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 ) only discussed SI engines, while this paper includes CI engines in dual-fueling. In addition, in reference  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  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 ). 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 . As a result, total energy consumption has grown by about 36% over the last 15 years . 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 . 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 . 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 . 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 . 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) . 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 ). Fig. 2. Fossil fuel consumption from 1983 to 2008 (from ) with approximate current reserves-to-production ratios in remaining years. Fig. 3. Brent Europe crude oil prices June 2007eMay 2009 (from ). 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 . 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) . Where automotive powerplants are concerned, natural gas has already been used with reasonable success . 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 . 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 . 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 . 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 ). 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 . 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 . 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 . 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 . 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 , 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 ), 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 , 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 . 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 . _ 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 . 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 , 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 . 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 . 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 . 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  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 ). 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 . Fig. 9 also shows a typical rate of energy change of the working ﬂuid for multiple cycles of an SI engine (from ). 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 : 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 ). 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 ). 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 . 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  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  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% . 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 . 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) . 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 ), 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 . 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 ). 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  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% . 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 , 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  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 ). Fig. 11. Coefﬁcient of variation in IMEP versus equivalence ratio (from ). 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)) . 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 . 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 . 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 , 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 , 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 . 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 ), 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) . Efﬁciency is increased, going up to 21% for a 30% hydrogen substitution by volume, compared with 12% for operation with pure natural gas . 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 . 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% . 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 . 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 . 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 . 3.2. Exhaust emissions Fig. 13. COV of IMEP with increasing hydrogen addition by volume in natural gas (from ). the conditions of the tests in . 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 . 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 . Both these trends correlate with Fig. 7. Other work reports exactly the opposite, where total unburned HC emissions increase  (by about 160%). However, 90% of these HC emissions were unburned methane, and non-methane HC emissions were lowered (by about 70% ). 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) . The increase of methane emissions is typical of converted gasoline engines to run on natural gas when the engines are not optimally modiﬁed . 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 ). 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 . 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 . Exhaust gas catalysts can be used to reduce these high NOx emissions, sometimes by more than 90% . 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 , 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 . However, this only occurs at certain operating conditions . Previous work  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 ). Fig. 16. Increase of unburnt HC in a natural gas engine with varying hydrogen and EGR levels (from ). 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 . 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 . 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% . 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 . Contrary to this, reference  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  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 , 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  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 ). Fig. 19. NOx variation with beginning of injection timing in crank angle degrees BTDC (from ). 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 . 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 . 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 . 4. Natural gas in compression-ignition engines Fig. 17. Power output (Pe) variation with beginning of injection timing in crank angle degrees BTDC (from ). While SI engines are the dominant automotive powerplants , 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 102 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%) . 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 . 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 . 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  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 ). 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 ). 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  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 . 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  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 104 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 ). 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 . 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 . 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 . In addition, no effect on the engine-fuel consumption as a result of hydrogen addition was noted . 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 . However, other work  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 ). 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 . 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 . 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 . 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) . 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 . 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 . 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 . 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 . 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 ). Fig. 25. Comparison of CO emissions for dual-fuel operation as function of BMEP (from ). 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  (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 . 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 . 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 . 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 . 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 ). Fig. 27. HC emission variation with equivalence ratio for different EGR rates during dual-fuel operation (from ). 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 , 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 . 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) . 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 . 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 . Comparatively lower emissions of NOx and particulate matter were also recorded . 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 ). 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 . 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  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 . 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 . 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 . 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 . 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 . 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 . 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 . 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 . Combined experimental and computational studies are needed in order to develop better models to predict emissions for all engine fuels. Experimental studies similar to  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 References  Cho HM, He B. Spark ignition natural gas engines e a review. Energy Conversion and Management February 2007;48(2):608e18.  Population Division United States Census Bureau International Database. World population: 1950e2050. on-line statistics. Available on: www.census. gov, visited on 2010may20 at, http://www.census.gov/ipc/www/idb/ worldpopgraph.php.  British Petroleum. BP statistical review of world energy. BP Annual Review; June 2009.  Energy Information Administration-Ofﬁcial Energy Statistics from the U.S. Government. Brent-Europe crude oil spot prices 1987e2009; June 2009.  The European Parliament and the Council of the European Union. Regulation (EC) No 715/2007 of the European parliament and of the council of 20 June 2007 on type approval of motor vehicles with respect to emissions from light passenger and commercial vehicles (Euro 5 and Euro 6) and on access to vehicle repair and maintenance information. Ofﬁcial Journal 20 June 2007;OJL 171.  Korakianitis T, Wilson DG. Models for predicting the performance of Braytoncycle engines. Transactions of the ASME, Journal of Engineering for Gas Turbines and Power April 1994;116(2):381e8.  Korakianitis T, Grantstrom J, Wassingbo P, Massardo A. Parametric performance of combined-cogeneration power plants with various power and efﬁciency enhancements. Transactions of the ASME, Journal of Engineering for Gas Turbines and Power January 2005;127(1):65e72.  Korakianitis T, Meyer L, Boruta M, McCormick HE. Introduction and performance prediction of a nutating-disk engine. Transactions of the ASME, Journal of Engineering for Gas Turbines and Power April 2004;126(2):294e9.  Korakianitis T, Boruta M, Jerovsek J, Meitner PL. Performance of a single nutating disk engine in the 2 to 500 kW power range. Applied Energy October 2009;86(10):2213e21.  Massardo AF, Scialò M. Thermoeconomic analysis of gas turbine based cycles. Transactions of the ASME, Journal of Engineering for Gas Turbines and Power 2000;122:664e71.  Traverso A, Massardo AF. Optimal design of compact recuperators for microturbine application. Applied Thermal Engineering 2005;25:2054e71. 111  Pachidis V, Pilidis P, Talhouarn F, Kalfas A, Templalexis I. A fully integrated approach to component zooming using computational ﬂuid dynamics. Transactions of the ASME, Journal of Engineering for Gas Turbines and Power July 2006;128(3):579e84.  Løvås T, Mauss F, Hasse C, Peters N. Development of adaptive kinetics for application in combustion systems. Proceedings of the Combustion Institute 2002;29:1387e93.  Huang J, Crookes RJ. Assessment of simulated biogas as a fuel for the spark ignition engine. Fuel December 1998;77(15):1793e801.  Digest of United Kingdom energy statistics. Department for Business Enterprise and Regulatory Reform; 2008.  Heywood JB. Internal combustion engine fundamentals. McGraw-Hill; 1988.  Natural gas vehicle statistics. International Association for Natural Gas Vehicles; on-line statistics. Available on: www.iangv.org, visited on 2010may20 at http://www.iangv.org/content/view/17/35.  Turrio-Baldassarri L, Battistelli CL, Conti L, Crebelli R, De Berardis B, Iamiceli AL, et al. Evaluation of emission toxicity of urban bus engines: compressed natural gas and comparison with liquid fuels. Science of The Total Environment February 2006;355(1e3):64e77.  Van Herle J, Membrez Y, Bucheli O. Biogas as a fuel source for SOFC cogenerators. Journal of Power Sources March 2004;127(1e2):300e12.  Jayaratne ER, Ristovski ZD, Meyer N, Morawska L. Particle and gaseous emissions from compressed natural gas and ultralow sulphur diesel-fuelled buses at four steady engine loads. Science of The Total Environment April 2009;407(8):2845e952.  Lopez JM, Gomez A, Aparicio F, Sanchez FJ. Comparison of GHG emissions from diesel, biodiesel and natural gas refuse trucks of the City of Madrid. Applied Energy May 2009;86(5):610e5.  Dondero L, Goldemberg J. Environmental implications of converting light gas vehicles: the Brazilian experience. Energy Policy September 2005;33(13):1703e8.  Porpatham E, Ramesh A, Nagalingam B. Investigation on the effect of concentration of methane in biogas when used as a fuel for a spark ignition engine. Fuel July 2008;87(8e9):1651e9.  Bade Shrestha SO, Narayanan G. Landﬁll gas with hydrogen addition-a fuel for SI engines. Fuel December 2008;87(17e18):3616e26.  Crookes RJ. Comparative bio-fuels performance in internal combustion engines. Biomass & Bioenergy 2006;30:461e8.  Henham A, Makkar MK. Combustion of simulated biogas in a dual-fuel diesel engine. Energy Conversion and Management December 1998;39 (16e18):2001e9.  Dimopoulos P, Bach C, Soltic P, Boulouchos K. Hydrogenenatural gas blends fuelling passenger car engines: combustion, emissions and well-to-wheels assessment. International Journal of Hydrogen Energy December 2008;33 (23):7224e36.  D’Agosto M, Ribeiro SK. Assessing total and renewable energy in Brazilian automotive fuels. A life cycle inventory (LCI) approach. Renewable and Sustainable Energy Reviews AugusteSeptember 2009;13(6e7):1326e37.  Korakianitis T, Meyer L, Boruta M, McCormick HE. One-disk nutating-engine performance for unmanned aerial vehicles. Transactions of the ASME, Journal of Engineering for Gas Turbines and Power July 2004;126(3):475e81.  Korakianitis T, Meyer L, Boruta M, McCormick HE. Alternative multi-nutatingdisk engine conﬁgurations for diverse applications. Transactions of the ASME, Journal of Engineering for Gas Turbines and Power July 2004;126(3):482e8.  Hirose K, Ueda T, Takaoka T, Kobayashi Y. The high-expansion-ratio gasoline engine for the hybrid passenger car. JSAE Review January 1999;20(1):13e21.  Summary of fuel economy performance. National Highway Trafﬁc Safety Administration, U.S. Department of Transportation; December 2009.  Korakianitis T, Sadoi T. Turbocharger-design effects on gasoline-engine performance. Transactions of the ASME, Journal of Engineering for Gas Turbines and Power July 2005;127(3):525e30.  Sahari BB, Adlan H, Wong SV, Hamouda AM. Gear ratios strategy of PROTON Waja CNG-DI vehicle for improved performance. Modern Applied Science August 2009;3(8):63e71.  Karim GA. Combustion in gas fueled compression: ignition engines of the dual fuel type. Journal of Engineering for Gas Turbines and Power July 2003;125 (3):827e36.  Kiannejad F. Assessment of vegetable oil and emulsions for use as fuels in diesel engines. PhD thesis, Queen Mary and Westﬁeld College, University of London; 1993.  Crookes RJ, Korakianitis T, Namasivayam AM. A systematic experimental assessment of the use of rapeseed methyl ester (RME) as a compression ignition engine fuel during conventional and dual-fuel operation. In: TAE 7th international colloquium on fuels, Stuttgart, 14e15 Jan, 2009.  Transport statistics Great Britain (TSGB). UK Department for Transport; 2009.  Zeng K, Huang Z, Liu B, Liu L, Jiang D, Ren Y, et al. Combustion characteristics of a direct-injection natural gas engine under various fuel injection timings. Applied Thermal Engineering June 2006;26(8e9):806e13.  Thurnheer T, Soltic P, Eggenschwiler PD. SI engine fuelled with gasoline, methane and methane/hydrogen blends: heat release and loss analysis. International Journal of Hydrogen Energy March 2009;34(5):2494e503.  Mello P, Pelliza G, Cataluña R, da Silva R. Evaluation of the maximum horsepower of vehicles converted for use with natural gas fuel. Fuel October 2006;85(14e15):2180e6.  Evans RL, Blaszczyk J. A comparative study of the performance and exhaust emissions of a spark ignition engine fuelled by natural gas and gasoline. 112                         T. Korakianitis et al. / Progress in Energy and Combustion Science 37 (2011) 89e112 Proceedings of the Institution of Mechanical Engineers Part D-Journal of Automobile Engineering 1997;211(1):39e47. Bach C, Lammle C, Bill R, Soltic P, Dyntar D, Janner P, et al. Clean engine vehicle. A natural gas driven EURO 4/SULEV with 30% reduced CO2 emissions. SAE paper 2004-01-0645; 2004. Kesgin U. Effect of turbocharging system on the performance of a natural gas engine. Energy Conversion and Management January 2005;46(1):11e32. Kahraman N, Bilge Çeper B, Orhan Akansu SS, Aydin K. Investigation of combustion characteristics and emissions in a spark-ignition engine fuelled with natural gasehydrogen blends. International Journal of Hydrogen Energy January 2009;34(2):1026e34. Ma F, Ding S, Wang Y, Wang Y, Wang J, Zhao S. Study on combustion behaviors and cycle-by-cycle variations in a turbocharged lean burn natural gas SI engine with hydrogen enrichment. International Journal of Hydrogen Energy December 2008;33(23):7245e55. Dimopoulos P, Rechsteiner C, Soltic P, Laemmle C, Boulouchos K. Increase of passenger car engine efﬁciency with low engine-out emissions using hydrogenenatural gas mixtures: a thermodynamic analysis. International Journal of Hydrogen Energy September 2007;32(14):3073e83. Hu E, Huang Z, Liu B, Zheng J, Gu X. Experimental study on combustion characteristics of a spark-ignition engine fueled with natural gasehydrogen blends combining with EGR. International Journal of Hydrogen Energy January 2009;34(2):1035e44. Wang J, Chen H, Liu B, Huang Z. Study of cycle-by-cycle variations of a spark ignition engine fueled with natural gasehydrogen blends. International Journal of Hydrogen Energy September 2008;33(18):4876e83. Ma F, Wang Y. Study on the extension of lean operation limit through hydrogen enrichment in a natural gas spark-ignition engine. International Journal of Hydrogen Energy February 2008;33(4):1416e24. White CM, Steeper RR, Lutz AE. The hydrogen-fuelled internal combustion engine: a technical review. International Journal of Hydrogen Energy August 2006;31(10):1292e305. Abd-Alla GH. Using exhaust gas recirculation in internal combustion engines: a review. Energy Conversion and Management May 2002;43(8):1027e42. Arcoumanis C, Flora H, Kim JW, Xu HM. Injection natural gas engine for lightduty applications. In: International conference on 21st century emissions technology. London, England. December 04e05, 2000; 2000. Zheng JJ, Wang JH, Wang B, Huang ZH. Effect of the compression ratio on the performance and combustion of a natural-gas direct-injection engine. Proceedings of the Institution of Mechanical Engineers Part D-Journal of Automobile Engineering January 2009;223(D1):85e98. Huang Z, Wang J, Liu B, Zeng K, Yu J, Jiang D. Combustion characteristics of a direct-injection engine fueled with natural gasehydrogen blends under different ignition timings. Fuel February 2007;86(3):381e7. Borman GL, Ragland KW. Combustion engineering. WCB/McGraw-Hill; 1998. Crookes RJ, Bob-Manuel KDH, Olsen J. Renewable fuelling of a diesel engine in single and dual fuel operation. Technische Akademie Esslingen (TAE) 6th international colloquium fuels, 10e11 January 2007. Kusaka J, Okamoto T, Daisho Y, Kihara R, Saito T. Combustion and exhaust gas emission characteristics of a diesel engine dual-fuelled with natural gas. JSAE Review October 2000;21(4):489e96. Papagiannakis RG, Hountalas DT. Combustion and exhaust emission characteristics of a dual fuel compression ignition engine operated with pilot diesel fuel and natural gas. Energy Conversion and Management November 2004;45 (18e19):2971e87. Carlucci AP, Ficarella A, Laforgia D. Control of the combustion behavior in a diesel engine using early injection and gas addition. Applied Thermal Engineering December 2006;26(17e18):2279e86. Sahoo BB, Sahoo N, Saha UK. Effect of engine parameters and type of gaseous fuel on the performance of dual-fuel gas diesel engines e a critical review. Renewable and Sustainable Energy Reviews AugusteSeptember 2009;13 (6e7):1151e84. Papagiannakis RG, Hountalas DT. Experimental investigation concerning the effect of natural gas percentage on performance and emissions of a DI dual fuel diesel engine. Applied Thermal Engineering February 2003;23 (3):353e65. Selim MYE. Pressureetime characteristics in diesel engine fueled with natural gas. Renewable Energy April 2001;22(4):473e89. Karim GA. A review of combustion processes in the dual fuel engine e the gas diesel engine. Progress in Energy and Combustion Science 1980;6 (3):277e85. Shenghua L, Longbao Z, Ziyan W, Jiang R. Combustion characteristics of compressed natural gas/diesel dual-fuel turbocharged compressed ignition engine. Journal of Automobile Engineering 2003;217(9):833e8. Duc PM, Wattanavichien K. Study on biogas premixed charge diesel dual fuelled engine. Energy Conversion and Management August 2007;48(8): 2286e308.  Selim MYE. Effect of engine parameters and gaseous fuel type on the cyclic variability of dual fuel engines. Fuel May 2005;84(7e8):961e71.  Namasivayam AM, Crookes RJ, Korakianitis T, Olsen J. Assessment of combustion in natural gas fuelled compression ignition engines with DME and RME pilot ignition. Proceedings of the Institution of Mechanical Engineering, International Journal of Engine Research 2009;10(3):165e74.  Selim MYE. Sensitivity of dual fuel engine combustion and knocking limits to gaseous fuel composition. Energy Conversion and Management February 2004;45(3):411e25.  Pirouzpanah V, Khoshbakhti Saray R, Sohrabi A, Niaei A. Comparison of thermal and radical effects of EGR gases on combustion process in dual fuel engines at part loads. Energy Conversion and Management July 2007;48 (7):1909e18.  McTaggart-Cowan GP, Jones HL, Rogak SN, Bushe WK, Hill PG, Munshi SR. Direct-injected hydrogenemethane mixtures in a heavy duty compression ignition engine. SAE paper 2006-01-0653; 2006.  Selim MYE. Effect of exhaust gas recirculation on some combustion characteristics of dual fuel engine. Energy Conversion and Management March 2003;44(5):707e21.  Abd-Alla GH, Soliman HA, Badr OA, Abd-Rabbo MF. Effects of diluent admissions and intake air temperature in exhaust gas recirculation on the emissions of an indirect injection dual fuel engine. Energy Conversion and Management May 2001;42(8):1033e45.  Srinivasan KK, Krishnan SR, Midkiff KC. Improving low load combustion, stability, and emissions in pilot-ignited natural gas engines. Journal of Automobile Engineering 2006;220(2):229e39.  Tree DR, Svensson KI. Soot processes in compression ignition engines. Progress in Energy and Combustion Science June 2007;33(3):272e309.  Nwafor OMI. Effect of advanced injection timing on emission characteristics of diesel engine running on natural gas. Renewable Energy November 2007;32 (14):2361e8.  Olsen J, Crookes RJ, Bob-Manuel KDH. Experiments in dual fuelling a compression ignition engine by injecting di-methyl ether as a pilot fuel to ignite varying quantities of natural gas. SAE paper 2007-01-3624; August 2007.  Papagiannakis RG, Hountalas DT, Rakopoulos CD. Theoretical study of the effects of pilot fuel quantity and its injection timing on the performance and emissions of a dual fuel diesel engine. Energy Conversion and Management November 2007;48(11):2951e61.  Namasivayam AM, Korakianitis T, Crookes RJ, Bob-Manuel KDH, Olsen J. Biodiesel, emulsiﬁed biodiesel and dimethyl ether as pilot fuels for natural gas fuelled engines. Applied Energy Mar 2010;87(3):769e78.  Crookes RJ, Bob-Manuel KDH. RME or DME: a preferred alternative fuel option for future diesel engine operation. Energy Conversion and Management November 2007;48(11):2971e7.  Banapurmath NR, Tewari PG. Comparative performance studies of a 4-stroke CI engine operated on dual fuel mode with producer gas and Honge oil and its methyl ester (HOME) with and without carburetor. Renewable Energy April 2009;34(4):1009e15.  Nwafor OMI. Effect of choice of pilot fuel on the performance of natural gas in diesel engines. Renewable Energy November 2000;21(3e4):495e504.  Banapurmath NR, Tewari PG, Hosmath RS. Experimental investigations of a four-stroke single cylinder direct injection diesel engine operated on dual fuel mode with producer gas as inducted fuel and Honge oil and its methyl ester (HOME) as injected fuels. Renewable Energy September 2008;33(9):2007e18.  Selim MYE, Radwan MS, Saleh HE. Improving the performance of dual fuel engines running on natural gas/LPG by using pilot fuel derived from jojoba seeds. Renewable Energy June 2008;33(6):1173e85.  Crookes RJ, Kiannejad F, Nazha G. Seed-oil bio-fuel of low cetane number: the effect of water emulsiﬁcation on diesel-engine operation and emissions. Journal of Institute of Energy 1995;68:142e51.  McTaggart-Cowan GP, Jones HL, Rogak SN, Bushe WK, Hill PG, Munshi SR. The effects of high-pressure injection on a compression-ignition, direct injection of natural gas engine. Journal of Engineering for Gas Turbines and Power April 2007;129(2):579e88.  Milton BE, Pianthong K. Pulsed, supersonic fuel jets e a review of their characteristics and potential for fuel injection. International Journal of Heat and Fluid Flow August 2005;26(4):656e71.  Mbarawa M, Milton BE, Casey RT. Experiments and modelling of natural gas combustion ignited by a pilot diesel fuel spray. International Journal of Thermal Sciences 2001;40(10):927e36.  Mbarawa M, Milton BE, Casey RT. An investigation of the effects of diesel pilot injection parameters on natural gas combustion under diesel conditions. Journal of the Institute of Energy September 2001;74(500):81e90.  Lu T, Law CK. Toward accommodating realistic fuel chemistry in large-scale computations. Progress in Energy and Combustion Science April 2009;35 (2):192e215.