WO2006079818A1

WO2006079818A1 - Internal combustion engine

Info

Publication number: WO2006079818A1
Application number: PCT/GB2006/000270
Authority: WO
Inventors: Jean-Pierre Pirault; Thomas William Ryan, Iii; Terrance Francis Ii Alger; Charles Edward Junior Roberts; Ryan C. Roecker; Manfred Amann
Original assignee: South West Research Institute
Priority date: 2005-01-26
Filing date: 2006-01-26
Publication date: 2006-08-03
Also published as: GB0501632D0

Abstract

A four stroke internal combustion engine has a cylinder (13) with first inlet port 1 directed to discharge in a direction tangential to the periphery of the cylinder (13) and a second inlet part directed to discharge towards the centre of the cylinder (13). A first injector has a spray targeted symmetrically on the up-stream side of a head of a first inlet valve (7) and a second injector has a spray targeted towards the side of a head of a second inlet valve (8) which is closest to the centre of the cylinder (13). There is also provided means by which more exhaust gas recirculation can be supplied to the first inlet port (1) than the second inlet port (2) and a means by which the flow rate through the second inlet port (2) can be restricted.

Description

Internal Combustion Engine

Field of the invention

The present invention relates to an internal combustion engine having a plurality of cylinders, at least two inlet valves per cylinder and separate inlet ports delivering air and fuel to each inlet valve, and to a method of operating such an engine.

Background to the Invention

Regulatory authorities are continually demanding that internal combustion engines have reduced emissions. At the same time, there is an economic demand to see engine fuel efficiency increase. Gasoline fuel engines are relatively fuel inefficient in comparison to diesel engines. However, they are able to use highly efficient 3-way catalyst systems, where conversion efficiencies of above 98% are readily achievable when the catalyst temperature exceeds the threshold light-off level, giving them very low emissions. By contrast, more fuel efficient diesel engines tend to have higher emissions. For example, there are serious challenges in the aftertreatment of particulate material (PM). NOx emissions are also particularly difficult to handle in a diesel exhaust due, at least in part, to the presence of excess oxygen, which makes the reduction of NOx dependent on added exhaust borne reductants, such as urea, or the use of NOx storage and subsequent programmed intermittent reduction with rich air/fuel ratios which are naturally reducing, ie oxygen stripping. This is complicated and expensive. So, there is current interest in increasing the fuel efficiency of gasoline engines while maintaining their low emissions. It is well known that a proportion of the exhaust gas from an internal combustion engine can be fed back through the inlet system into the combustion chamber to provide a reduction in combustion temperature. This feature is known as exhaust gas recirculation or EGR. The main aim and benefit of this is usually to reduce NOx emissions. However, there is also now evidence to show that cooled EGR can provide substantial mitigation of detonation, probably because EGR reduces reaction rates in the combustion chamber. Whilst this is not typically a problem in diesel engines, gasoline engines rely on a specific ratio of fuel to air, known as stoichiometric conditions, to achieve combustion. So, percentages of EGR above a minimum value, which may in some cases be 30%, can prevent combustion in gasoline engines.

US 2005/0011485 discloses a gasoline engine which enables the use of higher compression ratios than would normally be achievable with a conventional spark ignition (Sl) system, by the use of specific fast burn techniques in combination with high levels of cooled EGR. The fast burn allows late burning, with acceptable fuel efficiency, and both the late burning and the cooled EGR mitigate detonation, which is the critical phenomenon limiting compression ratio. The publication mentions that pilot fuel injection can be used to improve combustion stability of the engine. However, it does not recognise that the air-fuel mixture may vary spatially throughout the cylinder, nor how to ensure that ignition is maintained in the presence of high levels of EGR, particularly in light of such variations in the air-fuel mixture. US 5762041 discloses a gasoline engine having divided inlet ports. A first inlet port discharges substantially in a direction tangential to the periphery of the cylinder as viewed in the direction of the central axis (e.g. is a so-called tangential inlet port) and a second inlet port discharges substantially towards the centre of the cylinder as viewed in the direction of its central axis (e.g. is a so-called neutral port). This allows air flow in the cylinder to be controlled during compression to provide stable conditions at an ignition location at the centre of the cylinder. However, as a fuel injector is only provided in the neutral port, it is difficult, if not impossible, to control the air-fuel mixture throughout the cylinder. The present invention seeks to overcome these problems.

Summary of the invention According to one aspect of the present invention, there is provided a four stroke internal combustion engine comprising: a cylinder having a central axis along which a piston can travel with a reciprocating motion; a first inlet port for supplying the cylinder with fuel and air through a first inlet valve, the first inlet port being directed to discharge substantially in a direction tangential to the periphery of the cylinder as viewed in the direction of its central axis; and a second inlet port for supplying the cylinder with fuel and air through a second inlet valve, the second inlet port being directed to discharge substantially towards the centre of the cylinder as viewed in the direction of its central axis; wherein the engine has individual injectors for injecting fuel into each inlet port of the cylinder. So, as injectors are provided for each inlet port, the air/fuel mixture can be controlled throughout the cylinder. This enables faster burning combustion, which can also be ignited later in the engine cycle to avoid detonation and therefore permit higher compression ratios for a given fuel and air/fuel ratio versus conventional spark ignited SI combustion systems. The invention relates to an engine using predominantly gasoline fuel but with the possibility of using diesel fuel as an ignition source and to an engine in which a proportion of the exhaust gas is recirculated into the combustion chamber.

Expressed in different terms, according to another aspect of the present invention, there is provided a four stroke internal combustion engine containing a plurality of cylinders each of which has an axis along which a piston travels with a reciprocating motion and in which is a combustion chamber with at least two ignition sources and which is supplied with fuel and air through two inlet valves each communicating with a separate inlet port, the first of which is directed to discharge substantially in a direction tangential to the periphery of the cylinder, as viewed in the direction of its axis, and the second of which is directed to discharge substantially towards the centre of the cylinder, as viewed in the direction of its axis, and a first injector having a spray targeted substantially symmetrically on the up stream side of the head of the first inlet valve, and a second injector with a spray targeted substantially towards the side of the second inlet valve head which is closest to the centre of the cylinder, when viewed along its axis, and a means by which more exhaust gas recirculation is supplied to the first port than the second port, in some operational modes, and a means by which the flow rate through the second port is restricted, in some operational modes.

At least one injector may be mounted in the cylinder head, connecting to one port. Alternatively, both injectors may be mounted in the cylinder head, each connecting to different inlet ports; or at least one injector may be mounted in the inlet manifold.

At least one injector may use auxiliary air to improve fuel preparation. Typically, the injector in the first port uses auxiliary air to improve fuel preparation. Additionally or alternatively, the injector in the second port uses auxiliary air to improve fuel preparation. So, both injectors may use auxiliary air to improve fuel preparation.

The first and the second injectors may be timed, for some operational modes, to inject during the induction stroke. For example, the first injector may be timed, for some operational modes, to inject at different periods of the induction stroke, relative to the timing of the second injector. In other examples: the first and second injectors may be timed, for some operational modes, to inject during the same period of the induction stroke; the first injector may be timed, for some operational modes, to inject in advance of the second injector; the first injector may be timed, for operational modes when the flow through the first port is substantially larger than the flow through the second port, to inject in advance of the second injector; and/or the first injector may be timed, for operational modes when the exhaust gas recirculation flow through the first port is substantially larger than the exhaust gas recirculation flow through the second port, to inject in advance of the second injector. In these examples, the first port injector may be timed to begin injection between 0-140 deg ATDC induction. Finally, in another example, the first injector may be timed to inject during the expansion and exhaust strokes of the preceding cycle, and the second injector timed, for some operational modes, to inject during the induction stroke.

In some operational modes, the flow through the first port is substantially larger than the flow through the second port. For example, exhaust gas recirculation flow through the first port can be substantially larger, in some operational modes, than the exhaust gas recirculation flow through the second port. Total exhaust gas recirculation (EGR), defined as the percentage mass of carbon dioxide gas in the inlet charge relative to the total mass of carbon dioxide in the exhaust upstream of any catalysts, may be less than 70%. Similarly, Total exhaust gas recirculation, as previously defined, may be split so that up to 99% of the total EGR is in the first port.

A first spark plug may have its discharge electrodes within 10% cylinder bore radius of the centre of the major cylinder axis. A spark plug may also be located with its discharge electrodes substantially at the centre of major cylinder axis. A second spark plug may have its discharge electrodes substantially off centre of major cylinder axis. A second spark plug may have its discharge electrodes within 90% cylinder bore radius of the centre of major cylinder axis. The second spark plug may have its discharge electrodes on the same side of the cylinder head as the second port.

The combustion system may have 4-valves per cylinder. The bulk of compressed charge at TDC may be in the cylinder head clearance volume. The inlet and exhaust valves may have an included angle between their principle axes that is greater than 15 degrees. The combustion chamber clearance volume may be substantially within the envelope of the piston and the lower surface of the cylinder head is substantially flat. A piston bowl of the combustion system may be re-entrant in section. The piston bowl may have a protrusion, within the bowl, located centrally on the principal axis of the piston bowl. The inlet and exhaust valves may have an included angle between their principle axes that is less than 15 degrees. Alternatively, the inlet and exhaust valves' principle axes may be substantially parallel. The first and second injectors may have different spray plume patterns. The first and second injector sprays may have different Sauter mean droplet sizes.

The ignition source may be an in-cylinder injector delivering a relatively small proportion of a cetane based diesel fuel, in comparison to the gasoline fuel provided by the port injectors. The in-cylinder injector may inject the diesel fuel late in the compression stroke. Indeed, the in-cylinder injector may inject the diesel fuel at least 150 degrees after bottom dead centre of the induction stroke. At least one of the spark plugs may be supplied with an ignition system that has multi-strike capability. For example, the spark plug located near the second port may have the multi-strike capability.

In other examples: one or more of the ignition means may be of the corona discharge type; one or more of the ignition means may be of the rail plug type; and/or one or more of the ignition means may be of the capacitive discharge type.

The combustion system may be used with a turbocharged and intercooled engine. It might also be used with exhaust gas recirculation that is supplied between the outlet of the intercooler and the inlet ports; with cooled exhaust gas recirculation; or in conjunction with a compressor expander down stream of the first intercooler, and with an intercooler between the compressor and expander of the compressor-expander.

Preferred embodiments of the invention are now described, by way of example only, with reference to the accompanying drawings.

Brief Description of the Drawings

Figure 1 is a schematic view from above of a cylinder of an internal combustion engine according to a preferred embodiment of the present invention.

Figure 2 is a notional side view of the cylinder shown in Figure 1 , showing the main flow paths from two inlet ports of the cylinder. Figure 3 is a notional side view of the cylinder shown in Figure 1 , showing the main flow paths of fuel spray into and leaving a first of the inlet ports.

Figure 4 is a notional side view of the cylinder shown in Figure 1 , showing the main flow paths of fuel spray into and leaving a second of the inlet ports.

Figure 5 is a schematic view from above of three adjacent cylinders of the internal combustion engine shown in Figure 1.

Detailed Description of the Preferred Embodiments

With reference to Figures 1 to 4, a cylinder 13 is arranged in an internal combustion engine with a first inlet valve 7, a second inlet valve 8, a first exhaust valve 11 , a second exhaust valve 12, a first inlet port 1 connecting an inlet manifold (not shown) to the first inlet valve 7 and cylinder 13, a second inlet port 2 connecting the inlet manifold (not shown) to the second inlet valve 8 and cylinder 13, a first injector 120 located in the first inlet port 1 and a second injector 121 located in the second inlet port 2. The spray 115 of the first injector 120 is targeted symmetrically on the back of the valve head 117 of the first inlet valve 7. The path of the spray 115 is depicted by arrow 6 in Figure 1. Air flow in the first inlet port 1 is depicted by arrow 15 in Figure 1 and air flow in the first valve 7 is depicted by arrow 3 in Figure 2. Flow of the spray 115 from the valve 7 is also depicted by arrows 122 in Figure 3. The spray 118 of the second injector 121 is targeted towards the side of the valve head 123 of the second inlet valve 8. The path of the spray 118 is depicted by arrow 5 in Figure 1. Air flow in the second inlet port 2 depicted by arrow 16 in Figure 1 and air flow from the second valve 8 is depicted by arrow 4 in Figure 2. Flow of the spray 115 from the valve 8 is also depicted by arrows 119 in Figure 4. The cylinder 13 has at least two sources of high energy ignition, which in this particular embodiment are spark plug locations 9, 10, one spark plug location 9 being substantially on or within 10% radially of the major cylinder axis, the other spark plug location being within 90% of the cylinder bore radius relative to the cylinder centre.

The first inlet port 1 is arranged to provide air flow into the cylinder 13 that is substantially tangential to the cylinder bore, as depicted notionally by the arrow 3 in Figure 2, whilst the second inlet port 2 discharges air with a bias towards the central axis of the cylinder 13, as notionally depicted by arrow 4 in Figure 2. Both of the inlet ports 1, 2 may receive hot or cooled exhaust gas recirculation (EGR), with the first inlet port 1 receiving relatively more or all the EGR under some operating conditions. The second inlet port 2 has some means of reducing airflow. In this embodiment, this means comprises by a throttle valve or obturator 14. Alternative means of reducing the airflow include reducing the maximum lift of second inlet valve 8, or use of a sliding throttle plate located between the entry 17 to the second inlet port 2 and the manifold flange. Each injector 120, 121 may optionally have its own particular spray pattern and droplet size characteristics, as defined by the Sauter Mean Droplet Size (SMDS). As an example, the first injector 120 may have a spray included angle of approximately 20° or less, and an SMDS of 100 microns or less, whilst the second injector 121 may have a spray included angle of approximately 15° or less and an SMDS of 180 microns or less. Each injector 120, 121 may be timed to deliver all its fuel during the inlet valve open period, with the first injector 120 timed to begin injection between 0 and 140 crankangle degrees (CA) after top dead centre (ATDC) induction in some operating modes, and with the second injector 121 delivering its fuel later than the first injector 120 in some operating modes, e.g 60 to 160 CA ATDC start of injection. In other operating modes, the first injector 120 may deliver its fuel onto a closed inlet valve during the preceding expansion and exhaust strokes, e.g with start of injection at 0 to 220 CA after inlet valve closing.

The ignition sources may comprise either two spark plugs supplied with high levels of electrical energy, or an in-cylinder injector delivering a very small quantity of diesel fuel which acts as a multiple ignition source to the gasoline fuel provided by the port injectors. In the case of the two spark plugs, one of the spark plugs is located at or within 10% of the radius distance from the centre of the cylinder, when viewed along its axis, and one of these spark plugs is located well off the cylinder centre axis, typically up to 90% of the radius distance from the cylinder centre. In one embodiment, as shown in Figurei , this other spark plug is located between the outer peripheries of the inlet and exhaust valve heads and the periphery of the cylinder 13. In some embodiments it will be preferable to position the second spark plug on the same side of the cylinder 13 as the second inlet port 2.

The principle behind this arrangement is to generate locally different mixture strengths in the cylinder 13, depending particularly on speeds and loads, the key operating modes now being described.

It should be noted that in the following description references will be made to a total EGR figure given as a percentage. Total EGR of 40%, for example, means in effect that EGR carbon dioxide content is approximately 40% of the content of carbon dioxide leaving the exhaust system. The remainder of the charge typically consists of air and fuel.

It should also be noted that the invention applies equally to turbo- charged engines and those which are turbo-charged and intercooled. The EGR is typically supplied to the inlet system at a point between the outlet of the intercooler, or turbo-charger, and the inlet ports 1 , 2. Another possible arrangement is to have a compressor-expander downstream of the first intercooler and to have a second intercooler in between the compressor and expander. In the latter case, EGR may be added after the compressor- expander. The invention also allows for the use of cooled EGR. In one mode, corresponding to relatively steady-state full load across the speed range, both inlet ports 1 , 2 will be flowing substantially similar amounts of charge, as the second inlet port 2 will have no flow restriction in this mode. The first inlet port 1 will have the bulk of the EGR, eg 80% of the total of perhaps 40% cooled EGR (ie EGR which is cooled and which represents 40% of the total charge) at this condition. The fuel in the first inlet port 1 could either be injected during the preceding exhaust stroke, or during the open inlet valve period, whilst the fuel in the second inlet port 2 would be injected only during the open valve period, and as late as possible within the constraints of the dynamic capability of the injector. Ignition will begin readily and rapidly in the relatively undiluted mixture near the spark plug, with the start of combustion aided by the high energy ignition, and this flame will spread to the relatively diluted EGR/stoichiometric mixture that is in the outer core of the cylinder 13, and has a reduced tendency to detonate, due to the high levels of cooled EGR diluent. There will be some residual tumble and swirl bulk charge motion during compression, but this will be relatively low as the first and second inlet port flows will tend to interact. The second ignition source may be optionally used to help to accelerate the burn rate, particularly during the expansion stroke, but use of the second spark plug will depend on the rate of pressure rise, which may be unacceptably high at full load.

In another mode corresponding to full load transients across the speed range, both inlet ports 1 , 2 will be flowing substantially similar amounts of charge, as the second port 2 will have no flow restriction in this mode. The first port 1 will have the bulk of the EGR, eg 80% of a total of perhaps 40% cooled EGR at this condition. As the engine speed is changing moderately, the fuel in the first inlet port 1 would be injected during the open inlet valve period, and the fuel in the second inlet port 2 would be injected during the open valve period, and as late as possible within the constraints of the dynamic capability of the injector. The injection during the open valve period is important to generate charge stratification, as well as helping to maintain better dynamic control of air/fuel ratio, by reducing the effects of wall wetting. Ignition will begin readily and rapidly in the relatively undiluted mixture near the spark plug, with the start of combustion aided by the high energy ignition, and this flame will spread to the relatively diluted EGR/stoichiometric mixture that is in the outer core of the cylinder 13, and has a reduced tendency to detonate, due to the high levels of cooled EGR diluent. There will be some residual tumble and swirl bulk charge motion during compression, but this will be relatively low as the first and second port flows will tend to interact. The second ignition source may be optionally used to help to accelerate the burn rate, particularly during the expansion stroke, but this will depend on the rate of pressure rise, which may be unacceptably high at full load.

In another mode, corresponding to moderate part load operation at moderate speeds, e.g. less than 50% load and less than 75% rated speed, the first inlet port 1 will be flowing the bulk of the working charge, eg 80% of total charge, and the second inlet port 2 will have a flow restriction in this mode, flowing the balance of the charge, ie 20% in this example. The first inlet port 1 will have the bulk of the EGR, eg 80% of a total of perhaps 60% EGR, which may be cooled or uncooled at this condition, depending on the NOx emission strategy. The fuel in the first inlet port 1 would be injected during the open inlet valve period, eg 70 CA ATDC, and the fuel in the second inlet port 2 would also be injected during the open valve period, e.g. with start of injection and as late as possible within the constraints of the dynamic capability of the injector, e.g. with start of injection 130 CA ATDC. The fuel in the first inlet port 1 is transported readily by the relatively high velocities existing in the first inlet port 1 , with 80% of the charge passing through it, and with the injection occurring during the period of highest gas velocity in the intake ports, ie 60- 120 CA ATDC. This fuel, which is gradually evaporating as it travels, impinges substantially symmetrically on the back of the open inlet valve head, enters the cylinder 13 with the charge having a combination of high tangential velocities and low levels of axial velocity towards the retreating piston; the resultant charge motion is a type of slow spiralling towards the retreating piston crown, occurring mainly in the outer core of the cylinder charge. By contrast, the partial closure of the second inlet port 2 will restrict the charge velocities in the second inlet port 2, and the cylinder centred discharge focus of this port 2 results in the charge entering the centre portion of the cylinder 13, where charge motion is low, the retreating piston sucking more of this centre core into the cylinder 13, surrounded by the bulk spiralling motion of the charge from the first inlet port 1. As the fuel is injected relatively late in the second inlet port 2, it will tend to remain in the upper regions of the cylinder core, ie near the axis of central ignition source. At these part load conditions, ignition will begin readily and rapidly in the relatively undiluted mixture near the spark plug, with the start of combustion aided by the high energy ignition, and this flame will spread to the relatively diluted EGR/stoichiometric mixture that is in the outer core of the cylinder 13. As the second inlet port flow is highly reduced by the upstream throttling device in the second port, the first inlet port 1 generates a relatively strong bulk swirl, compared to the high to full load conditions, and this swirl can be arranged, with the appropriate open valve injection timing, to carry the first port injected fuel around the cylinder periphery such that the first port fuel cloud, during the final stages of compression, arrives in the region of the second ignition source, and ignites readily, as the fuel in this cloud will only have been partially diffused by the swirling motion and evaporation. A well known feature of an ignition source in a high swirling flow is that the resultant burning from this kernel is characterised by the burn products of combustion being centrifuged towards the cylinder centre, due to the relative density effects of burned and unburned gases, and creating a strong turbulent field in the flamefront and unburnt mixture, ie rapidly accelerating the burning of the diluted mixture in this zone of the cylinder 13. This additional ignition source therefore provides a useful means of extending the burning capability of the combustion system in highly diluted flows, such as those using high levels of EGR with a stoichiometric air/fuel ratio. It should be noted that the bulk swirling of this mode reduces almost entirely as the flow to the second inlet port 2 is increased, substantially counteracting the swirl motion of the first inlet port 1.

An alternative means of achieving at least two ignition sources is by in- cylinder injection of a very small quantity of high cetane fuel, such as diesel fuel, into the compressed gasoline fuel/air and EGR charge, late in the compression stroke. Injection may occur at least 150 degrees after bottom dead centre after completion of the induction stroke. With suitable compression ratios, the high cetane fuel autoignites, and provides ignition sites for the surrounding gasoline mixture. The in-cylinder injector, which may be used without any spark plugs, or with one or two spark plugs, may be arranged to have at least two sprays in order to increase the spatial distribution of the injected high cetane fuel, and these sprays can, in some embodiments, be targeted at one or two spark plugs that are away from the cylinder centre, the spark plugs providing additional energy to assist the autoignition of the high cetane fuel, and ensuring, by appropriate location of the spark plugs, that ignition starts in the evaporated portion of the relatively fuel rich mixture zones. This method of injecting a high cetane fuel into a predominantly high octane fuel is sometimes referred to as "micropilot" injection. The in-cylinder injector is usually positioned substantially centrally in the cylinder, as viewed along its axis.

Another important enabling feature of the claimed invention is the use of high energy ignition, in combination with the other features described. The high energy ignition provides high energy levels to ionise the charge and creates high temperature intense heat for local breakdown of the fuel in the dilute mixtures, particularly when ignition is very retarded and when an ignition kernel has to be established as the cylinder contents are expanded as the piston descends from top dead centre (TDC). Various high energy ignition means may be used, the highest energy being provided chemically in the form of micropilot injected high cetane fuel. Other high energy means may use electrical energy, such as electrical ignition discharges that have a high capacitive content, so called Corona discharge systems that use radio frequency discharges, or "rail" sparking plugs that generate a large annular plasma in a spark plug tunnel, or prechamber type ignitors which are arranged to receive and ignite small amounts of suitably rich mixture in a confined volume with at least one small passage connecting with the bulk cylinder contents. The resulting large pressure rise in the small prechamber volume creates a high velocity burning jet from the connecting passage with the main cylinder contents, helping to create rapid burning.

Another embodiment uses an ignition system that has a multi-strike capability. That is to say the electrical energy delivery systems provides one or more of the spark plugs with multiple sparks within one piston stroke. With reference to Figure 5, adjacent cylinders 13, 13a, 13b may be arranged so that the second inlet port 2 of a first of the cylinders 13, 13a, 13b is adjacent the second inlet port 2a of a second of the cylinders 13, 13a, 13b and the first inlet port 1 of the first cylinder 13 is adjacent the first inlet port 1b of a third of the cylinders 13, 13a, 13b. This can simplify the means used to control the relative airflow and EGR between pairs of ports 1 , 1a, 1b, 2, 2a, 2b for each cylinder 13, 13a, 13b, as is required for some operational modes. Similarly, the first inlet port 1a of the second cylinder 13a would be adjacent to the first inlet port (not shown) of the next cylinder (not shown) and likewise the second inlet port 2b of the third cylinder 13b would be adjacent to the second inlet port (not shown) of the next cylinder (not shown) and so on.

The aforementioned port and injector configuration may be contained in a cylinder head in which the bulk of the compressed cylinder charge at top dead centre (TDC) is in the cylinder head clearance volume, the piston being either substantially flat or with either slightly convex or concave surfaces. An alternative arrangement (not shown), still with the ports, valves and ignition means contained substantially within the cylinder head, is to have a deep concave shape, known as a bowl, in the top of the piston such that the bulk of the clearance volume at TDC is within the piston and the lower surface of the cylinder head is substantially flat. The piston bowl will in some embodiments have a re-entrant shape, that is to say a shape which is narrower at its uppermost point (i.e. the point closest to the cylinder head) than at its widest point. It may also have a shape including a raised protrusion substantially at the centre of the piston when viewed on the cylinder axis.

It can also be noted that the angle α between the inlet and exhaust valves 7, 8 of the cylinder 13 can in some embodiments be at or close to zero, less than 15 degrees or it may be preferable in some embodiments for the angle to exceed 15 degrees.

The injectors described previously are preferably mounted in the cylinder head such that they protrude into the inlet ports 1 , 2. However in an alternative arrangement they may be mounted in the inlet manifold, upstream of the inlet ports. It may also be arranged for one or more of the injectors, whether mounted in the cylinder head or in the manifold, to be supplied with pressurised auxiliary air to improve the fuel spray characteristics. The described embodiments of the invention are only examples of how the invention may be implemented. Modifications, variations and changes to the described embodiments will occur to those having appropriate skills and knowledge. These modifications, variations and changes may be made without departure from the spirit and scope of the invention defined in the claims and its equivalents.

Claims

1. A four stroke internal combustion engine comprising: a cylinder having a central axis along which a piston can travel with a reciprocating motion; a first inlet port for supplying the cylinder with fuel and air through a first inlet valve, the first inlet port being directed to discharge substantially in a direction tangential to the periphery of the cylinder as viewed in the direction of its central axis; and a second inlet port for supplying the cylinder with fuel and air through a second inlet valve, the second inlet port being directed to discharge substantially towards the centre of the cylinder as viewed in the direction of its central axis; wherein the engine has individual injectors for injecting fuel into each of the first and second inlet ports of the cylinder.

2. The engine of claim 1 , wherein a first of the injectors has a spray targeted substantially symmetrically on an up stream side of a head of the first inlet valve.

3. The engine of claim 1 or claim 2, wherein a second of the injectors has a spray targeted substantially towards a side of a head of the second inlet valve which is closest to the centre of the cylinder when viewed along its axis.

4. The engine of any one of the preceding claims, wherein at least one of the injectors is mounted in an inlet manifold of the engine.

5. The engine of any one of the preceding claims, wherein at least one of the injectors is arranged to use auxiliary air to improve fuel preparation.

6. The engine of any one of the preceding claims, wherein the injectors have different spray plume patterns.

7. The engine of any one of the preceding claims, wherein the injectors have sprays with different Sauter mean droplet sizes.

8. The engine of any one of the preceding claims, wherein the cylinder has two ignition sources.

9. The engine of any one of the preceding claims, wherein the cylinder has a first ignition source centred within 10% of the radius of the cylinder from the central axis of the cylinder.

10. The engine of any one of the preceding claims, wherein the cylinder has a second ignition source centred away from the central axis of the cylinder, but within 90% of the radius of the cylinder from the central axis of the cylinder.

11. The engine of any one of claims 8 to 10, wherein a/the second of the ignition sources and the second port are located on the same side of the central axis of the cylinder.

12. The engine of any one of the preceding claims, having an ignition source comprising an in-cylinder injector for delivering a proportion of a cetane based diesel fuel relatively small in comparison to the gasoline fuel provided by the other injectors.

13. The engine of claim 12, wherein the in-cylinder injector is arranged to inject the diesel fuel late in a compression stroke of the piston.

14. The engine of claim 12 or claim 13, wherein the in-cylinder injector is arranged to inject the diesel fuel at least 150 crank angle degrees after bottom dead centre of an induction stroke of the piston.

15. The engine of any one of the preceding claims, having an ignition source comprising a spark plug with multi-strike capability.

16. The engine of any one of the preceding claims, wherein the first and second inlet valves are inclined relative to an exhaust valve of the cylinder by an angle greater than 15 degrees.

17. The engine of any one of the preceding claims, further comprising a supply of recirculated exhaust gas located upstream of the inlet ports.

18. The engine of claim 17, wherein the supply of recirculated exhaust gas is less than 70% by mass of carbon dioxide in the exhaust upstream of any catalyst.

19. The engine of claim 17 or claim 18, comprising means for splitting the supply of recirculated exhaust gas between the first inlet port and the second inlet port.

20. The engine of claim 19, wherein the splitting means can vary the split of recirculated exhaust gas between the first inlet port and the second inlet port.

21. The engine of any one of the preceding claims, comprising means for individually varying the injection of fuel by the injectors into each inlet port of the cylinder.

22. A method of operating the engine of any one of the preceding claims, comprising individually varying the injection of fuel by the injectors into each inlet port of the cylinder.

23. The method of claim 22, comprising varying the time at which a/the first injector injects fuel into the first port relative to the time at which a/the second injector injects fuel into the second port.

24. The method of claim 22 or claim 23, comprising varying the time at which a/the first injector injects fuel into the first port to be in advance of the time during a/the induction stroke of the piston at which a/the second injector injects fuel into the second port when flow through the first port is larger than flow through the second port.

25. The method of any one of claims 22 to 24, comprising varying the time at which a/the first injector injects fuel into the first port to be in advance of the time during a/the induction stroke of the piston at which a/the second injector injects fuel into the second port when flow of recirculated exhaust gas through the first port is larger than flow of recirculated exhaust gas through the second port.

26. The method of claim 22 or claim 23, comprising varying the time at which a/the first injector injects fuel into the first port to be between 0 and 140 Crankangle degrees after top dead centre of a/the induction stroke of the piston.

27. The method of claim 22 or claim 23, comprising varying the time at which a/the first injector injects fuel into the first port to be during an expansion or exhaust stroke of the piston and the time at which a/the second injector injects fuel into the second port to be during the induction stroke of the piston.

28. An internal combustion engine substantially as described with reference to any of the accompanying drawings.

29. A method of operating an engine substantially as described with reference to any of the accompanying drawings.