Title of Invention

METHOD AND APPARATUS OF FUELLING AN INTERNAL COMBUSTION ENGINE WITH HYDROGEN AND METHANE

Abstract

ABSTRACT METHOD AND APPARATUS OF FUELLING AN INTERNAL COMBUSTION ENGINE WITH HYDROGEN AND METHANE A gaseous-fuelled internal combuslion engine and a method of operating same arc provided for improving combustion stability and reducing emissions of NOx, PM, and unbumcd hydrocarbons. The method comprises fuelling an internal combuslion engine with hydrogen and natural gas, which can be directly injected into the combuslion chamber together or introduced separately. Of the total gaseous fuel delivered to ihe engine, at least 5% by volume at standard temperature and pressure is hydrogen. For at least one engine operating condition, the ratio of fuel rail pressure to peak in-cylinder pressure is al least 1.5: 1. The engine comprises a combustion chamber defined by a cylinder, a cylinder head, and a piston movable within the cylinder; and a fuel injection valve operable for introducing the gaseous fuel mixture directly into the combustion chamber, or two separate fuel injection valves for introducing the methane and hydrogen separately. An electronic controller is in communication with actuator(s) for the fuel injection valve{s) for controlling timing for operating the fuel injection valvc(s). thc engine has a compression ratio of at least 14:1.

Full Text

METHOD AND APPARATUS OF FUELLING AN INTERNAL COMBt/STrON ENCrrVE WITH HYDROGEN AND METHANE Field of the Invention [0001] The present invention relates to a method and apparatus of fiielling adiesel-cycle internal combustion engine with hydrogen and methane to improve combustion stability and reduce emissions of nitrogen oxides (NOx), unbumed hydrocarbons and particulate matter (PM). Backt»round of the Invention 10002) Because gaseous fuels such as natural gas, propane, hydrogen, and blends thereof are cleaner burning fuels compared to liquid fiiels such as diesel, recent attention has been directed to developing engines that can bum such fiiels while matching the power and performance that engine operators are accustomed to expecting from diesei engines. [0003] Natural gas ftielled engines that use lean-bum spark-ignition ("LBSi") introduce the fuel into the intake air manifold or intake ports at relatively low pressures. To avoid engine knock caused by the premature detonation of the fiiel inside the combustion chamber, such engines typically operate with a compression ratio no greater than about 12:1, wiiich is lower compared to diesel-cycle engines whichhavecompressionratiosof at least 14:1, and this affects engine performance and efficiency. Consequently, while the exhaust gases from the combustion chambers of LBSl engines can have lower emissions of NOx:, and PM compared to an equivalently sized diesel engine, such LB A engines also have lower performance and energy efficiency, which means that to do the same amount of work, more fuel is consumed on an energy basis, and to match the full range of power and performance of a diesel engine, a larger LBSI engine is needed. [0004| Recently, research has been directed towards blending natural gas and hydrogen for use in homogeneous charge, spark-ignition engines. Representative publications relating to such research include, "The Effects of Hydrogen Addition On Natural Gas Engine Operation", SAE Technical Paper 932775, by M,R. Swain, M.J. Yusuf, Z. Dulgerand M,N. Swain, which was published by the Society of Automotive Engineers ("SAE") in 1993; "Variable Composition Hydrogen/Natural Gas Mixtures for Increased Engine Efficiency and Decreased Emissions", ASME Journal of Engineering f6r Gas Turbines and Power, Vol. 122, Pp. 135-140, by R. Sicrens and E. Rousseel, published in 2000; "Hydrogen Blended Natural Gas Operation of a Heavy Duty Turbocharged Lean Bum Spark Ignition Engine", SAE Technical Paper 2004-01-2956, by S. R. Munshi, C. Nedelcu, J- Harris, et a!., published m 2004; "Hydrogen Enrichment: A Way to Maintain Combustion Stability in a Natural Gas Fuelled Engine with Exhaust Gas Recirculation, the Potential of Fuel Reforming", Proceedings of ihe Institution of Mechanical Engineers, Part D. Vol. 215 2001, Pp. 405-418, by S. Allenby, W-C. Chang, A. Megaritis and M.L. Wyszynski; "Emission Results from the New Development of a Dedicated Hydrogen-Enriched Natural Gas Heavy-Duty Engine", SAE Technical Paper 2005-010235, by K. Collier, N. Mulligan, D. Shin, and S. Brandon which was published in 2005; "Compari.sons of Emissions and Efficiency of a Turbocharged Lean-Bum Natural Gas and Hythane-Fuelled Engine", ASME Journal of Engineering for Gas Turbines and Power, Vol. 119,1997, Pp. 218-226, by J,F. Larsen and J.S. Wallace; "Effect of hydrogen addition on the performance of methane-fuelled vehicles Part 1: effect on S.I, engine performance", tnttmalionaJ Journal of Hydrogen Energy, Vol, 26, 2001, Pp. 55-70, by C.G. Bauer and T.W. Forest; "Methane-Hydrogen Mixtures as Fuels", International Journal of Hydrogen Energy, Vol. 21 No. 7, 1996, Pp. 625-631, by G.A. Karim, L Wierzba and Y. Al-Alousi; and "Internal Combustion Engines Fuelled by Natural Gas-Hydrogen Mixtures", International Journal of Hydrogen Energy, Vol. 29, 2004, Pp. 1527-1539, by S.O. Akansu, Z. Duiger, N. Kahraman and T. Veziroglu. The results reported in these papers have shown that at stoichiometric operation, the addition of hydrogen lends to reduce power density and increase NOx, while slightly reducing hydrocarbon and carbon monoxide emissions, A more significant effect is reported under lean premixed conditions, where a substantia! increase in the lean limit is observed. This has been attributed to enhanced combustion rate and shorter ignition delay. For a given air-fuel ratio, NOx emissions are higher with hydrogen addition, due to the higher flame temperature, while CO and unbtimed hydrocarbons are substantially reduced. However, due to hydrogen's ability to extend the lean limit, lower NOx emissions can be achieved by running at leaner air-fuel ratios with hydrogen addition. Flame stability in the presence of exhaust gas recirculation (EGR) is also improved. Efficiency effects can depend upon the tested operating condition, with some studies such as those reported in the Swain, Sierens, and Akansu papers, showing improved efficiency with hydrogen addition and other studies, such as those reported in the Larsen and Bauer papers, showing reduced efficiency. Such contradictory results show that while a considerable amount of research has been done to investigate the effects of blending natural gas and hydrogen for use in homogeneous charge spark-ignition engines, the combustion process is complex, that the effect of combusting such fiiel mixtures in an engine can be very dependent upon the engine operating conditions, and that the effect of adding hydrogen and the magnitude or such effects, if any, are not obvious.or easy to predict, furthermore, all of the published papers referenced herein relate to homogeneous charge spark-ignition engines, and while some laboratory experiments have been reported, such as shock-tube studies and non-premixed counterflow methane/heated air jet experiments, the inventors are not aware of any publications relating to experiments involving fuelling a direct injection internal combustion engine with a blended fuel mixture comprising methane and hydrogen. |0005] Engines that are capable of injecting a gaseous fuel directly into the combustion chamber of a high compression internal combustion engine are being developed, but are not yet commercially available. Engines fuelled with natural gas that use this approach can substantially match the power, performance and efficiency characteristics of a diesel engine, but with lower emissions of NOx, unbumed hydrocarbons, and PM. NOx are a key component in the formation of photochemical smog, as well as being a contributor to acid rain. PM emissions, among other detrimental health effects, have been linked to increased cardiovascular mortality rates afld impaired lung development in children. However, with direct injection engines that are fuelled with natural gas, it has been found that there is a trade-off between NOx emissions and emissions of unbumed hydrocarbons and PM. ITiat is, later timing for injecting the natural gas is beneficial for reducing NOx but results in higher emissions of unbumed hydrocarbons and PM. Environmental regulatory bodies in North America and around the world have legislated substantia] reductions in NOx and PM emissions from internal combustion engines. As a result, because it is necessary to reduce the emissions of each one of NOx, PM and unbumed hydrocarbons, for a direct injection engine fiielled with natural gas, the higher PM emissions associated with later combustion timing effectively limits how much the timing for fuel injection can be retarded. [0006] Since published technical papers have reported that under specific operating conditions there can be benefits arising from fuelling a homogeneous charge, spark-ignition engine with a gaseous fuel mixture comprising methane and hydrogen, and since environmental regulatory bodies have legislated substantial reductions in NOx and PM emissions from intemai combustion engines, and since the combustinn process is complex and the effect of adding hydrogen to a fiiel mixture delivered to a direct injection intemai combustion engine is unpredictable, there is a need to determine whether it is possible to improve combustion stability and reduce engine emissions by fuelling a direct injection intemai combustion engine with hydrogen and natural gas, and if so, the method of operating a direct injection engine that is flieiled with such fuels to achieve improvements in combustion stability and reductions in engine emissions. Suimnarv of die Invention [00071 A method is provided of operating, a direct injection intemai combustion engine. The method comprises introducing a gaseous fuel mixture directly into a combustion chamber of the engine. The gaseous fijel mixture comprises methane and between 5% and 50% hydrogen by volume at standard temperature and pressure. For at least one engine operating condition, the method comprises maintaining a fuel rail to peak in-cylinder pressure ratio of at least 1,5:1 when introducing the gaseous fuel mixture into the combustion chamber. A preferred embodiment of the method comprises maintaining a fiiel rail to peak in-cylinder pressure ratio of at least 1,5 ;1 when introducing the gaseous fuel mixture into the combustion chamber for all engine operating conditions. When the constituent parts of the gaseous fuel mixture are described herein as percentages by volume, unless noted otherwise this is defined to be the percentage by volume at standard temperature and pressure (STP). fO0O8| In preferred methods, the gaseous fiiel mixture can comprise between 10% and 50%, between 15% and 40% hydrogen by volume, or between 20% and 35% hydrogen by volume at standard temperature and pressure. The methane can be a constituent part of natural gas. The method can further comprise premixing the gaseous fuel mixture and storing it as a blended fuel within a storage lank from which it can be delivered to the engine. In a preferred method, methane is the largest constituent of the gaseous fuel mixture by volume at standard temperamre and pressure. [00091 The method can further comprise controlling fuel injection timing so that the mid-point of integrated combustion heat release occurs between 2 and 30 crank angle degrees after top dead center. An advantage of adding hydrogen to natural gas is that the combustion timing can be delayed to a later time in the combustion cycle compared to an engine that is fuelled with natural gas alone. A preferred method comprises controlling fuel injection timin so that in at least one engine operating condition the mid-point of integrated combustion heat release occurs between 5 and ] 5 crank angle degrees after top dead center. lOOlO) The method can comprise introducing a pilot fuel to assist with ignition of the gaseous fuel mixture. A preferred method comprises injecting a pilot fuel directly into the combustion chamber about 1 millisecond before start of injection of the gaseous fuel mixture. The pilot fuel can be a liquid fuel with a cetane number between 40 and 70. A pilot fuel with a cetane number between 40 and 50 is preferred in most cases, with conventional road grade diesel being a suitable fuel with a ccte number in this range. Over an engine operating map the pilot fuel is on average between 3% and 10% of the fuel that is consumed by the engine on an energy basis, and more between 4% and 6%. The pilot fuel is more easily ignited compared to the gaseous fijel mixture, and the pilot fuel ignites first to trigger the ignition of the gaseous fuel mixture. Because the gaseous fiiel mixture is preferably cleaner burner than the pilot fiiei, the pilot fuel preferably represents only a small portion of the fuel that is consumed by the engine on an energy basis. (OOIII Instead of employing a pilot fud, the method can comprise heating a hot surface inside the combustion chamber to assist with igniting the gaseous fuel mixture. In a preferred method the hot surface is provided by a glow plug and the method further comprises electrically heating the glow plug. In yet another embodiment, the method can comprise spark igniting the gaseous fuel mixture inside the combustion chamber. 100121 The method can ftsrthei comprise storing Uie hydrogen separately from the methane and mixing the hydrogen and methane to form the gaseous fuel mixture. The method can further comprise controlling the proportions of hydrogen and methane in the gaseous fuel mixture as a function of ;ngine operating conditions. |0013) The method can further comprise maintaining a fuel rail to peak in*cylinder pressure ratio of at least 2:1 when introducing the gaseous fiiei mixture into the combustion chamber for at least one engine operating condition. Preferred methods comprise maintaining a choked flow condition at a nozzle orifice of a fuel injection valve when introducing the gaseous fuel mixture into the combustion chamber. While experiments have proven that satisfactory engine operation can be achieved by injecting the gaseous fuel mixture into the combustion chamber with an injection pressure that is at least 16 MPa (about 2350 psia), higher fuel injection pressures of at least 20 MPa (about 2900 psia) are more preferred. 100141 According to the method, in the course of a compressioti stroke, an intake charge inside the combustion chamber is compressed by a ratio of at least about 14:1. Compression ratios higher than !4:I are associated with diesel-cycle engines, which can deliver higher performance and efficiency than conventional Otto-cycle engines, otherwise known as spark-ignition engines, which use a pre-mixed homogeneous tharge which limits them to lower compression ratios to avoid engine knock. [OOlSj In another preferred method of fueUing an intemai combustion engine, the method comprises introducing a gaseous fuel mixture directly into a combustion chamber of the engine, wherein the gaseous fuel mixture comprises methane, introducing hydrogen into the combustion chamber, thereby adding hydrogen to the gaseous fuel mixture, wherein the hydrogen represents at least 5% by volume of the gaseous fiiel mixture at standard temperature and pressure; and maintaining a gaseous fuel mixture rail to peak in-cylinder pressure ratio of at least 1.5:1 when introducing the gaseous fuel mixture into the combustion chamber for at least one engine operating condition. That is, the hydrogen can be introduced into the combustion chamber separately from the gaseous fuf mixture and becoming part of the gaseous fuel mixture inside the combustion chamber or the method can comprise premixing the hydrogen with the gaseous fuel mixture comprising methane, and introducing tlie gaseous fuel mixture and the hydrogen directly into the combustion chamber. In further embodiments, the method can comprise premixing the hydrogen with intake air and mtroducing the hydrogen into the combustion chamber during an intake stroke of the piston or introducing the hydrogen directly into the combustion chamber separately from the gaseous fuel mixture.. [0016] An interna! combustion engine is provided that can be fuelled with a gaseous fuel mixture comprising methane and between 5% and 60% hydrogen by volume at standard temperature and pressure. The disclosed engine comprises a combustion chamber defined by a cylinder, a cylinder bead, and a piston movable within the cylinder; a fiiel injection valve with a nozzle that is disposed within the combustion chamber, the fuel injection valve being operable to introduce the gaseous fuel mixnire directly into the combustion chamber; a pressurizing device and piping for delivering ihe gaseous fuel mixture to the injection valve with a ratio of fuel rail to peak in-cylinder pressure being at least I.5;l for at least one engine operating condition; and, an electronic controller io communication with an actuator for the fuel injection valve for controlling timing for operating the fuel injection valve. The engine preferably has a compression ratio of at least 14. |0017| The electronic controller is preferably programmable to time introduction of the gaseous fuel mixture into the combustion chamber so that the mid-point of an integrated combustion heat release occurs between 2 and 30 crank angle degrees afler top dead center, and in another embodiment, between 5 and IS crank angle degrees after top dead center. [0018] The fuel injection valve can be mounted in the cylinder head with the fuel injection valve comprising a nozzle disposed within the combustion chamber. The engine can further comprise a second fiiel injection valve that is operable to introduce a pilot fuel directly into the combustion chamber. The second fuel injection valve can be integrated into a valve assembly that also comprises the fuel injection valve for iniToducing the gaseous fuel mixture. The second ftiel injection valve and the fiiel injection valve for introducing the gaseous fiiel mixture are preferably independently acmated and the gaseous fiiel mixture is injectable into the combustion chamber through 3 first set of nozzle orifices, which are different from a second set of nozzle orifices through which the pilot fuel is injectable into the combustion chamber. [0019] Instead of employing a second fuel injection valve to introduce a pilot fuel to assist with ignition of the gaseous fuel mixture, the engine can comprise an ignition plug disposed within the combustion chamber that is operable to assist with ignition of the gaseous fiiel mixture. The ignition plug can be a glow plug that is electrically beatable to provide a hot surface for assisting with ignition of the gaseous fuel mixture or the ignition plug can be a spark plug. [0020! The engine can further comprise a storage vessel for storing the gaseous flic! mixture in a substantially homogeneous mixture with predetermined proportions of hydrogen and methane, in another embodiment, the engine can comprise a first storage vessel within which the hydrogen can be stored, a second storage vessel within which a gaseous iiiel comprising methane can be stored, and valves associated with each one of the first and second storage vessel that are operable to control respective proportions of hydrogen and methane in the gaseous fuel mixture that is introducible into the combustion chamber. If the hydrogen is stored separately from the gaseous fiiel mixture that comprises methane, then the electronic controller can be programmable to change respective proportions of hydrogen and methane in the gaseous fiiel mixture to predetermined amounts responsive to detected engine operating conditions. [0021] Another embodiment of an internal combustion engine is provided that can be fuelled with a gaseous fuel mixUire comprising methane and hydrogen. In this embodiment, the engine comprises a con. -ustion chamber defined by a cylinder, a cylinder head, and a piston movable within the cyhnder; a first fuel injection valve with a nozzle disposed within the combustion chamber, wherein the fuel injection valve is operable to mtroduce methane directly into the combustion chamber; a second fuel injection valve with a nozzle disposed within an intake air manifold, wherein the second fiiel injection valve is operable to introduce hydrogen into the intake air manifold from which the hydrogen can flow into the combustion chamber; and an electronic controller in communication with an actuator for each one of the first and second fuel injection valves for controlling respective timing for operating the first and second fijef injection valves. In this embodiment, the engine can further comprise a pressurizing device and piping for delivering the methane to the first injection valve with a ratio of fiiel rail pressure to peak in-cylinder pressure being at least 1.5: i for at least one engine operating condition. Like the other embodiments, the engine preferably has a compression ratio of at least 14:1, and compression ratios as high as 25:! are possible as well as ratios therebetween, such as 18; 1, 20; 1 and 22:1. Brief Description of the Drawings [00221 Figure 1 is a schematic drawing illustrating an apparatus for direct injection of a gaseous fuel mixture into the combustion chamber of an internal combustion engine. |0023] Figure 2 is a schematic drawing illustrating a second embodiment of an apparatus for direct injection of a gaseous fuel mixture into the combustion chamber of an internal combustion engine. |0024] Figure 3 shows four graphs that plot engine emissions against timing for the mid-point of a combustion heat release for an engine that is fuelled with 100% compressed natural gas, a gaseous fiiei m-xture of 10% hydrogen and 90% compressed namral gas, and 23% hydrogen and 77% compressed natural gas, with all percentages measured by volume. The plotted data was collected from an engine operating at 800 RPM, 6 bar GIMEP, O-Si]), 40% exhaust gas recirculation {by mass), and with a fuel injection pressure of 16 MPa. [0025) Figure 4 shows four graphs that plot engine performance characteristics against timing for the mid-point of a combustion heat release for an engine that is ftielled with 100% compressed natural gas, a gaseous fuel mixture of 10%. hydrogen and 90% compressed natural gas, and 23% hydrogen and 77%o compressed natural gas, with al! percentages measured by volume. The engine operating conditions were the same as for the data plotted in Figure 3. [0026] Figure 5 shows two bar graphs tha. plot pilot and gaseous fuel ignition delay for an engine that is fuelled with 100% compressed natural gas, a gaseous fticl mixture of 10% hydrogen and 90% compressed natural gas, and 23% hydrogen and 77% compressed natural gas, with all percentages measured by volume. The engine operating conditions were the same as for the data plotted in Figures 3 and 4. [0027] Figure 6 shows plots of in-cylindcr pressure and heat release rate for an engine that is ftielled with 100% compressed natural gas, and a gaseous fuel mixture of 10% hydrogen and 90% compressed naWral gas with percentages for the gaseous Eiiel mixture measured by volume and the timing for the mid-point of the integrated heat release occurring at 10 crank angle degrees after top dead center. The engine operating conditions were the same as for the data plotted in Figures 3 through 5. [0028) Figure 7 is plots of in-cylinder pressure and heat release rate for an engine that is fuelled with 100% compressed napiral gas, and a gaseous fiiel mixture of 23% hydrogen and 77% compressed natural gas, with percentages for the gaseous fael mixture measured by volume. In the upper two graphs, the timing for the mid-point of the integrated heat release occurs at 5 crank angle degrees after top dead center and in the lower two graphs the timing for the mid-point of the integrated heat release occurs at 15 crank angle degrees after top dead center. The engine operating conditions were the same as for the data plotted in Figures 3 through 6, |0O29| Figure 8 plots in-cylinder pressure and heat release rate for constant and adjusted timing conditions for an engine fuelled with 100% compressed natural gas, and a gaseous fuel mixture of 23% hydrogen and 77% compressed natural gas, with percentages for the gaseous fuel mixture measured by volume. The engine operating conditions were the same as for the data plotted in Figures 3 through 7. [0030] Figure 9 is a bar graph that plots gaseous fuel ignition delay in crank angle degrees against timing for the raid-point of integrated heat release for 100% natural gas, and a gaseous fiiel mixture of 23% hydrogen and 77% compressed natural gas, with percentages for the gaseous fuel mixture measured by volume. The plotted data shows the gaseous fuel ignition delay when the gaseous fijel mixture is injected with the same timing that is employed when the engine is fuelled with 100% natural gas, and the effect on gaseous fuel ignition delay when timing is adjusted. The engine operating conditions were the same as for the data plotted in Figures 3 through 8, [0031] Figure 10 plots engine emissions against the timing for the mid-point of the integrated heat release, showing the effect of increasing injection pressure from 16 MPa to 20 MPa for an engine fuelled with 100% compressed natural gas, and a gaseous fitel mixture of 23% hydrogen and 77% compressed natural gas, with percentages for the gaseous fuel mixture measured by volume. Other than the different fuel injection pressures, the engine operating conditions were the same as for the data plotted in Figures 3 though 9. |0032] Figure 11 plots in-cylinder pressure and heat release rate against crank angle degrees for fuel injection pressures of 16 MPa and 20 MPa for an engine fuelled with 100% compressed natural gas, and a gaseous fuel mixture of 23% hydrogen and 77% compressed natural gas, with percentages for the gaseous fuel mixture measured by volume. The engine operating conditions were the same as for the data plotted in Figures 3 through 9. Detailed Description of Preferred Embodiment(s) |0 cools the gaseous fuel mixture after it has been compressed by compressor 112. Aftercooler 113 can be a heal exchanger with the cooling fluid being engine coolant diverted from the engine cooling system. In a preferred embodiment, the fuel supply system is a common rail system, meaning that the gaseous fuel is delivered to fiiel injection valve assembly [20 at injection pressure. In such a common rail system, pressure sensor 1! 5 can be employed to measure the fuel pressure in gaseous-fijel supply rdil 116 so that compressor 112 can be operated to maintain gaseous fiiel injection pressure between a predetermined low and high set point. [0034| In a preferred embodiment, a liquid pilot fuel is also directly injected into combustion chamber 122 to assist with igniting the gaseous fuel mixture. In such an embodiment, injection valve assembly 120 car comprise two separate valve needles that are independently operable, with one valve needle controlling the injection of a gaseous fiiel mixture and the second valve needle controlling the injection of the liquid pilot fuel. Pilot fuel is deliverable to fiiel injection valve assembly 120 from pilot fuel rail US. Pilot fuel can be delivered to pilot fuel rail US at injection pressure by a conventional diesel common rail fuel supply system (not shown). |003S] Fuel injection valve 120 introduces the gaseous fiiel mixture directly into combustion chamber 122, which is generally defined by a bore provided in cylinder block 124, the cylinder head, and piston 126, which is movable up and down within the boce. The flow of air into combustion -.hamber 122 from intake air manifold 130 is controlled by intake valve 132, which can be opened during intake strokes of piston 126. Like conventional diesel engines, the disclosal engine can employ a turbocharger (not shown) to pres-iurize the intake air or the engine can be naturally aspirated. Combustion products can be expelled from combustion chamber 122 into exhaust manifold 140 through exhaust valve 142, which can be opened during exhaust strokes of piston 126. [0036] Electronic controller 150 is programmable to control the operation of compressor 1 i 2 and control valve 114 to control the pressure of the gaseous fuel mixture in gaseous ftiel supply rail 116. Controller 150 is also programmable to command the timing for opening and closing of the fuel injection valve needles that respectively control the injection of the gaseous fiiel mixture and the pilot fuel. For example, electronic controller 150 can be programmed to control the pilot fUel injection valve so that the pilot fbel is introduced about 1 millisecond before the gaseous fiiel injection valve is commanded to open. Furthermore, electronic controller 150 can be programmed to time the opening and closing of the gaseous fuel injection valve. The fuel injection timing can be predetermined responsive to the engine operating conditions determined from measured parameters that are inputted into electronic controller 150, and the input of such parameters is represented by arrow 152. [0037] Figure 2 is a schematic drawing of another preferred embodiment for an engine apparatus that is adapted to be fiielled with a gaseous fuel mixture comprising methane and hydrogen. In this embodiment, fiiel injection valve 220 injects only the gaseous fijel mixture into combustion chamber 222. A pilot fuel is not required by this engine because ignition assistance is provided by ignition plug 228. Ignition plug 228 can be an electrically heated glow plug that is adapted for sustained operation during engine operation. This is unlike a conventional glow plug, which is normally activated only under certain engine conditions such as start-up when the engine block is below a predetermined temperature. Compared to an engine that is fuelled with natural gas without any added hydrogen, an advantage of using a gaseous fuel mixture comprising hydrogen is that because hydrogen is easier to ignite compared to natural gas, the glow plug temperature can be kept at a lower temperamre compared to the temperature that is needed to assist with ignition of natural gas which is not mixed with hydrogen. This is advantageous because lower glow plug temperatures are generally associated with improved durability and longer service life. In yet another embodiment (not shown), the ignition plug can be a spark plug, [0038] In the illustrated embodiment of Figure 2, other than using ignition plug 228 to assist initiating fiiel combustion instead of a pilot fuel, the shown engine apparatus is essentially the same. That is, a bore in cylinder block 224, cylinder head 225, and piston 226, which is movable up and down within the cylinder bore, all cooperate to define combustion chamber 222. Air car flow into combustion chamber 222 through intake air manifold 230 when intake valve 232 is open and electronic controller 250 is programmable to control the timing for opening and closing fiiet injection valve 220, and to mntrol the temperature of ignition plug 228. [0039| Figure 2 also shows an optional secondary fiel injection valve 240 which can he employed to inject some of the gaseous fuel into the intake air manifold. A port fuel injection valve is shown, hut a single fuel injection valve can be disposed further upstream in the intake air manifold for introducing gaseous fuel into all of the combustion chambers. Secondary fuel injection valve 240 can be employed to introduce hydrogen into the combustion hamber separately from a gaseous fuel mixture comprising methane, such as natural gas. With such an embodiment, the methane and hydrogen mixes inside the combustion chamber, but with the hydrogen mote evenly dispersed within the combustion chamber. AK advantage of this arrangement is that the hydrogen need not be compressed to as high a pressure as it would need to be pressurized for direct injection. Another advantage is that separately injecting the methane and hydrogen allows the proportions of each fuel to be adjusted for different engine operating conditions. However, a disadvantage of this approach is that a secondary fijel injection valve adds complexity and capital and maintenance costs to the engine. [0040] To determine a desired method of operating an internal combustion engine with direct injection of a gaseous fuel mixture comprising methane and hydrogen, experiments were conducted using a single cylinder engine. The single cylinder engine was a Cummins™ ISX series heavy-duty six-cylinder, four stroke, direct injection diesel engine, modified to operate on only one cylinder. The engine was further adapted for gaseous fuel operation using Westport'"" fuel injection and fuel supply systems. The engine cylinder bote diameter was 13? millimeters, the piston stroke was 169 millimeters, and the displacement of the single cyiinder was about 2.5 liters. The connecting rod length was 261 millimeters and the compression ratio was 17:1. [0041] Because the experimental engine was a single cylinder engine, the energy in the exhaust stream was too small to drive a turbocharger to compress the intake air. To simulate the conditions for a turbocharged engine, in the experiments an air compres,sor was provided for the combustion air supply. The air compressor was equipped \u\\h a refrigerated air dryer to remove water vapour (de%w point -40*0) ai\d filters to remove contaminants. The EG loop comprised an EGR cooler and a variable flow-control valve. Maintaining the exhaust stream pressure approximately 10 kPa above the intake pressure drove the recirculation of the exhaust gas. [0042] The fuelling system provided gaseous fuel and diesel to the engine's internal fuelling rails. The fuel injection valve was a dual fUel injection valve operable to separately and independently inject the gaseous fuel mixture as the main fuel, and diesel fuel as the pilot fuel, with flow of the main, and pilot fuels being controlled by two concentric valve needles. Separate solenoid actuated control valves were operable to control the actuation of each valve needle to control the timing and duration of the respective pilot and main fuel injection events. The pilot fuel injection valve comprised a nozzle with 7 orifices, and the gaseous fuel injection valve comprised a nozzle with 9 orifices, and the injection angle was 18 degrees below the tiredeck. Two separate gaseous fiiel supplies were used in the experiments. Commercial natural gas (-96 moI% CHA, 2% CiH, Q-aces Ni, COi, CjHgal! Pollution Instruments). A second infrared analyzer (California Analytical) was used to measure the CO2 concentration in the intake stream, from which the EGR fraction was determined. A chilled water separator removed water vapour (dew point -SC) upstream of the non-dispersive infrared instruments. Repeatability studies on the gaseous emissions sampling have shown uncertainties of 5% in NOx and 10% intHC and CO, including both instrumentation uncertainty and variations in engine operating condition. Particulate matter was measured using a micro-dilution system, where a fraction of the exhaust stream was separated and diluted at a factor of 15:1. The particulate loading in this diluted sample was then measured either using a tapered element oscillating microbalance ("TEOM"). Rupprecht & Pataschnick Model 1105, or with gravimetric filters. Pallflex Emfr'i' filters were used to collect the samples, and were then weighed (accuracy ± 5 |ig) to caiculate the mass concentration m the exhaust stream. TEOM results were found to be, on average, 8% below the gravimetric filter readings (correlation coefTicienl 0.96). Where TEOM results are used in this work, they are identified by the caption "TEOM PM". [0044] Due to the single cylinder engine's high internal friction, brake-performance parameters are not representative of the in-cylinder conditions. As a result, the engine operation was measured on the basis of the gross-indicated power - the integral of the in-cylinder pressure versus volume curve, over the compression and power strokes only, as defined in J.B. Heywood in "Internal Combustion Engine Fundamentals, published in 1988 by McGraw-Hill, New York, The gross-indicated power, normalized by engine speed and displaced volume, provided the gross-indicated mean effective pressure (GIMEP). The indicated power was used to normalize both fuel consumption and emissions measurements. The gross-indicated specific fuel consumption (GISFC) reported the total fuel mass flow, with the gaseous component represented as-an equivalent mass of dies J on an energy basis (lower heating values: diesel, 42.8MJ/kg; NG, 48.8MJ/kg; 10%H2, 50.6MJ/kg; 23%H2, 52,5MJ/kg), (0045) The iji-cylindcr presswe trace can also be used to estimate the net heat-release rate, as given by: where is the crank angle, pis the in-cylinder pressure at a given crank angle, Kis the cylinder volume at that point, and yis the specific heat ratio {Cp/Cv - assumed constant). The net heat release rate represents the rate of energy release from the combustion processes less wall heat transfer and crevice flow losses. By integrating the heat-release rate up to a given crank-a gle and normalizing by the total energy released over the fijll cycle, the fraction of the energy released up to that point in the cycle can be determined. The midpoint of this curve is 50% of the integrated heat release (50% IHR), and can be used to define the combustion timing. 10046| The engine operation was also defined on the basis of the equivalence ratio {fy-ratio of actual to stoichiometric fiiel/oxidizer ratio). The amount of dilution of the intake air is defined by the intake oxygen mass fraction (Yi„io2) (0047! The experimental test conditions selected for testing the gaseous fuel mixture comprising methane and hydrogen were based on a desire to reduce fuel consumption while increasing operating condition realism. Specifically, an operating condition with high emissions associated with natural gas operation was of interest, to determine how effectively hydrogen could enhance poor natural gas combustion. The selected operating condition had the following characteristics: a high EGR fraction, namely 40% by mass; an intake oxygen mass fraction (Yu,io2)of0.175; an engine speed of 800 RPM; a low load, namely 6 bars gross indicated mean effective pressure ("GIMEP"); and, a moderate ij) of 0.5 (oxygen-based). Experiments were conducted with a fuel injection pressure of 16 MPa and 20 MPa. Natural gas with a 94% methane concentration by volume was the source of methane for the gaseous fiiel mixture, and mixtures with 10% hydrogen and 23% hydrogen by volume were tested. To establish influences over a range of conditions while minimizing the required changes to the operating condition, a range of combustion timings were used. By varying combustion timing, highly stable conditions (early timings) and very unstable conditions (late timings) could be tested at the same baseline (EGR, load, speed) condition. To improve experimental precision, it was decided to use a paired-testing approach, where a single point was tested using first natural gas and then the gaseous fuel mixture (or in the opposite order). By fixing the operating condition, then varying the timing, it was possible to minimize variations due to non-repeatability of the operating condition setpoint. Repiication of timing sets was used to establish repeatability. Most of the testing was carried out with a fuel injection pressure of 16\fPa, to ensure that the commanded injection opening durations were repeatable (in excess of 0.9 ms). As this pressure is below the pressures typically used in other gaseous fuelled direct injection internal combustion engines, such as engines that are fuelled with 100% natural gas, a set of tests, with both natural gas and the gaseous fijel mixtures, were carried out at 20 MPa to ensure that the trends were not being influenced by this parameter. [0048] The effects of mixing 10% and 23% (by volume) hydrogen in methane oa emissions are shown in Figure 3. Compared to the data from the same engine fiielled with natural gas alone, the data from the tests using a gaseous fuel mixture comprising 10% hydrogen showed that for the injectio" timings tested, the measured emissions were either the substantially the same or reduced. For example, the measured data indicated that die emissions of PM, tHC and CO were reduced on the order of 5% to 10%. Furthermore, it is noteworthy that there were no detrimental effects to the engine operation or the measured emission levels, resulting from the addition of hydrogen into the fuel. That is, the addition of hydrogen had no significant effect on the emissions of NOx. |l)049| It should be noted that the error bars presented in the plotted data are based on Ihe long-term uncertainty estimates, including both analyzer sensitivity and variations in engine operating condition. PM errors are based on calculated uncertainty for the gravimetric samples. [y050| The addition of23Vo hydrogen had a greater impact on the emissions than did 10% hydrogen. NOx emissions were increased slightly but were substantially unchanged, while CO, tHC, and COj (not shown) emissions were reduced. Due to uncertainties in the PM measurements, the only observed significant influence was at the latest timings, where a substantial reductit . in PM was observed with 23%H2 compared to the same timings for the engine fuelled with 100% natural gas or a gaseous fuel mixhirewith lQ%hydrogen. The presence ofhydrogen in die combustion zone may have affected pollutant emissions due to an increased concentration of the OH radical. This highly reactive molecuie would provide more rapid oxidation of unbumed fuel and partial-combustion species such as CO and tHC. Hydrogen has also been shown to effectively reduce local flame extinctions induced by high turbulent strain-rates, events that are thought to generate substantial pollutant emissions. That NOx emissions were slightly increased by hydrogen addition is possibly due to an increase in the prompt-NO mechanism resulting from higher OH concentrations. It may also be due to the more intense combustion with the hydrogen addition. [0051] The low. levels of PM being measur i were near the detection limit of the instruments. However, the results shown in Figure 3 show that even for the gaseous fUel mixture with only 10% hydrogen a smalt reduction in PM was consistently observed. For the gaseous fuel mixture with 23% hydrogen, for earlier injection timings a similar small reduction in PM was observed, but as the injection timing was delayed, more significant reductions in PM emissions were achieved. This is a significant difference in PM emissions from what normally occurs and that is expected from engines fiielled with 100% natural gas when later injection timings are tested. These results show (hat, unlike an engine fijcUed with only methane or natural gas, by using a gaseous fuel nuxture comprising methane and at least 23% hydrogen, for a low-load, low-speed engine condition it is possible to delay the timing for fuel injection to achieve significant reductions in NOx emissions without the noimal consequence of significantly increasing the emissions of PM. [00521 The effects of 10% and 23% hydrogen mixed with natural gas are compared to the natural gas fuelling case in terms ofburn duration (10-90% of integrated heat release), gross indicated specific fuel consumption (GISFC), peak heat-release rate, and coefficient of variation (GOV) of the GIMEP in Figure 4, The GISFC showed no significant influence of either timing or fliel composition. The bum duration was substantially reduced for the hydrogen-fuelling cases at late timing, especially with 23% hydrogen. Interestingly, there was no change in bum duration for the earlier timings. This suggests that different mechanisms may restrict the combustion rate at early and late timings, with a chemical kinetic limit at late timings, compared to a mixing-limited condition for early timings. The peak heat-release rate {corresponding roughly to the maximum rate of chemical energy being released from the (uel) averaged approximately 20% higher for the engine when Juelied with the gaseous fuel mixture comprising 23% hydrogen by volume, compared to when the engine was fuelled with 100% natural gas. The difference when the engine was fuelled with a gaseous fuel mixture comprising only 10% hydrogen" was less significant, although there was a slight increase in peak heat release rate (HRR) at most timings. The use of a gaseous fuel mixture comprising hydrogen and methane also substantially reduced the combustion variability (as measured by the COV GIMEP). For the gaseous fiiel mixture that comprised 10% hydrogen, a siificant reduction in variability was observed at the later combustion timing. For the gaseous fuel mixture that comprised 23% hydrogen, reduced variability was seen at all combustion timings, althou the reduction in variability was greatest for later combustion timings. This reduction in combustion variability can be due to increased flame stability caused by the addition of hydrogen, which can contribute directly to the observed reduction m CO and tHC emissions. |0 the point at which a rapid further increase in heat-rclease rate was observed. Examples of these locations are shown in the heat-release plot in Figure 6. J0054J The observed shorter gas ignition delay time is consistent with premixed and iion-premixed auto-ignJtion of methane tests, previously reported in 1997 by C.G. Fotache, T.G. Kreutz and C.K. Law in, "Ignition of Hydrogen-Enriched Methane by Heated Air", published in Combustion and Fiame, Vol 110, Pp. 429-440, which showed that hydrogen addition couid substantially reduce ignition delay times. However, the work of Fotache etal. does not relate to a non-premixed jet being ignited by a pilot flame, and therefore is not directly comparable to the presently disclosed method and apparatus. Contrary to the work of Fotache el al, that suggested that even at 10% H;, a noticeable reduction in ignition delay occurred, the experimental data shown in Figure 5 indicates that for the subject internal combustion engine, which employed pilot fuel to assist with ignition of the directly injected main fuel, a more substantial quantity of hydrogen was required before a significant effect was detected. Because the combustion process is complex, the shorter gas ignition delay can have a number of effects on the combustion process. First, the time available formixing is substantially reduced. While hydrogen can mix somewhat faster, due to its higher difSisivity, the methane diffusion rate is essentially constant. This can lead to less methane being over-mixed during the ignition delay period, resulting in a reduction in tHC emissions. The shorter ignition delay can also result in less air mixing into the gaseous jet during the pre-combustion period, resulting in a richer jet core during the combustion process. This richer jet can result in an increase in soot formation. The reduction in PM (which is not as substantial as the reductions in CO and tHC) may be a resuU of increases in both the soot formation (caused by the richer non-premixed jet) and oxidation through the OH radical processes. 100551 Figure 6 shows that for an engine fuelled with a gaseous fbel comprising 10% hydrogen and 90% natural gas, there was no significant difference observed in the in-cylinder conditions, as represented by the pressure trace and heat-release rate compared to when the engine was fuelled with 100% atura1 gas. hi this example, the timing shown by 601 is when the pilot fuel injection begins, while 602 shows the liming for when the injection of the gaseous fuel mixture begins. The first increase in net heat release rate at the timing shown by 603 indicates the start of combustion for the pilot fuel and the second increase in the net heat release rate shown by 604 indicates ihc timing for start of combustion for the gaseous fuel mixture. While the pressure traces and heat-release rates for the 0, 5, and 15°ATDC timings are not shown, similar results, were observed at these other timings. When the engine was fuelled with a ga.seous fuel mixture comprising 23% hydrogen and 77% natural gas a more significant effect on the in-cyiinder conditions was observed. For the data plotted in Figure 7, to maintain iJie same combustion timing for both of the plotted fijelling conditions (100% natural gas and a gaseous fuel mixture comprising 23% hydrogen and 77% natural gas), to compensate for the shorter ition delay the timing for injecting the gaseous fuel mixture was delayed by about 4 crank angle degrees. Figure 7 shows that for an engiiie fuelled with a gaseous fiiel mixture comprising hydrogen and natural gas, the heat release rate changes as a function of both fuel composition and fuel injection timmg. That is, the peak heat-release rate was substantially higher at all the combustion timings when the engine was fuelled with a gaseous fuel mixture comprising hydrogen, with peak heat-release rate increasing with increasing proportions of hydrogen in the fuel mixture, ITie effect of peak heat-release rates being higher for engines flielled with fuel mixtures comprising hydrogen was relatively consistent, although the increase in heal release rate is more substantial at 15'ATDC than at the earlier timings. The effect of fuel injection liming was observed to be consistent for both natural gas and gaseous fuel mixtures of hydrogen and natural gas, in that retarding injection timing resulted in reductions in the heat release rale. |0056| For the bulk of the testing, the mid-ooint of the heat release (50%IHR) was hdd constant by varying the start-of-mjection timing (both pilot and main fuel timings shifted equivalently, as the relative delay between the gas and diesel injections was held constant). While ttiis technique maximized comparability of the combustion timing, it resulted in variations in the combustion timing. To study this, experiments were conducted to collect two sets of data. One set of data was collected from the engine when it was operated with the same start-of-injection timing (pilot and gas) as for when the engine was fuelled with 100% natural gas, except that the engine was fuelled with a gaseous fiiel mixture comprising 23% hydrogen and 77% natural gas. A second set of data was collected with the same fuelling condition but with adjustments to the timing for start-of-injection to maintain a constant combustion liming for the mid-point of the integrated heat release. |0057| The effects of these timing adjustments on the in-cyiinder performance are ihown in Figure 8, which shows the in-cyHnder pressure and heat-release rate for the Allowing three conditions; (1) 100% natural gas; (2) a gaseous fuel mixture comprising 13% hydrogen and 77% natural gas, using the same timing as for IO0% natural gas; and, 3) a gaseous fuel mixture comprising 23% hydrogen and 77% natural gas, but with the iming for start-of-injection adjusted to maintain the same timing for the mid-point of the ntegrated heat release (50%EHR) as for 100% natural gas. This data is for the condition vhere the 50%IHR was set to 10 crank angle degrees after top dead center ("ATDC), for he engine fuelled with 100% natural gas and for the engine fuelled with 23% hydrogen vith the adjusted timing for start-of-injection. The addition of hydrogen to the fuel lubstantially reduced the ignition delay time, as shown by the significantly earlier nain combustion event, wfiile the pilot start-of-combustion (shown by the first increase )n the heat-release plot) was substantially constant for all three conditions. Similar esults were seen at all timings for the mid-point of the integrated heat release, as shown n Figure 9. A shorter gas ignition delay w s observed for the engine when it was fuelled vith a gaseous fiiel mixture comprising hydrogen under both fixed and adjusted timings, t is thought that the gas ignition delay was shorter for the fixed timing condition because he ignition was occurring earlier in the cycle. The mid-point for the integrated heal elease was advanced by approximately 4 crank angle degrees ("CA) for all the constant njection timing cases. The effects on emissions {not shown) were consistent with the ;ffects of advancing the timing by approximately ACA. 0058| ITie injection pressure of 16 MPa that was used to collect most of the experimental data is lower than what is normally used for gaseous-fiielled engines that lirectly inject gaseous fuels such as natural gas into the combustion chamber of an ntem combustion engine. Generally, higher injection pressures are considered to be nore desirable and injection pressures between 19 MPa and 30 MPa are more typical, To lest the effect of injection pressure on the observed results, some of the experiments vere repeated wilh a fuel injection pressure at 20 MPa. While this was still substantially lelow the highest achievable injection pressures, it provided a reasonable injection rail / peak cylinder pressure ratio, due to the low in-cy!inder pressure. The minimum liiel / cylinder pTessure ratio at the earliest combusticpn timing (where the peak cylinder pressurewas highest) was 2:1 at 20 Mpa, compared to 1,6:1 for the 16 MPa injection. For later combuMion timings, the ratio was increased to as much as 3.3 (compared to 2.7 for the 16 MPa case). These ratios do not represent the actual ratio between the ftie! at the injector nozzle and the in-cylinder condition, as the cylindCT pressure changed over the injection period, while the pressure of the gas exiting the nozzle was substantially lower than the rail pressure due to flow losses within the injector body and gas dynamics at the nozzle outlet. However, these ratios do provide a means for characterizing the effect of injection pressiae and provide a basis for comparing such effects between engines Urat are ftielled with 100% natural -las. and engines that are fiielled with a gaseous fuel mixture comprising hydrogen and methane. |0I)59] The effect of increasing the injection pressure on emissions is shown in Figure 10. The higher injection pressure tended to increase CO, PM, and tHC emissions, while l\Oxand GISFC were not affected. The results can be seen to be consistent for both natural gas and hydrogen-methane blend fiielling. That the injection pressure had little impact on the in-cylinder perforaiance is shown in Figure II, which plots a pressure trace and the heat release rate for 16 and 20 MPa, for an engine fiielled with a gaseous fiiei mixture comprising 23% hydrogen and 77% natural gas by volume. It was surprising that the higher injection pressure resulted in slightly increased levels of PM, tHC, and CO emissions compared to engine fiielled widi the same gaseous tuel mixture but with lower injection pressures. However, the experimental results do show that hydrogen addition resulted in reductions in the emissions of PM, total hydrocarbons (tHC) and carbon monoxide (CO), without any negative impact on emissions of NOx, and that this result was generally consistent at both injection pressures. Accordingly, these results indicate that hydrogen additic . has a positive impact on emissions over a range of fuel injection pressures. (0060] From the experimental data collected it is possible to determine certain trends relating to engine emissions and combustion stability arising from fuel composition and combustion timing. That is, these trends can be extrapolated from the data that was collected when the engine was operated with gaseous file! mixtures comprising 100% natural gas (and 0% hydrogen), 90% natural gas and 10% hydrogen, and 77% natural gas and 23% hydrogen. When the engine was fuelled with 90% natural gas and 10% hydrogen, improvemenls were observed in combustion stability and engine emissions were substantially the same or slightly reduced compared to when the same engine was fijelied with 100% natural gas. When the same engine was fuelled with 77% natural gas and 23% hydrogen, there were greater improvements in combustion stability and more substantial improvements in engjne emissions. Although the results are not plotted in the figures, experiments WCTe also conducted in which the engine was fueled with up to 35% hydrogen by volume (at STP), and at such higher hydrogen percentages the effect on emissions continued to be beneficial. However, hydrogen is harder to compress compared to natural gas and the higher volume occupied by hydrogen compared to meane for the same amount of energy introduces volumetric flow capacity challenges for gaseous fiiei mixtures with higher percentages of hydrogen. From the experimental data collected, the levels of emissions observed fix)ra the conducted experiments, and the pre-existing knowledge be relating to the combustion of gaseous fiiel mixtures in other engines, it can be reasonably determined that, compared to an engine flieiied with 100% natural gas, improved combustion stability and improved engine emissions can be achieved with gaseous fuel mixtures comprising hydrogen in concentrations from 5% to at least 60% by volume. From the observed trends plotted in Figure 3 and 4, higher hydrogen concentrations can yield better combustion stability (reduced combustion variability) and lower emissions, but these advantages can be offset by other factors such as higher hydrogen percentages requiring increased volumetric flow requirements, or the cost and availability of hydrogen. For higher percentages of hydrogen, the properties of the gaseous fiiel mixture can also change because hydrogen has a lower lubricity compared to natural gas. In some cases, the preferred gaseous fijel mixnjre can be between 10% and 50% hydrogen or an even narrower ranges, such as between 15% and 40% hydrogen mixed with natural gas or between 20% and 35% hydrogen mixed with natural gas. By way of specific examples, &ie gaseous fuel mixture can comprise methane and hydrogen with hydrogen content expressed as a percentage by volume being one of 12%, 14%, 16%, !8%, 20%, 22%, 23%, 24%, 25%, 26%, 28%. 30%, 32%, 34%, 35%, 36%, 38%, 40%, 42%. 44%, 46%, 48%, 50% and any percentage therebetween. [00611 From the experimental data, trends can also be determined relating to combustion timing. For an engine fiielled with a gaseoiis fuel mixture comprising hydrogen and methane, combustion stability can be achieved over a broader range compared to the same engine fuelled with 100% natural gas. For an engine fuelled with a gaseous fliei mixture comprising 10% hydrogen, this improved stabihty was observed to occur when timing for the mid-point of the integrated heat release occurred 10 crank angle degrees after lop dead center and later. For the same engine fuelled with a gaseous fuel mixture comprising 23% hydrogen, an improvement in combustion stability was observed as early as when the mid-point of the integrated heat release occurred 5 crank angle degrees after top dead center, with improvements to combustion stability increasing fiirther still for later combustion timings. From the experimental data it can be corctuded that an engine fiielled with a .aseous fuel mixture comprising methane and at least 10% hydrogen by volume, can equal or better the combustion stability and emissions from the same engine fuelled with 100% natural gas. Even though most of the data was collected for one engine operating condition, since the selected engine operating condition was one that is normally associated with hi engine emissions it is expected that the tested gaseous f\ie\ mixtures comprising at least 10% hydrogen and a majority of methane by volume will produce similar or better emissions and combustion stabihty compared to the same engine fuelled with 100% natural gas, when the engine is operated at ditlerent engine conditions. [0062] In summary, the experimental results show that an internal combustion engine with direct injection of a gaseous fuel mixture comprising hydrogen and methane can be operated to reduce emissions and improve combustion stability compared to the same enne fuelled with i00% natural gas. The graph of coelScient of variation of the GIMEP against combustion timing in Figure 4 shows Uiat the addition of hydrogen results in a substantial reduction in the combustion variability. The experimental results also show that -while hydrogen addition ca.. increase the peak combustion beat release rate, indicatmg higher combustion temperatures, the addition of hydrogen did not result in mcreascd levels of NOx emissions compared to when the engine was operated under the same conditions but tueiled with 100% natural gas. The results further show that hydrogen addition can allow later combustion timings because the level of PM emissions at later combustion timings are reduced compared to when the engine was fuelled with 100% natural gas. The experimental data confirmed that like engines fuelled with 100% natural gas, the levels of NOx emissions decrease with later combustion timings for engines fuelled with gaseous mixtures comprising hydrogen and methane. 'Whereas with engines fuelled with 100% natural gas, the steep increase in PM emissions for later combustion timings establishes a limit to h w much combustion timing can be retarded, the experimental results show that for engines fuelled with a gaseous fuel mixture comprising hydrogen and natural gas, later combustion timings are possible because PM emissions increase at a much shallower slope as combustion timing is delayed. In addition, it was found that a characteristic of gaseous fuel mixtures comprising hydrogen and methane that were directly injected into a combustion chamber of an internal combustion engine was that the gaseous fije! mixtures ignited with a shorter ignition delay compared to that of natural gas without the addition of hydrogen. For the tested engine condition the shorter ignition delay results in the combustion timing being advanced about 4 crank angle degrees, which resulted in higher peak in-cylinder pressures and higher peak heat release rates if the same injection liming used for a natural gas engine was maintained. It was determined that timing adjustments can be made so that combustion characteristics match those of engines fuelled with 100% natural gas. (0063] While particular elements, embodiments and applications of the present invention have been shown and described, it will be understood, that the invention is not limited thereto since modifications can be made by those skilled in the art without departing from the scope of the present disclosure, particularly in light of the foregoing teachings. WE CLAIMS 1. A method is provided of operating a direct injection internal combustion engine, said method comprising introducing a gaseous fuel mixture directly into a combustion chamber of said engine, wherein said gaseous fuel mixture comprises methane and between 5% and 60% hydrogen by volume at standard 5 temperature and pressure, and for at least one engine operatmg condition, maintaining a fuel rail to peak in-cylinder pressure ratio of at least 1.5:1 when introducing the gaseous fiiel mixture into said combustion chamber. 2. The method of claim 1 wherein said gaseous fuiel mixture comprises between 10% and 50% hydrogen by volume at standard temperature and pressure. 3. The method of claim 1 wherein said gaseous fuel mixture comprises between 15% and 40% hydrogen by volume at standard tempierature and pressure. 4. .The method of claim 1 WR rein said gaseous fuel mixture comprises between 20% and 35% hydrogen by volume at standard temperature and pressure. 5. Themcthodof claim I wherein said methane is a constituent part of natural gas. 6. The method of claim 1 further comprising premixing said gaseous fuel mixture and storing it as a blended fuel within a storage tank from which it can be delivered to said engine. 7. The method of claim 1 further comprising controlling fiicl injection timing so that the mid-point of integrated combustion heat release occurs between 2 and 30 craidc angle degrees after top dead center. S, The method of claim I further comprising controlling fuel injection liming so that in at least one engine operr Ing condition the mid-point of integrated combustion heat release occurs between 5 and 15 crank angle degrees after top dead center. 9. The method ofclaim I further comprising injecting a pilot fuel directly into said combustion chamber about I millisecond before start of injection of said gaseous fuel mixture. 10. The method ofclaim 9 wherein said pilot fuel is a liquid fuel with a cetane number between 40 and 70, 11. The method ofclaim i 0 wherein said liquid fuel is diesel fuel, 12. The method ofclaim 9 wherein over an engine operating map said pilot fuel is on average between 3% and 10% of the fuel that is consumed by said engine on an energy basis, 13. The method ofclaim 9 wherein over an engine operating map said pilot fuel is on average between 4% and 6% of the fuel that is consumed by said engine on an energy basis. 14- The method of claim 1 further comprising heating a hot surface inside said combustion chamber to assist with igniting said gaseous fuel mixture. 15, The method of claim 14 wherein said hot surface is provided by a glow plug and said method further comprises electrically heating said glow plug. 16, The method of claim 1 further comprising spark igniting said gaseous fuel mixmre inside said combustion chambor. 17, The method ofclaim I further comprising storing said hydrogen separately from said methane and mixing said hydrogen and methane to form said gaseous fuel mixture. 18. The method of claim 1 further comprising controlhng the proportions of hydiogen and methane in said gaseous iiiel mixture as a function of engine operating conditions. 19. The method of claim ! further comprising mamtaining a fuel rail to peak in-cylinder pressure ratio of at least 1.5:1 when introducing said gaseous fuel mixture into said combustion chamber for all engine operating conditions. 20. The method of claim I further comprising maintaining a fuel rail to peak in-cylinder pressure ratio of at least 2:1 when introducing said gaseou-s fuel mixture into said combustion chamber for at least one engine operating condition. 21. The method of claim 1 fiirther comprising maintaining a choked flow condition at a nozzle orifice of a fuel injection valve when introducing said gaseous fiiel mixture into said combustion chamber. 22- The method of claim I further comprising injecting said gaseous fuel mixture into said combustion chamber with an injection pressure that is at least i 6 MPa (about 2350 psia). 23. The method of claim I further comprising injecting said gaseous fijel mixmre mto said combustion chamber with an injection pressure that is at least 20 MPa (about 2900 psia). 24. The method of claim 1 wherein in the course of a compression stroke, an intake charge inside said combustion chamber is compressed by a ratio of at least about 14:1. 25. The method of claim 1 wherein methane is the largest constituent of said gaseous fuel mixture by volume at standard temperature and pressure. 26. A method is provided of fuelling an internal combustion engine, said method comprising: introducing a gaseous fuel mixture directly into a combustion chamber of said engine, wherein said gaseous fuel mixture comprises methane, and 5 introducing hydrogen into said combustion chamber, thereby adding hydrogen to said gaseous fuel mixture, wherein said hydrogen represents between 5% and 60Vo by volume of said gaseous fuel mixture at standard temperature and pressure; and maintaining a gaseous fuel mixture rail to peak in-cylinder pressure ratio of at least 1.5;1 when introducing the gaseous fuelmixture into said combustion chamber 10 for at least one engine operating condition. 27. The method of claim 26 further comprising premixing said hydrogen with said gaseous fiiel mixture comprising methane, and introducing said gaseous fuel mixture and said hydrogen directly into said combustion chamber. 28. The method of claim 26 further comprising premixing said hydrogen with intake air and introducing said hydrogen into said combustion chamber during an intake stroke of said piston, 29. The method of claim 26 further comprising introducing said hydrogen direcdy into said combustion chamber separately from said gaseous fuel mixture. 30. An internal combustion engine is provided that can be fuelled with a gaseous fuel mixture comprising methane and between 5% and 60% hydrogen by volume at standard temperature and pressure, said engine comprising: a combustion chamber defined by a cylinder, a cylinder head, and a 5 piston movable within said cylinder; a fuel injection valve with a nozzle that is disposed within said combustion chamber, said fuel injection valve being operable to mtroduce said gaseous fuel mixture directly into said combustion chamber; a pressurizing device and piping for delivering said gaseous fuel 10 mixture to said injection valve with a ratio of fuel rail to peak in-cylinder pressure being at least 1.5:1 for at least one engine operating condition; and an electronic controller in communication with an actuator for said field injection valve for controlling timing for operating said fijel injection valve, 31. The engine of claim 30 wherein said engine has a compression ratio of at least 14. 32. The engine of claim 30 wherein said electronic controller is programmable to time introduction of said gaseous f\iel mixture into said combustion chamber so that the mid-point of an integrated combustion heat release occurs 5 between 2 and 30 crank angle degrees after top dead center. 33. The engine of claim 30 wherein said electronic controiier is progranunable to time introduction of said gaseous fuel mixture into said combustion chamber so that the mid-point of an integrated combustion heat release occurs between 5 and 15 crank angle degrees after top dead center. 34. The engine of claim 30 wherein said fuel injection valve is mounted in said cylinder head with said fuel injection valve comprising a nozzle disposed withm said combustion chamber. 35. The engine of claim 30 further comprising a second fiiel injection valve that is operable to introduce a pilot fuel directly into said combustion chamber. 36. The engine of claim 35 wherein said second fuel injection valve is integrated into a valve assembiy that also comprises said fuel! injection valve for introducing said gaseous fuel mixture. 37. The engine of claim 36 wherein said second fuel injection valve and said fiiel injection valve for introducing said gaseous fiiel mixture can be independently actuated and said gaseous fuel mixture is injectable into said combustion chamber through a first set oi nozzle orifices, which are different from a 5 second set of nozzle orifices through which said pilot fuel is injectable into said combustion chamber. 38. The engine of claim 30 fiirther comprising an ignition plug disposed within said combustion chamber that is operable to assist with ignition of the gaseous fuel mixture. 39. The engine of claim 38 wherein said ignition plug is a glow plug that is electrically beatable to provide a hot surface for assisting with ignition of said gaseous fiiel mixture. 40. The engine of claim 38 wherein said ignition plug is a spark plug. 41. The engine ofclaim 30 further comprising a storage vessel forstonng said gaseous fuel mixture in a substantially homogeneous mixture with predetermined proportions of hydrogen and methane, 42. The engine of claim 30 further comprising a first storage vessel within which said hydrogen can be stored, a second storage vessel within which a gaseous fuel comprising methane can be stored, and valves associated with each one of said first and second storage vessel that are operable to control respective proportions of 5 hydrogen and methane in said gaseous fiiel mixture that is introducible into said combustion chamber, 43. The engine of claim 42 wherein said electronic controller is progiammable to change respective proportions of hydrogen and methane in said gaseous fuel mixture to predetermined amounts responsive to detected engine operating conditions. 44. An internal combustion engine is provided that can be fuelled with a gaseous fuel mixture comprising methane and hydrogen, said engine comprising: , a combustion chamber defined by a cylinder, a cylinder head, and a piston movable within said cylinder; 5 a first fuel injection valve with a nozzle disposed within said combustion chamber, wherein said fuel injection valve is operable to introduce methane directly into said combustion chamber; a second fuel injection valve with a nozzle disposed within an intake air manifold, wherein said second fuel injection valve is operable to introduce 10 hydrogen into said intake air man-fold from which said hydrogen can flow into said combustion chamber; and an electronic controller in communication with an actuator for each one of said first and second fuel injection valves for controlling respective timing for operating said first and second fuel injection valves. 45. The engine of claim 44 further comprising a pressurizing device and piping for delivering said methane to said first injection valve with a ratio of fuel rail pressure to peak in-cylinder pressure being at least 1.5:1 for at lea.st one engine operating condition. 5 46. The engine of claim 44 wherein said engine has a compression ratio of at least 14.

Documents:

5245-CHENP-2008 AMENDED CLAIMS 23-05-2014,,.pdf

5245-CHENP-2008 CORRESPONDENCE OTHERS 03-02-2014.pdf

5245-chenp-2008 correspondence others 27-12-2010.pdf

5245-CHENP-2008 EXAMINATION REPORT REPLY RECEIVED, 23-05-2014.pdf

5245-chenp-2008 form-3 27-12-2010.pdf

5245-CHENP-2008 FORM-3 03-02-2014.pdf

5245-CHENP-2008 FORM-3 07-05-2012.pdf

5245-CHENP-2008 POWR OF ATTORNEY 23-05-2014.pdf

5245-CHENP-2008 CORRESPONDENCE OTHERS 07-05-2012.pdf

5245-CHENP-2008 EXAMINATION REPORT REPLY RECEIVED 03-02-2014.pdf

5245-chenp-2008 abstract.pdf

5245-chenp-2008 assignment.pdf

5245-chenp-2008 claims.pdf

5245-chenp-2008 correspondence-others.pdf

5245-chenp-2008 description(complete).pdf

5245-chenp-2008 drawings.pdf

5245-chenp-2008 form-1.pdf

5245-chenp-2008 form-3.pdf

5245-chenp-2008 form-5.pdf

5245-chenp-2008 pct.pdf