lower EGR rates. EGR substitutions of more
than 15% have a greater effect on combustion (in terms of flame
speed) than a fixed mass hydrogen substitution of natural gas. In
addition, for any EGR substitution level, increasing substitutions of
hydrogen in the intake charge produce lower unburned HC levels
than the pure natural gas case [27].
3.3. Natural gas direct injection in SI engines
The majority of NGVs use port fuel-injected natural gas engines.
However, as natural gas displaces some of the air in the intake
T. Korakianitis et al. / Progress in Energy and Combustion Science 37 (2011) 89e112
manifold upon injection, a reduction in volumetric efficiency
occurs, which in turn leads to proportional reduction in power, as
indicated by equations (8)e(11). One way to avoid this is to inject
fuel directly into the combustion chamber [53]. Port fuel injectors
have operating pressures of the order of 0.5 MPa, which are lower
than the pressures of 8 MPa (or higher) used to overcome operating
cylinder pressures for direct gas injection. Direct natural-gas
injection in SI (and CI) engines requires development of specialty
high-pressure gaseous injectors [1,39,53] which currently are not
available in the open market.
In addition, direct natural gas injection has been shown to extend
the fuel-lean operating limit of normal engine operation compared to
port fuel injection [1,39,53]. For example, a port-injected natural gas
engine running very fuel-lean mixtures (f z 0.6) had a COV of IMEP
of more than 10%, while the same engine using an in-cylinder injector
built into the spark plug was operating at the same equivalence ratio
with a COV of IMEP of less than 5% [1]. This is a result of increased
mixture turbulence in the cylinder, in addition to locally fuel-rich
mixtures that become available close to the spark plug. Flame propagation is also accelerated, resulting in higher rates of energy change
of the working fluid and higher thermal efficiencies. However, higher
levels of NOx were recorded, caused by high combustion chamber
temperatures [1]. Contrary to this, reference [53] reports that NOx is
lower for direct injection engines because of increased charge stratification. It is likely that other engine parameters play a significant
role, as reference [1] goes on to say that retarding the spark timing
(relative to combustion TDC) can reduce NOx emissions while
affecting power output. Similar performance effects are also seen
with increasing compression ratio in direct injection engines [54],
where a compression ratio of 12:1 is reported as the optimum
compromise value between performance and emissions for direct
injection natural gas SI engines.
When fuel-injection timing is varied care must be taken not to
start injecting the fuel into the cylinder too late. The experiments
reported in reference [39] indicate early injection timing (during
the intake stroke) increases the combustion chamber pressure and
rate of energy change of the working fluid, resulting in favorable
power output. These trends are shown in Fig. 17, where 150 BTDC
to 180 BTDC (TDC here relates to combustion TDC) are the start-ofinjection timings during the compression stroke and 180 BTDC to
210 BTDC are the start-of-injection timings during the intake
stroke. The optimum injection timing in this work is shown to be at
180 before combustion TDC. Retarded injection timing during the
compression stroke does not allow sufficient time for the fuel to
mix and oxidize, resulting in poor flame propagation as well as
101
Fig. 18. Rates-of-reaction variation with injection timing advance with regards to
combustion TDC (from [39]).
Fig. 19. NOx variation with beginning of injection timing in crank angle degrees BTDC
(from [39]).
reduced and delayed peak rate of energy change of the working
fluid. This can be seen in Fig. 18. Exhaust emissions follow a similar
trend to that of engine power output [39]. For example, in Fig. 19
NOx emissions increase significantly with injection timing earlier
in the compression stroke, while there is little effect with further
injection timing advance into the intake stroke. HC emissions
follow the opposite trend to NOx emissions while CO emissions did
not change significantly with fuel-injection timing [39].
Hydrogen addition in direct-injected natural-gas fueled SI
engines has similar trends to those mentioned earlier in Section 3.2.
For a fixed spark timing, thermal efficiency increases with increasing
hydrogen fraction in the natural gaseair fuel mixture, while
combustion duration decreases. For a fixed hydrogen fraction, HC
concentration decreases and exhaust NOx concentration increases
with advancing spark timing (relative to combustion TDC), while CO
emissions do not vary significantly with spark timing [55].
4. Natural gas in compression-ignition engines
Fig. 17. Power output (Pe) variation with beginning of injection timing in crank angle
degrees BTDC (from [39]).
While SI engines are the dominant automotive powerplants
[38], there is still a significant percentage of passenger vehicles in
Europe that use diesel fuel, i.e. CI engine-powered (about 34% of the
fleet). In addition, almost 100% of goods vehicles as well as buses
use CI engines in Europe; and diesel-electric trains are extensively
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T. Korakianitis et al. / Progress in Energy and Combustion Science 37 (2011) 89e112
used in the USA. Natural gas as an alternative fuel should be usable
in CI and SI engines. CI engines have inherently higher compression
ratios than SI engines, and correspondingly higher thermal efficiencies. Natural gas has about 5.9% higher LHVf and 19.0% lower (F/
A)st than diesel (c.f. Table 2) with corresponding effects on power
indicated by equations (8)e(11). In most applications to date
natural gas is inducted or injected in the intake manifold, resulting
in lower volumetric efficiencies and the expected reduction in
power indicated by equation (11). In some applications natural gas
is directly injected in the cylinder.
Ignition quality is related to octane number (a measure of
“knocking” characteristics, used in SI engine-fuel combinations)
and cetane number (a measure of ignition delay characteristics,
used in CI engine-fuel combinations) [16,56]. For a given fuel there
is an inverse but not proportional relationship between octane and
cetane numbers. Methane, the main constituent of natural gas, has
a higher octane number than gasoline and a lower cetane number
than diesel; the cetane number of methane is so low that it is not
usually quoted. Ignition quality of methane is significantly different
from that of typical gasoline and diesel fuels, as indicated by octane
and cetane numbers in Table 2. Thus in both port-injected and
direct-injected natural-gas CI engines, natural gas does not spontaneously ignite under typical CI compression ratios (and the corresponding temperatures) like diesel, but needs a source of
controlled ignition. Thus multiple sources of ignition are provided
by diesel, biodiesel or other high-cetane fuel that is injected into
the cylinder via the traditional fuel injector in CI engines, and which
acts as a pilot fuel initiating natural gas combustion (in what is
defined as dual-fueling in CI engines).
Typically about 30% of the total fuel energy is supplied by the
pilot fuel. If the inducted natural gas is admitted via a small pipe in
the intake manifold near the intake valve, then experimental data
suggests that the resultant volumetric efficiency is only reduced by
2e5%. In higher engine speeds this results in more cycles per
second and less natural gas per cycle, and therefore lowers power
output because of insufficient fuel supply per cycle. This effect is
less pronounced at lower engine speeds. Like in SI engines,
hydrogen can be added to the fuel mix to increase flame speed and
combustion progress. Compared to diesel fuel emissions in
conventional CI-engine operation, natural gas use with pilot fuel
injection in CI engines results in reduced NOx at low loads and
about the same levels at high loads. However, CO and HC at low
loads are increased due to incomplete combustion resulting from
lower charge temperatures. CO and HC levels at high loads reduce
to normal CI-engine levels. CO2 levels are lower throughout the
load range because of the higher hydrogen-to-carbon ratio of
natural gas. The following sections review the experimental data
justifying the above statements. The modifications required for
engine optimization are also presented.
4.1. Performance
Natural gas does not autoignite under compression alone with
typical CI-engine compression ratios. For use in CI engines, a special
mode of operation known as dual-fueling is required. This ignition
source is provided by a spontaneously igniting “pilot” fuel. A small
amount (“pilot” injection) of a high-cetane fuel is injected directly
into the combustion chamber, where the spray mixes with a premixed natural gaseair charge. After an initial ignition delay period,
the pilot fuel ignites and begins to burn. Depending on the level of
mixing between the natural gas and air, a second delay period
occurs after which the natural gas ignites, producing most of the
energy released during combustion. Dual-fueling was conceived in
an effort to reduce NOx and smoke emissions that are common in
conventional CI engines. Dual-fuel engines are fairly effective in
doing so, while at the same time maintaining acceptable thermal
efficiency levels over the operating range [35,57e65]. While the
reduced flame speed during natural gas operation is a problem in SI
engines, it does not prove to be as significant an obstacle in CI
engines. This is because of a number of factors. With typical sparkignition engines only one source of ignition can be produced (the
spark plug). When the spark plug ignites, the flame propagates
from this single point outward through the charge. In conventional
CI-engine operation the fuel penetrates into the cylinder during
injection and mixes into the charge. When charge conditions allow
(pressure, temperature, fueleair mixing levels etc) these dispersed
droplets of fuel spontaneously ignite. This produces multiple ignition points throughout the chamber, resulting in multiple flame
fronts and a comparatively faster burn rate. In dual-fuel engines
(where the pilot fuel is injected directly into the cylinder), this large
spread of points compensates for the slow flame speed of natural
gas and allows faster combustion.
CI engines lack intake throttle valves along the inlet tract and
therefore their volumetric efficiency is higher than that of throttled SI engines. Load control in CI engines is affected by changing
the amount of fuel admitted to the cylinder. This reduces the
pumping losses, which are a common efficiency loss in SI engines.
Previous work [37,66] found that dual-fueled CI engines only have
slightly reduced volumetric efficiencies compared with normal CI
engines, of the order of 1e4%. This minimized effect is because of
the apparatus design, where the natural gas enters the intake
manifold via a steel tube placed close to the intake valve. This
tube, which is permanently installed in the engine, takes up
a fixed space in the intake manifold at all times, and therefore
affects the flow of air inducted into the engine even during normal
engine operation. This lowers the engine’s volumetric efficiency
during normal-fueling and marginally so during dual-fueling (as
a result of further air displacement by the natural gas). This
apparatus also limits the maximum flow rate of natural gas in
these dual-fuel studies, affecting power output at high engine
speeds as discussed below.
Power output during dual-fuel operation is matched with
normal CI-engine operation at lower speeds (1000 r/min) but,
with this particular induction method, during dual-fuel operation
this engine produces significantly less power at high speeds
(>1000 r/min) compared with normal CI-engine operation. This is
because more intake cycles occur per second at higher speeds. For
a certain natural gas flow rate through the intake tube, this results
in relative fuel starvation with regards to the natural gas supply
compared to lower speeds. Other work [59,62] found that this
effect does not occur when a mixing chamber was used to mix the
natural gas and air mixture prior to entering the intake manifold.
With a mixing chamber, the engine will simply induct more
fueleair mixture at higher speeds. Thus the method of natural gas
induction into the engine has significant effects on engine operation at different conditions.
Cycle-to-cycle variations of dual-fuel engines are comparable to
normal CI engines (both produce COV of IMEP levels of about 1%)
[67]. Any slight differences can result from a number of factors,
most of them commonly encountered in SI as well as CI engines. For
example, there are variations in the amount of natural gas present
in the intake charge per cycle, varying degrees of mixing between
fresh charge and residual cylinder gases, as well as mixture motion
variations in the cylinder per cycle [16,67].
In [37,68] the pilot fuel ignition delay was extended slightly
during dual-fuel operation, by about 0.08 ms, which could result
from the slightly higher specific heat capacity of the natural gaseair
mixture. Dual-fueled engines are noisier than normal CI engines
throughout the normal operating range [37,68,67], and knocking is
encountered at very high loads [69]. This knocking is caused by the
T. Korakianitis et al. / Progress in Energy and Combustion Science 37 (2011) 89e112
comparatively short duration of natural gaseair combustion (after
pilot fuel ignition), which is shorter than normal diesel combustion
[37,68]. This can be seen in Fig. 8 where the dual-fuel rate of energy
change of the working fluid reduces to zero at about 5 ATDC while
the normal rate of energy change of the working fluid reduces to
zero after 30 ATDC. For normal operation, the diesel fuel that has
mixed with air to about stoichiometric levels during the ignition
delay period burns first (shown by the peak occurring at about
7.5 ATDC). This is followed by a longer, lower peak (occurring at
about 2.5 ATDC) which results from combustion of fueleair
mixture that is just becoming available, otherwise known as mixing-controlled combustion. For dual-fuel operation, the same figure
shows a similar first peak being produced, which is a result of pilot
fuel mixing with air during the ignition delay period. Following this
there is a second increase in the rate of energy change of the
working fluid, which signifies the start of natural gas combustion.
This second combustion phase occurs on a much faster rate as the
natural gaseair mixture enters the combustion chamber already
premixed to a degree similar to conventional port-injected SI
engines. As a result, the crank angle at which this second increase in
the rate of energy of the working fluid occurs can be considered as
important as spark timing in SI engines, and therefore needs to
occur at the right crank angle.
However, the shorter combustion duration seen in Fig. 8 is not
maintained for all conditions in other literature [59]. This difference can be because of the different natural gas mass fractions
employed to hold a particular load in a particular engine. The
natural gas mass fraction can be as much as 86% in reference [59] or
no more than 70% in reference [37,68]. This difference in natural
gas levels is due to differences in the design of natural gas induction in the two sets of experiments (intake tube versus mixing
Fig. 20. Pressure and rate of energy change of the working fluid (heat release rate,
HRR) comparison of dual-fuelling and normal fuelling operation (from [59]).
103
Fig. 21. Comparison of computed rate of energy of the working fluid plots (“heatrelease rates”) for different EGR components during dual-fuel operation (from [70]).
chamber, as explained before). Fig. 20 shows this discrepancy,
where the dual-fuel rate of energy change of the working fluid
peaks are lower compared with normal fueling, while the reverse is
true in Fig. 8. The “Z” term in Fig. 20 indicates the level of natural
gas enthalpy substitution.
Some experimental studies [37,68] conclude that thermal efficiencies in dual-fuel operation are similar to those with normal CIengine operation (at max BMEP levels), while others [59] indicate
significantly lower efficiencies. The conclusion is that stable and
acceptable dual-fuel engine operation depends on the particular
engine and natural gas induction system used, as the original
engine design parameters affect dual-fuel combustion significantly.
While at high loads dual-fuel engine performance is similar to
normal engine operation, at low and intermediate loads lower
thermal efficiencies compared with normal CI-engine operation are
recorded. At these part-power conditions the pilot fuel fails to
ignite most of the main natural gaseair mixture. This is is the result
of a comparatively lower charge temperature (and pressure)
because of the lower fueleair ratio [58,70,72,73]. Uncooled EGR
improves low to intermediate load combustion, in addition to
reducing knocking tendencies in dual-fuel engines. Fig. 21 shows
the increase in computed peak rate of energy change of the working
fluid (in addition to shorter combustion duration) with a 2% mass
EGR substitution compared with conventional dual-fuel operation
[70]. The figure also shows the thermal (Th), chemical (Ch) and
radical (Ra) effects of EGR on the rate of energy change of the
working fluid. The thermal effect represents the effect of charge
temperature, the chemical effect represents the effect of charge
composition, and the radical effects represents the effect of active
radicals (partially oxidized products) present in EGR on the
combustion reactions. Compared with the radical effect (increased
rate of pre-ignition reactions) and the chemical effect (improved
mixture strength), it is the thermal effect of EGR (increasing charge
temperature) that is the main factor. This is because the dilution of
the charge by the EGR overcomes the potential improvement
brought on by the chemical and radical components of EGR.
Other work [72] concludes that relatively low EGR levels (of the
order of 5% by mass substitution) result in increased rates of
pressure rise (which reflects higher rate of energy change of the
working fluid) at low loads. This is because uncooled EGR increases
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T. Korakianitis et al. / Progress in Energy and Combustion Science 37 (2011) 89e112
Fig. 22. Peak rate of energy change of the working fluid (heat release rate, HRR) and COV variations with gaseous fuel-injection timing for dual-fuel operation with 40% EGR
(from [71]).
the temperature of the intake charge and accelerates start of ignition, though it also lowers volumetric efficiency (c.f. equation (11)).
At high loads, EGR alleviates knocking as it dilutes the intake
mixture, reducing the oxygen concentration and slowing combustion. This reverse trend is because the effect of EGR diluting the
charge is significantly greater than its heating at high loads, while
the reverse is true at low loads [70,72].
If higher levels of EGR are required in order to reduce NOx, then
this dilution of the charge with exhaust gas would also reduce
combustion progress. In these cases hydrogen gas addition
increases combustion stability and reduces ignition delay [71].
With 23% hydrogen substitution by volume of natural gas as well as
an EGR volume substitution of 40% of the intake oxygen reduced
COV of IMEP by about 1%. Fig. 22 shows peak rate of energy change
of the working fluid (in this case plotted against gaseous fuelinjection timing) increase with increasing hydrogen addition. In
Fig. 22, 50% “IHR” is used to define the combustion timing. By
integrating the rate of energy change of the working fluid curve up
to a crank angle where it has reached its maximum value, and
normalizing by the total combustion energy over the full cycle, the
fraction of the energy released in the cycle is determined. The
midpoint of this curve (50% of the integrated heat release or rate of
energy change of the working fluid curve, defined as the 50% IHR),
is plotted against crank angle and is used to define the combustion
timing in crank angle [71].
The same figure also shows the lower COV of IMEP with
advancing (with respect to combustion TDC) gaseous H2 fuelinjection timing. In this case the high flame speed of hydrogen
maintains combustion stability despite the high EGR rate. Ignition delay for the gaseous fuel combustion is reduced significantly (by about 20%) and the rate of energy change of the
working fluid was also increased [71]. In addition, no effect on
the engine-fuel consumption as a result of hydrogen addition was
noted [71].
4.2. Exhaust emissions
Dual-fueling in CI engines reduces NOx emissions significantly
compared with normal diesel engine operation [35,58,59,
57,60e65]. At low and intermediate loads the reduction is about
50% depending on engine type and conditions.
Fig. 23 shows the typical trend of NOx emissions from a dual-fuel
engine with diesel pilot fuel at two different speeds with varying
load levels (in terms of BMEP). In this case dual-fuel operation
lowers NOx emissions compared with normal engine operation. The
NOx reduction is primarily because of lower combustion temperatures, resulting from the slower flame speed of natural gas [59].
However, other work [35] reports that the NOx formed during the
combustion of the pilot fuel constitutes most of the total NOx
Fig. 23. Comparison of NOx emissions for dual-fuel operation as function of BMEP
(from [59]).
emissions of dual-fuel combustion. Therefore, it is the type and
quantity of pilot fuel that significantly affects final exhaust NOx
emissions during dual-fuel operation. Like conventional diesel
combustion, there is a critical time period between the start of pilot
fuel combustion and the time when peak combustion chamber
pressure is reached. In conventional CI engines, this critical time
period is usually the first 20 crank angle after the start of pilot fuel
combustion [16]. This is because peak combustion chamber
temperatures occur during that period. Here, the premixed pilot
fueleair mixture is compressed to a high temperature during the
pilot ignition delay period just prior to combustion, which accelerates the NOx formation rate during combustion. Combustion of
the natural gaseair mixture usually begins when the cycle is
already late in the expansion stroke (where the charge is being
cooled as the pressure drops) and NOx formation rates from this
second phase of combustion are lower. This charge cooling, as
a result of expansion in addition to mixing between hot and cold
gases, freezes NOx chemistry, preventing decomposition. Attempts
T. Korakianitis et al. / Progress in Energy and Combustion Science 37 (2011) 89e112
105
have been made to reduce NOx emissions further, by significantly
advancing pilot fuel-injection timing [74]. Here, the pilot fuel was
allowed time to mix with the natural gaseair mixture in order to
avoid locally rich mixtures, which increase NOx formation.
Smoke, soot and particulate emissions in dual-fuel engines are
very low and in some cases undetectable [58,59,57,60e65]. Natural
gas combustion produces little smoke, as natural gas has a lower
number of carbon atoms than diesel. Methane has no carbonecarbon bonds in addition to a high hydrogen-to-carbon ratio,
which result in lower sooting tendencies [75]. In addition, the
natural gaseair mixture is usually very well mixed just prior to
combustion. This results from the high residence time of the two
gases in traveling and mixing from the intake manifold to the
combustion chamber. Therefore any particulates recorded are
produced during the pyrolysis of the pilot fuel, in a way similar to
particulates formation in conventional diesel engines (i.e. in the
very rich mixture of the pilot fuel spray core) [16]. Some of the soot
so formed will be burned during combustion of the natural gaseair
mixture, further lowering particulate and smoke levels.
Part-load unburned HC emissions of dual-fuel engines are
significantly higher than in normal diesel engines [57e60,62,63,65]
at low to intermediate loads. Fig. 24 indicates HC concentrations of
the order of 6000 ppm, compared with significantly less than
100 ppm in conventional diesel operation [59]. This is caused by
unburned natural gas, i.e. mostly methane, surviving to the
exhaust. This is could be due to the stratified nature of the pilot fuel
that acts as an ignition source. It is suggested that the flame
(initiated by the pilot fuel) does not propagate through the charge
[37,68]. This results from local areas of gaseous fuel and air mixture
being too fuel-lean for combustion [72]. There is a equivalence ratio
threshold (f ¼ 0.4) for dual-fuel engines. Below f ¼ 0.4 unburned
HC emissions increase with increasing overall equivalence ratio
from f ¼ 0.2 to f ¼ 0.4 [37,61,68,76]. Above f ¼ 0.4 unburned HC
levels are lower and approach those of conventional diesel operation. It is possible that some of the natural gas is entering the
crevice volumes of the combustion chamber, cooling and escaping
combustion. Flame quenching as combustion of the natural gaseair
mixture proceeds into the expansion stroke may be a contributing
factor [77]. The above influences are amplified from the usual
expected results by the lower flame speed of natural gas. However,
at high engine loads unburned HC emissions in dual-fuel CI engines
are comparable to those in conventional diesel-fuel operation
[37,61,68], as the fuel-richer mixtures (though still sub-stoichiometric) result in sufficiently higher combustion temperatures to
oxidize most of the fuel.
Fig. 25 shows the variation of CO emissions with BMEP [59]. CO
emissions are significantly higher than normal CI-engine operation
at the speeds tested throughout the load range. Some of the literature confirm these trends to some degree, with even larger
increases with dual-fuel operation at part load [37,68]. In the
experimental data in references [37,68] CO emission trends start at
a comparatively higher value than normal operation at low loads,
and progressively approach normal diesel engine levels with
increasing equivalence ratio, as a result of higher combustion
temperatures at high loads. This indicates that combustion of the
natural gas is not complete at low loads.
A number of methods have been designed to improve low-load
emission trends of dual-fuel engines. One is the injection of a larger
proportion of the pilot fuel contributing to the total combustion
energy, which would provide more ignition points, so that
combustion of the natural gas air mixture would be more complete
[74,78]. While this approach is successful to some extent, it defeats
the objective of reducing diesel fuel demand in CI engines.
Another method to reduce low-load emissions is to induct low
amounts (5% by volume) of uncooled EGR. The hotter exhaust gas
helps increase combustion temperatures, resulting in lowered
Fig. 24. Comparison of HC emissions for dual-fuel operation as function of BMEP
(from [59]).
Fig. 25. Comparison of CO emissions for dual-fuel operation as function of BMEP
(from [59]).
106
T. Korakianitis et al. / Progress in Energy and Combustion Science 37 (2011) 89e112
unburned HC and CO emissions [58,70,72,73]. This follows from the
increased equivalence ratio (as a result of air substitution with EGR)
and a hotter charge. There is a further chemical effect of this EGR, in
the addition of active radicals (partially oxidized combustion
products) present in exhaust gas, which help to drive the chemical
reactions during combustion [73] (as explained in previous
sections). The diluting effect of EGR reduces the oxygen concentration in the charge, and reduces NOx emissions (Fig. 26) as well as
HC emissions (Fig. 27).
The increase of charge temperature as a result of EGR addition
offsets the dilution effect of EGR at low loads, while the opposite
effect is seen with high loads [73]. EGR substitution rates of more
than 5% of the intake air during dual-fuel combustion result in
excessively high unburned HC and CO coupled with low NOx
emissions [72]. At high EGR rates, the dilution of the intake charge
overcomes the positive effects of the hot EGR gases. An attempt to
extend the upper limit of EGR rates by adding hydrogen to the
charge has been reported in [71]. Hydrogen addition of 23% by
volume coupled with an EGR rate of 40% (also by volume) increased
NOx emissions only slightly, while unburned HC and CO emissions
were reduced by about 60% and 40% respectively [71]. These trends
are caused by the high flame speed of hydrogen, which helps
shorten the combustion duration despite the high EGR substitution.
Dual-fuel engines produce lower levels of CO2 than conventional
diesel-fuel engines (about 30%) [37,68]. This is caused by natural
gas having a higher hydrogen-to-carbon ratio than diesel. By stoichiometric combustion in air, 1 g of methane produces about 2.8 g
of CO2, while 1 g of diesel produces about 3.2 g of CO2.
4.3. Dual-fuel CI operation with natural gas and alternative
pilot fuels
Special consideration has to be given to the type of fuel used to
provide ignition, as the pilot fuel is of great importance to the
quality of dual-fuel combustion. While conventional diesel fuel has
been shown to be an adequate pilot fuel for natural gas dual-fuel CI
engines, many other alternative and sustainable fuels have also
been tested [37,57,68,79,80e84]. Among them are select varieties
of biodiesel, i.e. the transesterified ester made from vegetable/
organic oils (such as rapeseed oil) and methanol. The biodiesel fuels
are fairly similar to diesel fuel in conventional single-fuel operation
Fig. 26. NOx emission variation with equivalence ratio for different EGR rates during
dual-fuel operation (from [73]).
Fig. 27. HC emission variation with equivalence ratio for different EGR rates during
dual-fuel operation (from [73]).
in terms of their performance and exhaust emissions [37,57]. This
similarity to diesel fuel is extended to the pilot fuel role. Rapeseed
methyl ester (RME) performs very closely to diesel fuel as a pilot
fuel in terms of thermal efficiency [37,57]. The RME pilot fuel
ignition delay and peak combustion chamber pressure are also very
similar to those of diesel fuel at relatively high load [37,68]. RME
contains oxygen molecules in its chemical composition, which
allows a higher degree of fuel oxidation and compensates for the
lower combustion enthalpy. Similar overall trends in exhaust
emissions [37,57,68,79,80,82] are also recorded, where in these
works the natural gas contributes from 0% to 68% of the fuel energy
with increasing equivalence ratio and BMEP. A slight increase in
NOx and coinciding reduction in HC at high equivalence ratios is
caused by the oxygen present in the in RME pilot oxidizing on
a larger scale, producing higher combustion temperatures.
With other less well known esters, such as honge oil methyl
ester, and jojoba oil methyl ester [81,83,84], differences between
their respective performance with diesel fuel as well as between
the esters themselves are more apparent. Dual-fuel operation with
jojoba-seed methyl ester pilot fuel results in similar efficiencies to
the diesel pilot fuel, while dual-fuel operation with the honge oil
methyl ester result in lower efficiencies throughout the load range.
This is caused by the lower combustion enthalpy of the honge oil
methyl ester [81,83], as indicated by equations (8)e(11). Use of the
jojoba-seed methyl ester pilot fuel [84], with a significantly higher
cetane number of 63, results in lower rates of combustion pressure
rise, reducing noise and knock tendency. These methyl ester pilot
fuels generally produce higher levels of CO, smoke and HC levels
compared to diesel [83]. The differences primarily result from the
higher viscosities of the different methyl esters than diesel fuel (up
to 19.2 mm2/s, compared with about 5 mm2/s for diesel fuel)
[80,81,83,84]. The viscosity of the pilot fuel affects the spray characteristics and the distribution of flame front and combustion
progress across the cylinder.
A compilation of results from [37,68,79] and additional tests run
for this review comparing baseline diesel operation to dual-fueling
of natural gas with various alternative pilot fuels is shown in Figs.
28e32. Pure diesel and RME behave very similarly for performance and emissions in conventional CI operation as well as in
dual-fuel mode. Emissions are usually related to equivalence ratio,
T. Korakianitis et al. / Progress in Energy and Combustion Science 37 (2011) 89e112
Fig. 28. Specific NOx comparison for conventional CI-engine operation with diesel fuel
and natural gas dual-fueling with four pilot fuels at 1000 r/min.
Fig. 29. Specific HC comparison for conventional CI-engine operation with diesel fuel
and natural gas dual-fueling with four pilot fuels at 1000 r/min.
Fig. 30. Specific CO comparison for conventional CI-engine operation with diesel fuel
and natural gas dual-fueling with four pilot fuels at 1000 r/min.
while performance is usually related to BMEP, and the figures from
our own work have been plotted accordingly. Compared to normal
diesel CI operation NOx levels with dual-fueling are lower at low
and intermediate equivalence ratios (f < 0.6), and comparable at
high equivalence ratios (f > 0.7). In dual-fueling NOx levels rise at
lower equivalence ratios (f < 0.4), an effect which can be
compensated with addition of water to the fuel. HC and CO emissions are generally higher with dual-fueling. This effect is more
pronounced at lower equivalence ratios where the lean fuel
107
Fig. 31. Thermal efficiency comparison for conventional CI-engine operation with
diesel fuel and natural gas dual-fueling with four pilot fuels at 1000 r/min.
Fig. 32. Combustion chamber pressure comparison for conventional CI-engine operation with diesel fuel and natural gas dual-fueling with four pilot fuels at 1000 r/min,
0.6 MPa BMEP.
mixture causes lower charge temperatures; but at higher equivalence ratios HC and CO emissions approach those of conventional
CI-engine operation with diesel fuel. Adding water to the pilot fuel
reduces NOx but increases HC and CO at low loads. Thermal efficiency is maintained at high loads with water addition to the pilot
fuel, but it is reduced at lower loads. Overall dual-fueling results in
slightly lower thermal efficiencies than normal CI operation with
pure diesel.
Gaseous and emulsified fuels (liquid fuel mixed with water)
perform well at high loads but not as well at intermediate and low
loads, as indicated in the test in [37,68,77,79,80]. A high-cetane
gaseous pilot fuel allows more homogeneous mixing with the
natural gaseair charge during the ignition delay, which improves
combustion efficiency at low loads. Dimethyl-ether or DME (a highcetane gaseous fuel) used as a pilot fuel produces lower specific NOx
emissions, in addition to higher HC and CO specific emissions during
dual-fuel combustion [77,79,80]. DME vaporizes very quickly upon
injection, cooling the charge and lowering combustion temperatures, resulting in the recorded emission levels. Emulsions of water
and vegetable oil improve thermal efficiency during normal CIengine operation, caused by a phenomenon called “micro-explosions” (the emulsified fuel droplets explode violently during
combustion as a result of water vaporization, allowing more ignition
points to be distributed throughout the charge) [85]. When waterin-fuel emulsions are used as pilot fuels, specific NOx emissions are
lower than those with pure diesel or pure RME pilot fuels, at least
108
T. Korakianitis et al. / Progress in Energy and Combustion Science 37 (2011) 89e112
until an intermediate load (about f ¼ 0.55). Specific NOx levels rise
closer to RME pilot fuel levels [37,68]. A reverse trend is observed in
specific HC and CO emissions. This is because charge temperatures
are too low at the low and intermediate load range to exploit the
micro-explosion phenomenon.
DME as injected pilot fuel results in lower thermal efficiency
than RME at the same BMEP and engine speed, as shown in
Fig. 31. This is caused by the combined effect of 30% higher mass
flow rate and 9% lower LHVf of DME in order to keep the same
pilot energy input to the cylinder, as expected by the combined
effect of equations (5) and (13). Fig. 32 compares pressure traces
as indicators of combustion progress with various fuels. RME and
DME pilot fuels produce higher peak pressures and higher
pressure-rise rates than conventional CI operation with pure
diesel. Water-in-RME emulsions as pilot fuels produce reduced
peak pressures compared to pure RME pilot fuel. Compared to
operation with pure diesel in conventional CI operation, two
counteracting factors affect water-in-RME emulsions as pilot
fuels: microexplosions; and the cooling effect of water vaporization. The 5% water-in-RME emulsion does not produce enough
microexplosions and the water vaporization effect results in
reduced peak pressure compared to that with conventional diesel
CI operation. The 10% water-in-RME emulsion produces enough
microexplosions to overcome the water vaporization effect and
results in increased peak pressure compared to that with
conventional diesel CI operation. The additional microexplosions
with 10% water-in-RME emulsion increase the pressure-rise rate
shifting the peak pressure closer to TDC than the 5% water-inRME emulsion. Comparing conventional CI diesel operation with
natural gas dual-fueling with water-in-RME emulsions this
produces a minimal difference on thermal efficiencies (Fig. 31)
and on NOx emissions (Fig. 28).
Comparing natural gas dual-fueling with water-in-RME emulsion pilot fuels to the pure RME pilot fuels, the water vaporization
lowers charge temperature resulting in reduced specific NOx emissions (Fig. 28). The emulsified pilot fuel ignition delays are increased
compared to the RME pilot fuel, causing the peak combustion
chamber pressure to occur later in the expansion stroke [37,68,79].
The fuel properties affect the microexplosion phenomenon. In
normal CI operation with water in vegetable oil emulsions the
microexplosion effect is much more pronounced than it is in
normal CI operation with water in diesel emulsions. The density
and viscosity of vegetable oil is higher than those of diesel fuel,
which increases the intensity and resultant effects of microexplosions with vegetable oils [85].
4.4. Natural gas direct injection in CI engines
Most of the dual-fuel engine studies to date are conducted in
conventional diesel engines modified to induct natural gas into the
combustion chamber via the intake manifold, while maintaining
the original in-cylinder injector for pilot fuel injection. Like portinjected SI engines, inducting natural gas via the intake manifold
reduces volumetric efficiency, and therefore potential power
output (c.f. equations (8)e(11)). In some studies both the pilot fuel
as well as natural gas are injected directly into the cylinder via the
same injector [71,86]. Dual-fuel engines with direct natural gas
injection maintain power output and thermal efficiency levels
compared with conventional non-dual-fuel diesel engines [71].
Comparatively lower emissions of NOx and particulate matter were
also recorded [71].
Further improvement in direct natural gas injection dualfueled CI engines can be obtained by varying the injection pressure of the natural gas jet and the diesel pilot fuel. For normal
non-dual-fuel CI engines, increasing the fuel-injection pressure
Fig. 33. Specific emissions and fuel consumption trends with injection pressure at
1200 r/min (from [86]).
improves fuel atomization upon injection as well as fueleair
mixing rates prior to combustion [86,87]. A similar effect occurs in
direct natural gas injection dual-fueled engines [86]. Increasing
the injection pressure of both the diesel pilot fuel and the natural
gas injection (from 21 MPa to 30 MPa) results in a shortened
ignition delay of the pilot fuel [86] because of faster mixing
between the pilot fuel and air during the ignition delay period.
Higher combustion-progress rates are recorded, resulting in
a shorter overall combustion duration [86]. NOx emissions are
increased slightly compared with lower injection pressure conditions, in addition to lower HC, CO and significantly lower smoke
emissions. These emission trends are caused by better levels of
mixing between the pilot fuel, natural gas and air, in addition to
the faster combustion-progress rates. Thermal efficiency levels
were not significantly affected by varying injection pressure [86].
These emission trends are shown in Fig. 33 (which are plotted in
specific terms of mass per unit gross indicated kilowatt hour,
GikWhr, where GikWhr is the energy derived from the indicated
rather than the brake power). The data was obtained at a particular intake oxygen mass fraction (YintO2) of 0.19 in the total intake
charge, and combustion timing (50% IHR ¼ 17.5 ATDC). The figure
also shows the gross indicated specific fuel consumption (GISFC),
which is proportional to the inverse of thermal efficiency.
These trends vary significantly with the operating conditions,
especially with engine speed [86]. At low speeds, the higher
injection pressure (30 MPa) have more influence on combustion
quality than at high speeds. This is because the higher injection
pressures increase turbulence in the cylinder at low speeds, while
at high speeds cylinder turbulence is inherently high because of the
piston motion. In addition, at a particular engine speed, increased
turbulence brought on by the higher injection pressures is more
significant at higher loads [86]. This can result from a larger pressure difference between the higher fuel-injection pressure and
chamber pressure, compared with the lower injection pressures
(21 MPa) which are comparatively more effective at low loads.
These parameters influence the level and rate of mixing in the
T. Korakianitis et al. / Progress in Energy and Combustion Science 37 (2011) 89e112
cylinder, which in turn influence emission levels. For example,
particulate emissions are lowered to a larger extent at higher loads
than at lower loads [86].
5. Summary and conclusions
Natural gas is a practical fuel for SI engines, and for CI engines
in the dual-fuel mode, with varying degrees of success. Natural-gas
fueled SI engines can operate at higher compression ratios
resulting in similar or slightly higher thermal efficiencies
compared to gasoline-fueled engines. Natural gas injection or
induction in the intake manifold adversely affects volumetric
efficiency hv. The 2.2% lower LHVf and 17.2% higher (F/A)st of
natural gas compared to gasoline also affects power. Overall the
product of all three factors affects power output (equations (8)e
(11)) resulting in 10e15% reduction in power compared to gasoline-fueled engines. Direct in-cylinder injection of natural gas
avoids the volumetric efficiency effect. Injecting natural gas into
the cylinder requires high pressure (of the order of 30 MPa) and as
a result specialist injectors are required. In addition, ideal power
levels are obtained only at injection timings where high NOx
emissions are produced. The higher hydrogen-to-carbon ratio of
natural gas compared to conventional gasoline higher leads to
relatively minor reductions of CO2 emissions compared to gasoline
engines. These engines use high compression ratios and advanced
spark timing (compared to typical gasoline engines), which
generally increases NOx emissions. Corresponding reductions in
unburned non-methane HC and CO emissions are also reported.
Most of HC emissions are methane, so despite reductions in overall
HC emissions, the methane emissions of natural-gas fueled
engines are higher then those of gasoline engines. EGR can be used
to reduce NOx emissions but it also results in increasing HC and CO
emissions. The lean-burn strategy generally resolves emissions
issues, but unburnt methane emissions remain relatively high.
Ultra-lean operation results in misfire and unstable engine operation. Fuel-lean operation of natural gas engines is desirable in
order to reduce specific NOx emissions. Natural gas engines
running stoichiometric fueleair mixtures produce lower thermal
efficiencies compared to lean-burn natural gas engines because of
the lower specific heat ratio of the charge [1]. Comparable NOx
levels to lean-burn natural gas engines are reported if spark timing
is advanced relative to TDC appropriately (compared to typical
gasoline engine spark timing settings).
In CI engines natural gas is ignited by the use of a pilot fuel
(that can ignite with typical CI compression ratios) in a special
mode of operation known as “dual-fueling”. A small amount of
“pilot” high-cetane fuel is injected directly into the cylinder, which
provides an ignition source for the premixed natural gaseair
mixture. There is no significant loss in power in dual-fuel operation compared to conventional CI-engine operation provided
a sufficient amount of natural gaseair mixture can be admitted in
the chamber. In some of the literature the induction method
employed prevents a premixed natural gaseair mixture to form in
the inlet manifold (by keeping the natural gas supply separate
from the incoming air until very close to the intake valve). This
minimizes the volumetric efficiency penalty of natural gas
induction or injection in the intake manifold, but it also results in
a loss of power at higher speeds because comparatively lower
amounts of natural gas are inducted per cycle. Failure of the pilot
fuel to ignite the entire natural gaseair charge at the low and
intermediate loads (as a result of low charge temperatures) causes
lower thermal efficiencies. Low engine operating temperatures at
these loads (caused by pockets of local fueleair mixtures that are
too lean to support combustion as the flame propagates along the
chamber) also result in lower NOx emissions compared to normal
109
CI-engine operation. Conversely, HC and CO emissions are significantly increased, while at high loads NOx, HC and CO emissions
are comparable to normal CI-engine levels. A variety of alternative
high-cetane fuels can be used as pilot fuels, while water-in-fuel
pilot fuel emulsions and water injection can be used to reduce
emissions at select equivalence ratios.
In order to derive the full benefits of natural-gas fueled SI and CI
engines, extensive performance and emissions optimization of both
engine types is required. It is likely that a combination of different
engine operating modes is needed. SI engines operating at high
loads can employ high EGR rates as well as catalytic converters to
reduce NOx emissions; while advanced spark timing, high
compression ratios and forced induction can be used at all conditions to improve power output. For example a combination of
turbocharging, high compression ratio, catalytic converters and
engine control unit (ECU) reprogramming allows increased power
and efficiency while at the same time reduces NOx and CO2 emissions [43]. In dual-fueled CI engines, natural gas can be used in
smaller proportions compared to the pilot fuel at lower loads to
reduce emissions of HC and CO (i.e. the pilot fuel would provide
more than 50% of the total fuel energy input). High-pressure pilot
fuel injection (of the order of 100 MPa) can provide more ignition
points distributed more extensively throughout the natural gas air
charge. An increased number of smaller-than-standard diesel
injector holes allow better atomization and mixing of the pilot fuel
with the natural gas [88,89]. Uncooled EGR at low to intermediate
loads speeds up combustion progress and improves combustion
efficiency, and therefore reduces unburned HC as well as CO
emissions. In addition to these modifications, an additional fuel
that can improve the burning characteristics of natural gas can be
included in the intake charge. Hydrogen is effective at increasing
the flame speed of combustion in both SI and CI natural gas engines,
as well as reducing COV and increasing engine stability. Small
amounts of inducted hydrogen increase thermal efficiency and
improve EGR tolerance.
Storage of natural gas is an issue with NGVs. Natural gas requires
modified or new types of fuel storage and supply systems due to its
low density (typically stored in compressed gas tanks at z 20 MPa).
These fuel storage options generally carry less fuel energy per mass
or volume than the typical diesel or gasoline fuel tank (Fig. 5).
Exhaust gas catalysts are effective at improving NGV exhaust
emissions, but their maintenance and disposal costs should be
taken into account. NGVs that have secondary fuels on-board also
require secondary fueling systems (dual-fuel CI engines, engines
with hydrogen addition etc). As a result of limited fuel storage,
NGVs face the problem of a limited operating range. Operation over
long distances requires more-frequent refueling. Short distance
travel in inner cities and typical commuter routes helps reduce
photochemical smog as well as carcinogenic hydrocarbon emissions (but the increased methane emissions contribute to greenhouse gas buildup). These areas allow easier installation of a natural
gas refueling infrastructure.
As natural gas is not a renewable fuel, renewable sources of
methane are required in order to ensure its use in the long term. The
technology to produce renewable methane from waste biomass at
reasonable cost and quality has not yet been developed. Biogas
(harvested from landfills, for example) are the only major sources of
renewable methane. This biogas can be used in natural gas engines
fairly readily; however this would result in reduced power and
efficiency caused by contaminants in the biogas, such as CO2 and
sulphur dioxide (SO2). Biogas requires purification prior to use,
increasing production complexity and cost of the fuel in the process.
Overall, natural gas engines can supplement the existing engine
portfolio and assist in the conservation of the limited supply of
crude oil. Incorporation of other engine technologies (e.g.
110
T. Korakianitis et al. / Progress in Energy and Combustion Science 37 (2011) 89e112
turbocharging, higher compression ratios, control of autoignition)
is required to improve operating range and power output. Efforts to
improve the current position of renewable fuels should be
continued (where well-to-wheel life cycle analyses of the entire
production, supply and use is considered) in order to further reduce
dependence on conventional fossil fuels.
not renewable, its long-term use as an automotive fuel remains in
question. At best, NGVs running on fossil natural gas can reduce
the current demand on the crude oil supply, extending its supply
life.
5.1. Suggested future work
Latin
BMEP (or B.M.E.P.) brake mean effective pressure
isochoric specific heat capacity
cv
E; E_
energy, energy rate
chemical energy (Eqs. 23, 26)
Ech
net energy in cyclinder (Eqs. 26, 27)
En
rate of energy change of the working fluid (Eqs. 26, 28)
E_ n
(F/A)
fueleair ratio
IMEP
indicated mean effective pressure
HHV
higher heating value
k
power strokes per shaft revolution
L
piston stroke
LHV
lower heating value
_
m; m
mass, mass flow rate
N
shaft revolutions (Ns or Nm)
P
pressure
Q
heating
r
ratio
R
gas specific constant
t
time
T
temperature
S
mean velocity or mean speed
U, u
internal energy, specific internal energy
V
volume
_
W; W
work, work rate (power)
Details of laminar flame speeds for various fuels including
natural gas are well known. There is a need for experimental
studies of natural gas combustion simulating typical reciprocating
piston engine and gas turbine operating conditions. For instance
the turbulent flame speed, flame propagation characteristics, and
emissions generation characteristics of pure natural gas in these
engine operating conditions are not well known. The effects of
various additions such as EGR or water (to reduce NOx) or hydrogen
(to accelerate combustion progress) are even less well known. A
few computational studies are available in the open literature, but
quality experimental data to provide basic insight or to validate the
numerical work are currently not available. Similarly additional
experimental and numerical work is required to understand and
predict the emission trends, such as the NOx-HC tradeoff, correlated
to different engine configurations, engine parameter choices, and
combustion regimes. It is not well known how combinations of
different performance parameters (such as compression ratio,
equivalence ratio, ignition timing for SI engines or fuel-injection
timing for dual-fueled CI engines) affect the exhaust emissions of
natural gas engines. At the time of writing this paper, worldwide
quality of emissions computations are not reliable enough to be
included in this review in order to explain emissions trends [90].
Combined experimental and computational studies are needed in
order to develop better models to predict emissions for all engine
fuels. Experimental studies similar to [43] are required for CI
engines. Additional research is needed to optimize dual-fuel CIengine operation with natural gas. Studies to vary pilot fuel
amounts at different engine loads (or equivalence ratios) and to
vary injection timing and pressure are required in order to optimize
emissions characteristics.
Both SI and CI natural gas engines require direct in-cylinder fuel
injection in order to eliminate the low volumetric efficiency with
inlet-manifold induction or injection, and to increase power
output. Advancement of direct gaseous-injection technology is
needed. Currently most direct gaseous injectors are custom made
or prototypes. Mass production is hindered by reliability concerns
because of unknown lubrication and wear properties. Lubrication
concerns with gaseous fuels extends to engine components such as
intake valves and valve seats. Further research into different
component materials and alternative or additional lubricating
strategies for natural-gas fueled engines is necessary.
Advancement of gaseous fuel storage technology is required.
Typical compressed natural gas storage systems cannot approach
typical gasoline or diesel fuel energy densities. Liquid gasoline and
diesel fuels will remain dominant until they are unavailable or until
significant improvements or new storage technologies are introduced to approach the liquid-fuel energy densities per mass and
per volume of stored alternative fuel.
Currently there is limited natural gas supply infrastructure for
NGVs. It is costly to implement an NGV fuel-supply network to
compete with the gasoline and diesel network, especially considering the reduced-range issues due to the natural gas storage issues
identified above as applied to NGVs.
Renewable sources of methane that are reliable and cost
effective are currently lacking. Improvements of harvesting and
purifying biogas and landfill gas are needed. As fossil natural gas is
Nomenclature
Greek
D
h
q
l
r
f
difference operator
efficiency
angle (crank angle)
inverse of equivalence ratio
density
equivalence ratio
Subscripts
a
air
b
brake (or shaft)
c
compression
ch
chemical
d
displacement
f
fuel
fr
friction
i
indicated
in
inlet
in
into thermodynamic system
m
per minute
n
net (rate of energy change of the working fluid in the
combustion chamber)
me
mechanical
mn
minimum
mx
maximum
p
products, after reaction
pn
piston
r
reactants, before reaction
s
per second
st
stoichiometric
th
thermal
T. Korakianitis et al. / Progress in Energy and Combustion Science 37 (2011) 89e112
to
v
total
volumetric
Acronyms
ATDC
after top dead center
ATC
after top center
BDC
bottom dead center
BTDC
before top dead center
CA
crank angle
CI
compression-ignition
CNG
compressed natural gas
COV
coefficient of variation
DoE
design of experiments (computations)
DME
di-methyl ether
EGR
exhaust gas recirculation
GISFC
gross indicated specific fuel consumption
HC
hydrocarbon
HRR
heat release rate (rate of energy change of the working
fluid)
IC
internal combustion
IHR
integrated heat release
MBT
maximum brake torque
NMHC non-methane hydrocarbon
NGV
natural gas vehicle
PM
particulate matter
RME
rapeseed methyl ester
R/P
reserves-to-production ratio
rpm or RPM shaft revolutions per minute
SI
spark-ignition
TC
top center
TDC
top dead center
THC
total hydrocarbons
UHC
unburned hydrocarbons
VOC
volatile organic compound
WTW
well to wheels
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