GB2623804A

GB2623804A - An engine

Info

Publication number: GB2623804A
Application number: GB2215955.2A
Authority: GB
Inventors: Mccarthy Paul; Vowels Richard; Vickery Warren
Original assignee: Jcb Research
Current assignee: Jcb Research
Priority date: 2022-10-27
Filing date: 2022-10-27
Publication date: 2024-05-01
Also published as: GB202215955D0; WO2024089440A1

Abstract

A four-stroke gaseous fuel engine comprising one or more cylinder assemblies comprising: a cylinder including an inlet and an outlet having intake valve and an exhaust valves; and a fuel injector. Gaseous fuel (such as hydrogen) is injected from the fuel injector into an intake runner to supply fuel and air from the intake runner to the cylinder via the inlet during an intake stroke of the piston. The fuel is injected into the intake runner exclusively between first and second crank-angles of the piston, the first crank angle A1 corresponds to the intake valve being closed and the exhaust valve being open during the exhaust stroke, and the second crank angle A2 corresponds to the intake valve being open and the exhaust valve being closed during the intake stroke. The assembly allows for injecting gaseous fuel at a flow rate high enough to ensure that the required volume of fuel is supplied within each four-stroke cycle.

Description

AN ENGINE

FIELD

The present teachings relate to an engine, particularly a four-stroke gaseous fuel engine, a working machine, and a method for operating a four-stroke gaseous fuel engine.

BACKGROUND

Four-stroke gaseous fuel internal combustion (IC) engines are known in which a gaseous fuel, such as compressed natural gas (CNG) for example, is supplied to a cylinder of the engine via port fuel injection, in which the gaseous fuel is injected into an air supply passage upstream of the cylinder. This is in contrast to direct fuel injection, in which fuel is injected directly into the cylinder.

In such IC engines, a sufficient volume of the gaseous fuel is required to be injected into the air supply passage within each four-stroke cycle of the engine such that near-optimal combustion is achieved in the cylinder.

Hitherto, IC engines used in off-highway, and heavy commercial vehicle applications have typically used diesel as fuel, which is directly injected into the combustion chamber. To increase the efficiency of diesel and gasoline engines, it is known for intake valves to open at the end of the exhaust stroke before the piston reaches top dead centre (TDC) as it reduces the pumping losses of the engine. This potentially allows exhaust gases to flow into the air supply passage during the exhaust stroke within the cylinder causing gaseous fuel and air contained in the air intake passage to flow upstream away from the cylinder. If utilising hydrogen, for example, as an alternative fuel, and injecting the fuel into the inlet port there is an increased risk of the backflowing air and fuel in the inlet runner combusting outside the cylinder ("backfiring") or flowing into other cylinders of the engine, which may result in poor combustion and/or misfiring of those cylinders. It may also be costly to source fuel injectors capable of injecting gaseous fuel at a flow rate high enough to ensure that the required volume of gaseous fuel is supplied to the cylinder within each four-stroke cycle.

The present teachings seek to overcome or at least mitigate one or more problems

associated with the prior art.

SUMMARY

According to a first aspect of the present teachings, there is provided a four-stroke gaseous fuel engine comprising one or more cylinder assemblies, each cylinder assembly comprising: a cylinder including an inlet and an outlet selectively opened and closed by an intake valve and an exhaust valve respectively; a piston translationally movable within the cylinder; an intake runner leading to the inlet; and a fuel injector, wherein the cylinder assembly is configured to selectively inject a gaseous fuel from the fuel injector into the intake runner, wherein the engine is configured to supply gaseous fuel and air from the intake runner to the cylinder via the inlet during an intake stroke of the piston, and exhaust combustion gases from the cylinder via the outlet during an exhaust stroke of the piston, wherein, within a four-stroke cycle of the engine, the cylinder assembly is configured to inject the gaseous fuel into the intake runner exclusively during an interval starting with a first crank angle of the piston and ending with a second crank angle of the piston, and wherein at a maximum interval that the cylinder assembly can inject the gaseous fuel into the intake runner, the first crank angle corresponds to the intake valve being closed and the exhaust valve being open during the exhaust stroke, and the second crank angle corresponds to the intake valve being open and the exhaust valve being closed during the intake stroke.

Advantageously, injecting gaseous fuel into the intake runner over such a maximum interval may help ensure a sufficient volume of gaseous fuel is supplied to the cylinder for near-optimal combustion, whilst also enhancing fuel-air mixing, which may promote enhanced combustion.

A maximum second crank angle at which the cylinder assembly can inject the gaseous fuel into the intake runner may be a crank angle of 495 degrees or less after a top power dead centre of the piston.

The maximum second crank angle at which the cylinder assembly can inject the gaseous fuel into the intake runner may be a crank angle of 490 degrees or less after a top power dead centre of the piston.

Advantageously, it has been found that such maximum second crank angles help to inhibit backflow of gaseous fuel and air within the intake runner away from the inlet, which may cause poor combustion and misfiring of the engine having a plurality of the cylinder assemblies.

A minimum second crank angle at which the cylinder assembly can inject the gaseous fuel into the intake runner may be a crank angle of 440 degrees or more, optionally 450 degrees or more, after a top power dead centre of the piston.

The gaseous fuel may be hydrogen.

Advantageously, hydrogen produces less harmful emissions when combusted relative to hydrocarbon gaseous fuels for example.

The maximum interval may be in the range of 160 to 200 degrees, optionally 170 to 190 degrees, for example approximately 180 degrees.

Advantageously, such an interval may help to ensure a sufficient volume of gaseous fuel is supplied to the cylinder for near-optimal combustion.

The interval may be a function of an output torque and/or an engine speed of the engine.

The maximum interval may be at a maximum output torque at a predetermined engine speed.

The predetermined engine speed may be in the range of 1600 to 2200 RPM, optionally, in the range of 1800 to 2000 RPM.

A minimum first crank angle at which the cylinder assembly can inject the gaseous fuel into the intake runner may be at a maximum output torque at a predetermined engine speed. The predetermined engine speed may be in the range of 1500 to 2600 RPM, optionally, 1800 to 2200 RPM, for example, approximately 2000 RPM.

A maximum second crank angle at which the cylinder assembly can inject the gaseous fuel into the intake runner may be at a maximum output torque at a predetermined engine speed. The predetermined engine speed may be in the range of 1200 to 1800 RPM, optionally, 1300 to 1700 RPM, for example approximately 1500 RPM.

For a constant output torque of the engine, the interval may increase as the engine speed increases.

Advantageously, increasing the interval with increased engine speed may help ensure that a sufficient volume of gaseous fuel is supplied to the cylinder as the duration of each four-stroke working cycle of the engine reduces.

The fuel injector may be configured to inject the gaseous fuel continuously between the first and second crank angles.

Advantageously, such a configuration of the fuel injector may help to ensure a sufficient volume of gaseous fuel is supplied to the cylinder for near-optimal combustion.

The fuel injector may be configured to inject the gaseous fuel at a substantially constant injection rate between the first and second crank angles.

Advantageously, such a configuration of the fuel injector may help to simplify implementation of the engine.

Gaseous fuel may be injected into the intake runner exclusively via the fuel injector.

Advantageously, such a configuration of the engine may help to simplify and reduce the cost of implementing the engine.

Within a four-stroke cycle, the intake valve may move from a closed position to an open position at a crank angle of the piston in the range of 310 to 350 degrees, optionally, 320 to 340 degrees, for example approximately 335 degrees, after a top power dead centre of the piston.

Within a four-stroke cycle, the intake valve may move from an open position to a closed position at a crank angle of the piston in the range of 570 to 610 degrees, optionally, in the range of 580 to 600 degrees, for example approximately 585 degrees, after a top power dead centre of the piston.

Within a four-stroke cycle, the exhaust valve may move from an open position to a closed position at a crank angle of the piston in the range of 350 to 390 degrees, optionally in the range of 360 to 380 degrees, for example approximately 375 degrees, after a top power dead centre of the piston.

Within a four-stroke cycle, the exhaust valve may move from a closed position to an open position at a crank angle of the piston which may be in the range of 110 to 150 degrees, optionally, in the range of 120 to 140 degrees, for example approximately 130 degrees, after a top power dead centre of the piston.

Within a four-stroke cycle: the exhaust valve may move from a closed to an open position at a crank angle of the piston of approximately 140 degrees after a top power dead centre of the piston; the exhaust valve may move from an open to a closed position at a crank angle of the piston of approximately 380 degrees after a top power dead centre of the piston; the intake valve may move from a closed to an open position at a crank angle of the piston of approximately 335 degrees after a top power dead centre of the piston; and the intake valve may move from an open to a closed position at a crank angle of the piston of approximately 580 degrees after a top power dead centre of the piston.

Advantageously, such a configuration of the engine has been found to minimise backflow of gaseous fuel and air in the intake runner away from the inlet, resulting in an improved volumetric efficiency of the engine.

The cylinder may comprise a flat roof arranged substantially normal to an axis of piston motion. The intake valve and the exhaust valve may be arranged to open and close via movement along respective axes that are parallel to said axis of piston motion.

The cylinder assembly may be configured to inject the gaseous fuel into the intake runner from the fuel injector at an injection angle to a mean flow direction of air flowing through the intake runner and/or the horizontal, in use, in the range of 0 to 25 degrees, optionally, in the range of 10 to 20 degrees, for example 14 degrees.

Advantageously, such injection angles have been found to provide good mixing between fuel and air, and to inhibit backflow of fuel entering the intake runner via the fuel injector, in use.

Each cylinder assembly may comprise an intake port leading to the inlet. An outlet portion of the intake runner may lead to the intake port. The cylinder assembly may be configured to direct the injected gaseous fuel from the fuel injector along an injection axis in a direction substantially towards the outlet portion. The intake port may have a width along an axis of the intake valve. The injection axis may intersect the axis of the intake valve at a position greater than 50%, optionally greater than 60%, optionally greater than 70%, of the width of the intake port from the inlet.

Advantageously, such configurations of the injection port and the injection axis have been found to provide good mixing between fuel and air in the intake port, and to inhibit backflow of fuel entering the intake runner via the fuel injector, in use.

The cylinder may comprise a further inlet and a further outlet selectively opened and closed by a further intake valve and a further exhaust valve respectively. The intake runner may lead to the further inlet. The engine may be configured to supply gaseous fuel and air from the intake runner to the cylinder via the further inlet during the intake stroke of the piston, and exhaust combustion gases from the cylinder via the further outlet during the exhaust stroke of the piston.

The engine may further comprise an intake port including a downstream end leading to the inlet and the further inlet. The engine may further comprise an exhaust port including an upstream end leading from the outlet and the further outlet. The intake runner may lead to an upstream end of the intake port. The downstream end of the intake port may bifurcate to the inlet and the further inlet. The ports may be arranged in a tandem configuration.

According to a second aspect of the present teachings, there is provided a working machine comprising the engine according to the first aspect.

According to a third aspect of the present teachings, there is provided a method for operating a four-stroke gaseous fuel engine configured to supply gaseous fuel and air from an intake runner to a cylinder via an inlet of the cylinder during an intake stroke of a piston within the cylinder, and exhaust gases from the cylinder via an outlet of the cylinder during an exhaust stroke of the piston, the inlet selectively opened and closed by an intake valve, and the outlet selectively opened and closed by an outlet valve, the method comprising the step of: within a four-stroke cycle of the engine, injecting a gaseous fuel into the intake runner exclusively during an interval starting with a first crank angle of the piston and ending with a second crank angle of the piston, wherein at a maximum interval that the cylinder assembly can inject the gaseous fuel into the intake runner, the first crank angle corresponds to the intake valve being closed and the exhaust valve being open during the exhaust stroke, and the second crank angle corresponds to the intake valve being open and the exhaust valve being closed during the intake stroke.

BRIEF DESCRIPTION OF THE DRAWINGS

Embodiments are now disclosed by way of example only with reference to the drawings, in which: Figure 1A is a plan view of an internal combustion engine according to an embodiment of the present teachings; Figure 1B is a cross-sectional view of the internal combustion engine along section X-X shown in Figure 1A; Figure IC is a cross-sectional view of the internal combustion engine along section Y-Y shown in Figure 1A; Figure 2 shows plots of the movement of an intake valve and an exhaust valve, and an injection rate of a fuel injector, with respect to a crank angle of a piston, during a four-stroke cycle of the internal combustion engine of Figure 1A; Figure 3 is a contour plot of a first (lower) crank angle of a fuel injection interval with respect to an engine speed and an output torque of the internal combustion engine of Figure 1A; Figure 4 is a contour plot of a second (upper) crank angle of a fuel injection interval with respect to an engine speed and an output torque of the internal combustion engine of Figure 1A; and Figure 5 is a plan view of a cylinder, an intake port and an exhaust port of the internal combustion engine of Figure 1A.

DETAILED DESCRIPTION OF EMBODIMENT(S)

Referring firstly to Figures 1A to IC, an embodiment includes an internal combustion engine 1. Figure 1A shows a plan view of the engine 1, Figure 1B shows a cross-sectional view of the engine 1 along section X-X shown in Figure 1A, Figure 1C shows a cross-sectional view of the engine 1 along section Y-Y shown in Figure 1A.

The engine 1 is a four-stroke gaseous fuel engine configured to be powered by a gaseous fuel, such as hydrogen, compressed natural gas (CNG), landfill gas or the like.

The engine 1 may be suitable for use as the prime mover in a working machine (not shown), such as a telescopic handler, a forklift truck, a backhoe loader, a wheeled loading shovel, a dumper, an excavator or a tractor, for example. Such working machines are suitable for use in off-highway industries such as agriculture and construction. In these industries they are generally configured to perform tasks such as excavation, load handling, harvesting or planting crops. The engine may also be utilised in a genset -a self-contained unit to provide electrical power at off-grid locations. As such the engine 1 is typically required to have certain characteristics such as a high torque output over a wide engine speed band, with peak torque occurring at a relatively low engine speed.

which differ from light passenger vehicles, for example. In off-highway applications, this provides "torque backup" that enables working machines to continue to carry out working operations when encountering increased loads, or resistance to a working operation -e.g. an excavator encountering a particularly solid piece of earth to be excavated.

In this embodiment, the engine 1 has four cylinder assemblies indicated generally at 19.

As configured the engine has a maximum power output of around 55kW, although it will be appreciated that the present teachings are applicable to engines with a wide range of power outputs. Each cylinder assembly 19 includes a cylinder 5 including two inlets 6 and two outlets 9, a piston 20 translationally movable within the cylinder 5, an intake runner 16 leading to the two inlets 6, and a fuel injector 22. Each cylinder assembly 19 is configured to selectively inject a gaseous fuel from the fuel injector 22 into the intake runner 16. Each inlet 6 is selectively opened and closed by an intake valve 7i, such that there are two intake valves 7i. Each outlet 9 is selectively opened and closed by an exhaust valve 7e such that there are two exhaust valves 7e.

The engine 1 is configured to supply gaseous fuel and air from the intake runner 16 to the cylinder 5 via the inlets 6 during an intake stroke of the piston 20, and exhaust combustion gases from the cylinder 5 via the outlets 9 during an exhaust stroke of the piston 20.

In alternative embodiments (not shown), the engine 1 may have more or fewer cylinder assemblies 19, e.g. 2, 3, 6, or 8. In addition, in other embodiments the cylinders 5 may be oriented in a "V" or boxer configuration rather than inline as in the disclosed embodiment.

In alternative embodiments (not shown), each cylinder 5 may include only a single inlet 6 and/or a single outlet 9. Alternatively, each cylinder 5 may include more than two inlets 6 and/or more than two outlets 9.

Each piston 20 is configured to translate along an axis C of the respective cylinder 5.

Figures 1B and 1C show the typical orientation of the engine 1 when implemented in a vehicle such as a working machine, in use, with the axis C of the cylinders 5 substantially vertical. However, in some embodiments the cylinders 5 may be orientated at an inclined angle with respect to the vertical.

Each cylinder assembly 19 includes an intake port 4. A downstream end 4d of each intake port 4 leads to the corresponding inlets 6 of the cylinder assembly 19 (see Figure 3). An outlet portion 160 of each intake runner 16 leads to an upstream end 4u of the corresponding intake port 4 of the cylinder assembly 19.

The engine 1 comprises a cylinder block 2, a cylinder head 3, and an intake assembly 10.

The cylinder block 2 includes the cylinders 5. The cylinder head 3 includes the intake port 4 and a downstream portion 16d of the intake runner 16 of each cylinder assembly 19. The intake assembly 10 includes an upstream portion 16u of the intake runner 16 of each cylinder assembly 19. Each upstream portion 16u leads to one of the downstream portions 16d.

The cylinder head 3 is mounted to the cylinder block 2. The intake assembly 10 is mounted to the cylinder head 3. The engine 1 is configured such that the intake assembly 10 supplies a mixture of air and fuel to the intake ports 4 of the cylinder head 3 via the intake runners 16. The air-fuel mixture is then supplied from the intake ports 4 to the corresponding cylinders 5 of the cylinder block 2 via the inlets 6. As such, the engine 1 is a port fuel injection engine (i.e. fuel is provided to the cylinders 5 via port fuel injection).

The intake assembly 10 includes an intake manifold 11. The intake manifold 11 includes a first plenum 12, a second plenum 14, and the upstream portions 16u of the intake runners 16. The first plenum 12 includes an air supply inlet (not shown). The intake runners 16 are fluidly connected to, and extend from, the second plenum 14. In use, air supplied via the air supply inlet travels sequentially along the first plenum 12, the second plenum 14, and the intake runners 16 towards the inlets 6.

In alternative embodiments (not shown), air may be supplied to the intake runners 16 via any suitable arrangement.

In the illustrated embodiment, each fuel injector 22 is supplied with gaseous fuel via a fuel supply system 50 including a common fuel rail 52 in fluid communication with each fuel injector 22.

As the engine 1 utilises a gaseous fuel, e.g. hydrogen, a spark is required to initiate combustion. Thus, each cylinder assembly 19 comprises a spark plug 21 (illustrated schematically) mounted intermediate the intake and outlet ports 6, 9 in the cylinder head 3. If a hydrogen fuelled engine uses carry-over parts from a diesel engine (which does not require a spark plug) as a way of minimising costs, an advantage of port fuel injection is that the space for an injector in the cylinder head 3 is made available for a spark plug 21 instead, by repositioning the fuel injector 22 away from being directly above the corresponding piston 20.

The engine 1 includes a crankshaft (not shown) coupled to each piston 20. The engine 1 is configured such that translational movements of each piston 20 is converted into rotational movement of the crankshaft around a rotational axis thereof. Rotation of the crankshaft drives a camshaft (not shown), which moves the intake valves 7i and the exhaust valves 7e between open and closed positions. A rotation of the crankshaft by one degree is referred to as a degree of crank angle.

With reference to Figure 2, the engine 1 repeats a working cycle including four strokes of each piston 20. These four strokes are: a power stroke; an exhaust stroke; an intake stroke; and a compression stroke. The four-stroke working cycle of the engine 1 includes two complete revolutions of the crankshaft and thus 720 degrees of crank angle. Exactly one four-stroke working cycle of the engine 1 is shown in Figure 2.

Each piston 20 is translationally movable between a bottom dead centre and a top dead centre. Within each four-stroke cycle, there are two different types of top dead centre. A first of these types is the so-called top charge changing dead centre (TCDC), which occurs between the exhaust stroke and the intake stroke, and corresponds to a crank angle of 360 degrees in Figure 2. The second type is the so-called top power dead centre (TPDC), which occurs between the compression stroke and the power stroke, and corresponds to a crank angle of 0 degrees in Figure 2. Each piston 20 is at a bottom dead centre between the power stroke and the exhaust stroke at a crank angle of 180 degrees in Figure 2, and between the intake stroke and the compression stroke at a crank angle of 540 degrees in Figure 2.

Figure 2 shows exemplary plots of the movements of the intake valves 7i and the exhaust valves 7e, and the fuel injection rate of the fuel injector 22, with respect to the crank angle of the respective piston 20 during a four-stroke cycle of the engine 1 for each cylinder assembly 19.

The exemplary plots of the movements of the intake valves 7i and the exhaust valves 7e are typical for a conventional diesel engine. As such, if reconfiguring an existing diesel engine to operate with a gaseous fuel, e.g. hydrogen, or seeking to commonise parts across engines operating with different fuel, a common camshaft for controlling valve timing may be utilised for both diesel and at least hydrogen fuelled engines. Such engines may also use common engine blocks, cylinder heads, crankshafts, crankshaft connecting rods, bearings, lubrication systems, gear trains, cooling systems, flywheels, and parts of the valve train such as pushrods and rocker arms.

Figure 2 shows a first plot 100 of the amount of lift of each exhaust valve 7e away from the respective outlet 9 with respect to the crank angle of the piston 20. The outlets 9 are closed when the lift of the exhaust valves 7e away from the outlets 9 is zero, and open, at least partially, otherwise.

Figure 2 further shows a second plot 102 of the amount of lift of the intake valves 7i away from the respective inlets 6 with respect to the crank angle of the piston 20. The inlets 6 are closed when the lift of the intake valves 7i away from the inlets 6 is zero, and open, at least partially, otherwise.

The term "open" includes any position of the valves that is not closed -i.e. in a maximum open position or partially open.

In the exemplary embodiment, the maximum valve lift of each exhaust valve 7e away from the respective outlet 9 is in the range of 5 to 15 millimetres, and the maximum valve lift of each intake valve 7i away from the respective inlet 6 is in the range of 5 to 15 millimetres.

Figure 2 further shows a third plot 104 of the injection rate of gaseous fuel from the fuel injector 22 into the intake runner 16 with respect to the crank angle of the piston 20. In the illustrated embodiment, the gaseous fuel is hydrogen. Note that in Figure 2, the scaling of the first plot 100 and the second plot 102 is different to the scaling of the third plot 104.

Within a four-stroke cycle of the engine 1, the cylinder assembly 19 is configured to inject the gaseous fuel from the fuel injector 22 into the intake runner 16 exclusively during an interval X starting with a first crank angle Al of the piston 20, and ending with a second crank angle A2 of the piston 20.

The third plot 104 in Figure 2 shows the maximum interval X that the cylinder assembly 19 can inject the gaseous fuel into the intake runner 16. For this maximum interval X, the first crank angle Al corresponds to the intake valve 7i being closed and the exhaust valve 7e being open during the exhaust stroke, and the second crank angle A2 corresponds to the intake valve 7i being open and the exhaust valve 7e being closed during the intake stroke.

Advantageously, injecting gaseous fuel into the intake runner 16 over such an interval X may help ensure a sufficient volume of gaseous fuel is supplied to the cylinder 5 for near-optimal combustion using the single fuel injector 22. An extended angular interval X (and therefore an extended time proportional to engine speed) may also enhance the fuel air mixing, also promoting enhanced combustion.

In the illustrated embodiment, the first crank angle Al and the second crank angle A2 are both a function of an engine speed and an output torque of the engine 1. As such, the interval X is a function of the engine speed and the output torque of the engine 1. The output power (P) of the engine 1 is a function of the engine speed (S) and the output torque (T) of the engine 1; i.e. P = 2nTS/60, where P is in watts, T is in newton-metres, and S is in revolutions per minute. As such, in the illustrated embodiment, the first crank angle Al, the second crank angle A2, and thus the interval X, are a function of the output power of the engine 1.

In alternative embodiments (not shown), the first crank angle Al and the second crank angle A2, and thus the interval X, may not be a function of the engine speed and/or the output torque of the engine 1. For example, the first crank angle Al and the second crank angle A2, and thus the interval X, may be a function of the output torque or the engine speed of the engine 1 only.

In alternative embodiments (not shown), only one of the first crank angle Al and the second crank angle A2 may be a function of the output torque and/or engine speed of the engine 1. For example, the second crank angle A2 may be substantially constant.

Figure 3 shows an exemplary contour plot of the first crank angle Al with respect to the engine speed and the output torque of the engine 1. Similarly, Figure 4 shows an exemplary contour plot of the second crank angle A2 with respect to the engine speed and the output torque of the engine 1. The contour plots in Figures 3 and 4 were determined from on an optimisation of the first and second crank angles Al, A2 with respect to the engine speed and the output torque of the engine 1.

With reference to Figure 3, a minimum first crank angle Al at which the cylinder assembly 19 can inject the gaseous fuel into the intake runner 16 is at a maximum output torque of the engine 1 at a predetermined engine speed in the range of 1500 to 2600 RPM, optionally, 1800 to 2200 RPM, for example, approximately 2000 RPM. In the illustrated embodiment, the minimum first crank angle Al is approximately 316 degrees. In alternative embodiments (not shown), the minimum first crank angle Al may be in the range of 300 to 330 degrees, optionally, 310 to 320 degrees.

By 'maximum output torque at a predetermined engine speed' it is intended to mean that the output torque of the engine 1 is maximised for the given engine speed, and not necessarily that the output torque is at the maximum value capable of being produced by the engine 1 (i.e. the engine is operating at a maximum load). For example, in Figures 3 and 4, at an engine speed of 2000 RPM, the maximum output torque of the engine 1 is approximately 270 Nm. Whereas the maximum output torque capable of being produced by the engine 1 is approximately 450 Nm.

With reference to Figure 4, a maximum second crank angle A2 at which the cylinder assembly 19 can inject the gaseous fuel into the intake runner 16 is at a maximum output torque of the engine 1 at a predetermined engine speed in the range of 1200 to 1800 RPM, optionally, 1300 to 1700 RPM, for example approximately 1500 RPM.

The difference in engine speeds at the minimum first crank angle Al and the maximum second angle A2 in the exemplary contour plots of Figures 3 and 4 is due to the optimisation, which primarily determined the second crank angles A2 that optimise for fuel efficiency and emissions, and determined the first crank angles Al that ensure that sufficient fuel is injected into the cylinder 5 within each four-stroke working cycle of the engine 1, whilst integrity of the engine components are maintained.

In the exemplary plots shown in Figure 2, it can be seen that the intake valves 7i are open during the exhaust stroke, before the initiation of the intake stroke. As will be appreciated, during operation of the engine 1, this may result in a portion of the exhaust gases in the cylinder 5 flowing through the inlets 6 and into the intake runner 16. In diesel engines, opening the intake valves during an exhaust stroke can increase the efficiency of the engine as it reduces the pumping losses of the engine.

A problem with opening the intake valves 7i during the exhaust stroke in the engine 1 is that there is the potential for the exhaust gases flowing into the intake runner 16 to push air and fuel contained in the intake runner 16 upstream away from the inlets 6. In some cases, air and fuel may backflow into other intake runners 16 via the inlet manifold 11, causing misfiring of the corresponding cylinders 5 and/or poor combustion.

With reference to Figure 4, to help solve this problem, the maximum second crank angle A2 at which the cylinder assembly 19 can inject the gaseous fuel into the intake runner 16 is a crank angle of approximately 495 degrees after a top power dead centre of the piston 20. In alternative embodiments (not shown), the maximum second crank angle A2 may be a crank angle of less than 495 degrees after a top power dead centre of the piston 20.

For example, in some embodiments, the maximum second crank angle A2 at which the cylinder assembly 19 can inject the gaseous fuel into the intake runner 16 may be a crank angle of approximately 490 degrees or less after a top power dead centre of the piston 20.

Advantageously, it has been found that providing such an upper limit on the second crank angle A2 helps to inhibit backflowing of gaseous fuel and air within the intake runner 16.

In the illustrated embodiment, a minimum second crank angle A2 at which the cylinder assembly 19 can inject the gaseous fuel into the intake runner 16 is a crank angle of approximately 440 degrees after a top power dead centre of the piston 20. In alternative embodiments (not shown), the minimum second crank angle A2 may be a crank angle of more than 440 degrees, for example 450 degrees or more, after a top power dead centre of the piston 20.

The maximum interval X between the first crank angle Al and the second crank angle A2 that the cylinder assembly 19 can inject the gaseous fuel into the intake runner 16 is in the range of 160 to 200 degrees, optionally 170 to 190 degrees. In the illustrated embodiment, the maximum interval X is approximately 180 degrees.

Advantageously, such a maximum interval X may help to ensure a sufficient volume of gaseous fuel is supplied to the cylinder 5 for near-optimal combustion.

The maximum interval X is at a maximum output torque of the engine 1 at a predetermined engine speed, which may be in the range of 1600 to 2200 RPM, optionally in the range of 1800 to 2000 RPM.

For a constant output torque of the engine 1, the interval X increases as the engine speed of the engine 1 increases. Advantageously, this may help ensure that a sufficient volume of gaseous fuel is supplied to the cylinder 5 as the time duration of each four-stroke working cycle of the engine 1 reduces with increased engine speed.

In the illustrated embodiment, the engine 1 includes an engine control unit (not shown) configured to determine the interval X for each fuel injector 22 within each four-stroke cycle of the engine 1, and control each fuel injector 22 to inject fuel into the corresponding intake runner 16 exclusively during the determined interval X. The engine control unit may determine the interval X based on input engine speed and a torque output of the engine 1, for example using a look-up table or map stored in associated memory.

As shown in Figure 2, the fuel injector 22 is configured to inject the gaseous fuel continuously between the first crank angle Al and the second crank angle A2.

Advantageously, such a configuration of the fuel injector 22 may help to ensure a sufficient volume of gaseous fuel is supplied to the cylinder 5 for near-optimal combustion. In alternative embodiments (not shown), the fuel injector 22 may not be configured to inject the gaseous fuel continuously between the first crank angle Al and the second crank angle A2. For example, the fuel injector 22 may be configured to inject the gaseous fuel as a series of pulsed injections within the interval X. As shown in Figure 2, the fuel injector 22 is configured to inject the gaseous fuel at a substantially constant injection rate between the first crank angle Al and the second crank angle A2. In the exemplary embodiment, the fuel injector 22 injects the gaseous fuel at an injection rate in the range of 0.0005 to 0.0025 kilograms per second. Advantageously, such a configuration of the fuel injector 22 may help to simplify implementation of the engine 1.

In alternative embodiments (not shown), the fuel injector 22 may be configured to inject the gaseous fuel at a varying injection rate between the first crank angle Al and the second crank angle A2.

In the illustrated embodiment, gaseous fuel is injected into the intake runner 16 exclusively via the fuel injector 22.

In alternative embodiments (not shown), gaseous fuel may be injected into the intake runner 16 via a plurality of the fuel injectors 22, or via the fuel injector 22 in addition to any other suitable means.

Within a four-stroke cycle of the engine 1, each intake valve 7i moves from a closed position to an open position at a crank angle of the piston 20 in the range of 310 to 350 degrees, optionally in the range of 320 to 340 degrees, after a top power dead centre of the piston 20. In the illustrated embodiment, each intake valve 7i moves from a closed position to an open position at a crank angle of the piston 20 of approximately 335 degrees after a top power dead centre of the piston 20.

Within a four-stroke cycle of the engine 1, each intake valve 7i moves from an open position to a closed position at a crank angle of the piston 20 in the range of 570 to 610 degrees, optionally in the range of 580 to 600 degrees, after a top power dead centre (TPDC) of the piston 20. In the illustrated embodiment, each intake valve 7i moves from an open position to a closed position at a crank angle of the piston 20 of approximately 585 degrees after a top power dead centre of the piston 20.

Within a four-stroke cycle of the engine 1, each exhaust valve 7e moves from an open position to a closed position at a crank angle of the piston 20 in the range of 350 to 390 degrees, optionally in the range of 360 to 380 degrees, after a top power dead centre of the piston 20. In the illustrated embodiment, each exhaust valve 7e moves from an open position to a closed position at a crank angle of the piston 20 of approximately 360 degrees after a top power dead centre of the piston 20. In alternative embodiments (not shown), each exhaust valve 7e may move from an open position to a closed position at a crank angle of the piston 20 of approximately 375 degrees after a top power dead centre of the piston 20.

Within a four-stroke cycle of the engine 1, each exhaust valve 7e moves from a closed position to an open position at a crank angle of the piston 20 in the range of 110 to 150 degrees, optionally 120 to 140 degrees, after a top power dead centre of the piston 20. In the illustrated embodiment, each exhaust valve 7e moves from a closed position to an open position at a crank angle of the piston 20 of approximately 130 degrees after a top power dead centre of the piston 20.

Although not shown in the figures, in some embodiments, the engine 1 may be configured such that, within a four-stroke cycle of the engine 1: * the exhaust valve 7e moves from a closed to an open position at a crank angle of the piston 20 approximately 140 degrees after a top power dead centre of the piston 20; * the exhaust valve 7e moves from an open to a closed position at a crank angle of the piston 20 of approximately 380 degrees after a top power dead centre of the piston 20; * the intake valve 7i moves from a closed to an open position at a crank angle of the piston 20 of approximately 335 degrees after a top power dead centre of the piston 20; and * the intake valve 7i moves from an open to a closed position at a crank angle of the piston 20 of approximately 580 degrees after a top power dead centre of the piston 20.

Advantageously, such a configuration of the engine 1 has been found to minimise backflow of gaseous fuel and air in the intake runner 16, which may result in an improved volumetric efficiency of the engine 1.

As shown in Figure 2, within a four-stroke cycle of the engine 1, movement of each of the exhaust valve 7e and the intake valve 71 from a closed position to a maximum open position, and from the maximum open position to the closed position, is substantially symmetric with respect to the varying crank angle of the piston 20.

In alternative embodiments (not shown), movement of one or both of the exhaust valve 7e and the intake valve 7i from a closed position to a maximum open position, and from the maximum open position to the closed position may be asymmetric with respect to the varying crank angle of the piston 20.

With reference to Figures 1B and 1C, each cylinder 5 has a flat roof 30. In the illustrated embodiment, the flat roof 30 includes a face of the cylinder head 3 facing the cylinder 5.

The roof 30 and the piston 20 define a combustion chamber therebetween within the cylinder 5 when the piston 20 is at the top power dead centre position. The flat roof 30 is arranged substantially normal to the cylinder axis C. The intake valves 7i and the exhaust valves 7e are arranged to open and close via movement along respective axes that are parallel to the cylinder axis C. In the illustrated embodiment, the exhaust valves 7e and the intake valves 7i extend through the flat roof 30. Such arrangements are typical of diesel engines, so retaining this configuration may also allow for a common cylinder head, and certain valve train parts such as pushrods and rocker arms.

In alternative embodiments (not shown), the cylinders 5 may have e.g. a pent-roof or hemispheric roof and the valves may have axes that are inclined with respect to the axis C. With reference to Figures 1B and 1C, the cylinder assembly 19 is configured to inject the gaseous fuel into the intake runner 16 via the fuel injector 22 at an injection angle to a mean flow direction F of air flowing through the intake runner 16, in use, and/or the horizontal, in use, in the range of 0 to 25 degrees, optionally in the range of 10 to 20 degrees, for example 14 degrees.

Advantageously, such injection angles have been found to provide good mixing between fuel and air, and to inhibit backflow of fuel entering the intake runner 16, in use.

In the illustrated embodiment, the fuel injector 22 is received in a bore 40 in the intake assembly 10. The bore 40 is in fluid communication with a fuel injection passage 42 in the intake assembly 10. The fuel injection passage 42 includes a fuel orifice 44 within the intake runner 16. Fuel injected from the fuel injector 22 travels along the fuel injection passage 42 and into the intake runner 16 via the fuel orifice 44. The fuel injection passage 42 is configured to direct fuel exiting the fuel orifice 44 substantially along an injection axis A. In alternative embodiments (not shown), fuel may be injected into the intake runner 16 from the fuel injector 22 via any suitable arrangement. In some embodiments, the fuel injector 22 may be arranged in the cylinder head 3.

The cylinder assembly 19 is configured to direct the injected gaseous fuel from the fuel injector 22 along the injection axis A in a direction substantially towards the outlet portion 16o of the intake runner 16. The intake port 4 has a width W along an axis of the intake valve 7i closest to the intake runner 16. The injection axis A intersects said axis of the intake valve 7i at a position greater than 50%, optionally greater than 60%, optionally greater than 70%, of the width W of the intake port 4 from the corresponding inlet 6.

Advantageously, such configurations of the injection port 4 and the injection axis A have been found to provide good mixing between fuel and air in the intake port 4, and to inhibit backflow of fuel entering the intake runner 16 via the fuel injector 22, in use. Further, by having an injection axis orientated in this way, it may be less likely for gaseous fuel, particularly hydrogen that has a low autoignition temperature (around 500°C), to come into contact with hot exhaust gases, combust outside the cylinder and cause a backfire event.

With further reference to Figure 5, each cylinder assembly 19 includes an exhaust port 8 including an upstream end 8u leading away from the outlets 9. Each exhaust port 8 is arranged to transport exhaust gases away from the corresponding cylinder 5 towards a downstream portion 8d of the exhaust port 8. The upstream end 8u of the exhaust port 8 bifurcates to the outlets 9. Put another way, the upstream end 8u of the exhaust port 8 bifurcates into two branches, each branch leading to one of the outlets 9. In other embodiments (not shown) a non-bifurcating exhaust port 8 and a single outlet 9 may however be utilised.

The downstream end 4d of the intake port 4 bifurcates to the inlets 6. Put another way, the downstream end 4d of the intake port 4 bifurcates into two branches, each branch leading to one of the inlets 6. One branch is longer than the other so that the ports have a so-called "tandem" configuration. In this configuration the ports are arranged transverse to an axis defined by a crankshaft of the engine. The outlet ports are similarly arranged. Such an arrangement is typically used to generate a swirl motion of air in the cylinder that is desirable for efficient compression ignition combustion in diesel engines.

Spark ignition engines typically use a cross-flow configuration to generate a desired tumble motion in the cylinder for efficient combustion. It has however been found that the combination of port fuel injection of hydrogen, the valve and injection timing described above and this tandem configuration efficient combustion can nevertheless be achieved In other embodiments (not shown) a non-bifurcating intake port 4 and single inlet 6 may however be utilised.

As described above it is desirable to utilise common parts from pre-existing diesel engines for hydrogen fuelled engines to minimise tooling costs and inventory costs, as well as costs associated with repackaging an IC engine into a particular application. The use of common parts may similarly retain certain characteristics of diesel engines that are advantageous for off-highway and/or heavy commercial vehicle applications, in particular high and broad torque band starting at a low engine speed.

Claims

CLAIMS1. A four-stroke gaseous fuel engine comprising one or more cylinder assemblies, each cylinder assembly comprising: a cylinder including an inlet and an outlet selectively opened and closed by an intake valve and an exhaust valve respectively; a piston translationally movable within the cylinder; an intake runner leading to the inlet; and a fuel injector, wherein the cylinder assembly is configured to selectively inject a gaseous fuel from the fuel injector into the intake runner, wherein the engine is configured to supply gaseous fuel and air from the intake runner to the cylinder via the inlet during an intake stroke of the piston, and exhaust combustion gases from the cylinder via the outlet during an exhaust stroke of the piston, wherein, within a four-stroke cycle of the engine, the cylinder assembly is configured to inject the gaseous fuel into the intake runner exclusively during an interval starting with a first crank angle of the piston and ending with a second crank angle of the piston, and wherein at a maximum interval that the cylinder assembly can inject the gaseous fuel into the intake runner, the first crank angle corresponds to the intake valve being closed and the exhaust valve being open during the exhaust stroke, and the second crank angle corresponds to the intake valve being open and the exhaust valve being closed during the intake stroke.
2. The engine of claim 1, wherein a maximum second crank angle at which the cylinder assembly can inject the gaseous fuel into the intake runner is a crank angle of 495 degrees or less after a top power dead centre of the piston; optionally, 490 degrees or less after a top power dead centre of the piston.
3. The engine of claim 1 or claim 2 wherein a minimum second crank angle at which the cylinder assembly can inject the gaseous fuel into the intake runner is a crank angle of 440 degrees or more, optionally 450 degrees or more, after a top power dead centre of the piston.
4. The engine of any preceding claim, wherein the gaseous fuel is hydrogen.
5. The engine of any preceding claim, wherein the maximum interval is in the range of 160 to 200 degrees, optionally 170 to 190 degrees, for example approximately 180 degrees.
6. The engine of any preceding claim, wherein the interval is a function of an output torque and/or an engine speed of the engine.
7. The engine of claim 6, wherein the maximum interval is at a maximum output torque at a predetermined engine speed; optionally, in the range of 1600 to 2200 RPM, optionally in the range of 1800 to 2000 RPM.
8. The engine of claims 6 or 7, wherein a minimum first crank angle at which the cylinder assembly can inject the gaseous fuel into the intake runner is at a maximum output torque at a predetermined engine speed in the range of 1500 to 2600 RPM, optionally, 1800 to 2200 RPM, for example, approximately 2000 RPM.
9. The engine of any one of claims 6 to 8, wherein a maximum second crank angle at which the cylinder assembly can inject the gaseous fuel into the intake runner is at a maximum output torque at a predetermined engine speed in the range of 1200 to 1800 RPM, optionally 1300 to 1700 RPM, for example approximately 1500 RPM.
10. The engine of any one of claims 6 to 9, wherein, for a constant output torque of the engine, the interval increases as the engine speed increases.
11. The engine of any preceding claim, wherein the fuel injector is configured to inject the gaseous fuel continuously between the first and second crank angles.
12. The engine of claim 11, wherein the fuel injector is configured to inject the gaseous fuel at a substantially constant injection rate between the first and second crank angles.
13. The engine of any preceding claim, wherein gaseous fuel is injected into the intake runner exclusively via the fuel injector.
14. The engine of any preceding claim, wherein, within a four-stroke cycle, the intake valve moves from a closed position to an open position at a crank angle of the piston in the range of 310 to 350 degrees, optionally in the range of 320 to 340 degrees, for example approximately 335 degrees, after a top power dead centre of the piston.
15. The engine of any preceding claim, wherein, within a four-stroke cycle, the intake valve moves from an open position to a closed position at a crank angle of the piston in the range of 570 to 610 degrees, optionally in the range of 580 to 600 degrees, for example approximately 585 degrees, after a top power dead centre of the piston.
16. The engine of any preceding claim, wherein, within a four-stroke cycle, the exhaust valve moves from an open position to a closed position at a crank angle of the piston in the range of 350 to 390 degrees, optionally in the range of 360 to 380 degrees, for example approximately 375 degrees, after a top power dead centre of the piston.
17. The engine of any preceding claim, wherein, within a four-stroke cycle, the exhaust valve moves from a closed position to an open position at a crank angle of the piston in the range of 110 to 150 degrees, optionally in the range of 120 to 140 degrees, for example approximately 130 degrees, after a top power dead centre of the piston.
18. The engine of any preceding claim, wherein, within a four-stroke cycle: the exhaust valve moves from a closed to an open position at a crank angle of the piston of approximately 140 degrees after a top power dead centre of the piston; the exhaust valve moves from an open to a closed position at a crank angle of the piston of approximately 380 degrees after a top power dead centre of the piston; the intake valve moves from a closed to an open position at a crank angle of the piston of approximately 335 degrees after a top power dead centre of the piston; and the intake valve moves from an open to a closed position at a crank angle of the piston of approximately 580 degrees after a top power dead centre of the piston.
19. The engine of any preceding claim, wherein the cylinder comprises a flat roof arranged substantially normal to an axis of piston motion, and wherein the intake valve and the exhaust valve are arranged to open and close via movement along respective axes that are parallel to said axis of piston motion.
20. The engine of any preceding claim, wherein the cylinder assembly is configured to inject the gaseous fuel into the intake runner from the fuel injector at an injection angle to a mean flow direction of air flowing through the intake runner and/or the horizontal, in use, in the range of 0 to 25 degrees, optionally in the range of 10 to 20 degrees, forexample 14 degrees.
21. The engine of any preceding claim, wherein each cylinder assembly comprises an intake port leading to the inlet, wherein an outlet portion of the intake runner leads to the intake port, wherein the cylinder assembly is configured to direct the injected gaseous fuel from the fuel injector along an injection axis in a direction substantially towards the outlet portion, wherein the intake port has a width along an axis of the intake valve, and wherein the injection axis intersects the axis of the intake valve at a position greater than 50%, optionally greater than 60%, optionally greater than 70%, of the width of the intake port from the inlet.
22. The engine of any preceding claim, wherein the cylinder comprises a further inlet and a further outlet selectively opened and closed by a further intake valve and a further exhaust valve respectively, wherein the intake runner leads to the further inlet, and wherein the engine is configured to supply gaseous fuel and air from the intake runner to the cylinder via the further inlet during the intake stroke of the piston, and exhaust combustion gases from the cylinder via the further outlet during the exhaust stroke of the piston.
23. The engine of claim 22, further comprising: an intake port including a downstream end leading to the inlet and the further inlet; and an exhaust port including an upstream end leading from the outlet and the further outlet, wherein the intake runner leads to an upstream end of the intake port, the downstream end of the intake port bifurcating to the inlet and the further inlet, and wherein the ports are arranged in a tandem configuration.
24. A working machine comprising the engine of any preceding claim.
25. A method for operating a four-stroke gaseous fuel engine configured to supply gaseous fuel and air from an intake runner to a cylinder via an inlet of the cylinder during an intake stroke of a piston within the cylinder, and exhaust gases from the cylinder via an outlet of the cylinder during an exhaust stroke of the piston, the inlet selectively opened and closed by an intake valve, and the outlet selectively opened and closed by an outlet valve, the method comprising the step of: within a four-stroke cycle of the engine, injecting a gaseous fuel into the intake runner exclusively during an interval starting with a first crank angle of the piston and ending with a second crank angle of the piston, wherein at a maximum interval that the cylinder assembly can inject the gaseous fuel into the intake runner, the first crank angle corresponds to the intake valve being closed and the exhaust valve being open during the exhaust stroke, and the second crank angle corresponds to the intake valve being open and the exhaust valve being closed during the intake stroke.