TECHNICAL FIELD

The present disclosure relates to a piston for an internal combustion engine. Aspects of the invention relate to a piston, to an engine comprising such a piston, to a control system and to a vehicle comprising such an engine or control system.

BACKGROUND

Conventional engines comprise a cylinder block and cylinder head, which together define one or more cylinders of the engine within which respective pistons undergo reciprocating motion. As is shown in Figure 1 , a piston 10, a cylinder block 12 and a cylinder head 14 of an engine together define a combustion chamber 16, within which air and fuel is mixed and the fuel combusted to force the piston 10 downwardly in a respective cylinder 18. This motion causes rotation of a crankshaft (not shown) of the engine, to provide motive power to a vehicle for example. A component of the air-fuel mixture may comprise exhaust gas recirculation gases, which have been recirculated back into the engine from an exhaust system of the vehicle.

As is well known, during combustion of this air-fuel mixture in the combustion chamber 16 a large amount of heat is generated. Some of this heat is transferred from the air-fuel mixture to the surfaces of the combustion chamber 16, including all exposed surfaces of the cylinder block 12 and cylinder head 14, and a compression face 20 of the piston. Typically, at partial engine load, approximately 30% of the total fuel power is lost as heat in the combustion chamber 16 in this way, where 60% of this loss occurs through the piston 10. At full engine load, approximately 17% of the total fuel power is lost as heat in the combustion chamber 16. This heat loss has the effect of reducing the thermal efficiency of the engine, and reduces the temperature of the gases leaving the combustion chamber 16 and entering the exhaust system of the vehicle. In turn, this increases the time taken for downstream aftertreatment devices of the exhaust system, such as a catalytic converter, to reach an operating temperature, which can impact the effectiveness of these aftertreatment devices in processing and treating the exhaust gases. In addition, the heat loss reduces an amount of energy available for driving an exhaust-driven turbocharger or exhaust energy recovery device of the vehicle. It is known to provide thermal coatings on the exposed surfaces of the combustion chamber 16 to try to mitigate against this heat loss. However, application of such thermal coatings can often be costly and time consuming, and the coatings may degrade over time, oxidising or peeling off from the surface to which they are applied. The application of such thermal coatings may also result in an increase in the temperature of the air-fuel mixture in the combustion chamber 16 prior to combustion, reducing the density of the air-fuel mixture and thereby reducing the volumetric efficiency of the engine.

It is an aim of the present invention to address one or more of the disadvantages associated with the prior art.

SUMMARY OF THE INVENTION

Aspects and embodiments of the invention provide a piston, an engine, a control system and a vehicle as claimed in the appended claims. The engine may operate either a compression-ignition cycle or a spark-ignition cycle.

According to an aspect of the present invention there is provided a piston for an internal combustion engine, the piston having a piston head. The piston head comprises: a crown having a compression face; and a side wall having an outer surface, the side wall depending from the crown. The compression face defines a concave formation having an outer perimeter which adjoins the outer surface of the piston head side wall, to define an acute angle between the concave formation and the outer surface of the piston head side wall.

The term‘adjoins’ referred to herein is intended to mean‘meets’ or‘abuts’, such that the concave formation has an outer perimeter which shares a common boundary, or boundary line, with the outer surface of the piston head side wall. This does not exclude a chamfered region or radiused region being present between the outer perimeter and the outer surface of the piston head side wall, in view of manufacturing tolerances for example, and the outer perimeter and outer surface may meet at such a chamfered or radiused region. The boundary line may take the form of a chamfered region or radiused region, for example. Defined another way, unlike in conventional piston arrangements, the concave formation of the piston does not comprise an outer rim at the outer perimeter of the concave formation, where this rim is substantially perpendicular to a central longitudinal axis of the piston head.

The acute angle may be defined between the concave formation and the outer surface of the piston head side wall in a plane in which a central longitudinal axis of the piston head lies. Defined another way, in a plane in which a central longitudinal axis of the piston lies, a tangent to the concave formation at the outer perimeter may intersect a tangent to the outer surface of the piston head side wall, defining an acute angle therebetween. Alternatively, or in addition, an acute angle may be defined between a central longitudinal axis of the piston and the tangent to the concave formation at the outer perimeter of the concave formation, at the point at which this central longitudinal axis and tangent intersect one another.

The acute angle may be between 45 and 75 degrees, and may be between 65 and 70 degrees for example.

A piston arrangement in which a compression face defines a concave formation, and in which an outer perimeter of this concave formation adjoins an outer surface of a piston head side wall, advantageously results in a combustion chamber that is bound by surfaces having a reduced total surface area in comparison to a conventional piston arrangement. Since heat generated during combustion of an air-fuel mixture within a combustion chamber of an engine is readily transferred to these surfaces, a reduction in this total surface area results in a corresponding reduction in an amount of heat lost from the combustion chamber. The term‘concave’ is intended to mean a formation that is generally dish-shaped, and is intended to include formations in which the compression face additionally extends vertically upwardly in one or more regions of the compression face.

It follows that a larger proportion of the heat generated during combustion is advantageously retained in the gases contained in the combustion chamber, permitting this heat to be transferred into mechanical work done by the gas mixture as combustion occurs. Additionally, or alternatively, this heat may be retained in the gas mixture as the mixture is guided through an exhaust system, warming any aftertreatment devices such that these reach their operating temperature more quickly and can provide energy to an exhaust-driven turbocharger or an exhaust recovery device.

A height difference between the outer perimeter and the compression face radially inward of the outer perimeter may be at a maximum at a centre of the compression face. The height difference may be taken in a direction parallel to a central longitudinal axis of the piston. The centre of the compression face may be the point on the compression face at which a central longitudinal axis of the piston intersects the compression face.

Optionally, a diameter of the outer perimeter is at a ratio of 8:1 to a height difference between the outer perimeter and a centre of the compression face. As such, the compression face may advantageously have a shallow formation, reducing the surface area of the compression face and, in turn, reducing heat loss to the compression face of the piston.

The concave formation may define a continuous curve across an entire diameter of the outer perimeter. Advantageously, the concave formation defining a continuous curve means that the concave formation does not comprise any substantially vertical wall portions, parallel to a central longitudinal axis of the piston, as is the case with conventional piston arrangements. As such, any impediment to the movement and mixing of fuel injected or directed towards the compression face of the piston is reduced relative to conventional piston arrangements. The length of the continuous curve is optionally substantially 3 to 7% longer than the diameter of the outer perimeter. Again, such a percentage difference advantageously results in a compression face having a substantially shallow formation.

Advantageously, the concave formation may comprise a trough region and a convex region that extends radially inwardly from the trough region towards a centre of the compression face. A convex region around the centre of the compression face beneficially results in a piston that promotes improved mixing of an air-fuel mixture in contact with the compression face. A height difference between the outer perimeter and the compression face radially inward of the outer perimeter may be at a maximum within the trough region. This maximum height difference may be from 5 to 10 mm, optionally from 6 to 8 mm and optionally 7 mm. This maximum height difference may be 5 to 1 1 % of a diameter of the outer perimeter, the height difference being in a direction parallel to a central longitudinal axis of the piston.

A height difference between the outer perimeter and the compression face radially inward of the outer perimeter may be 2 to 7% of a diameter of the outer perimeter, in the convex region at a centre of the compression face, the height difference being in a direction parallel to a central longitudinal axis of the piston. The height difference at the centre of the compression face may be from 2 to 6 mm, optionally from 3 to 5 mm and optionally 3.5 to 4 mm.

The concave formation may extend from the outer perimeter simultaneously both radially inwardly towards a central longitudinal axis of the piston and downwardly towards a skirt of the piston before extending simultaneously radially inwardly towards the central longitudinal axis and upwardly away from the skirt to a centre of the compression face where the central longitudinal axis intersects the compression face. In other words, the trough region may extend around the entire circumference of the concave formation about the convex region, the convex region being located at the centre of the compression face.

The piston may comprise a chamfered region between the outer perimeter of the concave formation and the outer surface of the piston head side wall. Such a chamfered region may increase the ease of manufacture of the piston and/or may advantageously result in a piston that is more hardwearing.

The chamfered region is optionally a chamfer having a width of less than 1 mm along a radial axis, perpendicular to a central longitudinal axis of the piston. The chamfered region may have a width of less than 0.5 mm along a radial axis. The piston may comprise a radiused region between the outer perimeter of the concave formation and the outer surface of the piston head side wall. According to another aspect of the present invention there is provided an engine comprising a piston in accordance with a previous aspect of the invention, the piston being movable in the cylinder, the engine further comprising a fuel injector configured to inject fuel into the cylinder.

The fuel injector may be configured to direct fuel substantially towards a centre of the compression face of the piston. Advantageously, directing fuel substantially towards the centre of the compression face improves mixing of an air-fuel mixture in contact with the compression face. This advantage may be particularly apparent in combination with a high injection angle, such as an injection angle of greater than 155 degrees.

The fuel injector may comprise a nozzle defining a plurality of fuel channels. Optionally, the plurality of fuel channels includes a central fuel channel aligned with a central longitudinal axis of the fuel injector. The plurality of fuel channels may be arranged to provide an injection angle of between 140 and 160 degrees and optionally, of between 135 to 155 degrees. An injection angle of 135 to 155 degrees advantageously provides optimum mixing of an air-fuel mixture in contact with the compression face. An injection angle above this reduces an amount of mixing of the air-fuel mixture towards a centre of the compression face, whereas an injection angle below this reduces an amount of mixing of the air-fuel mixture towards the outer perimeter of the concave formation.

The plurality of fuel channels may be arranged to provide an injection angle of between 150 and 160 degrees and optionally of between 155 and 157 degrees. Advantageously, improvements in mixing of an air-fuel mixture in contact with the compression face that result from the plurality of fuel channels of the nozzle including a central fuel channel are particularly apparent in combination with a high injection angle, such as an injection angle of greater than 155 degrees.

Each fuel injector is optionally configured to inject fuel at an injection pressure, or injection rail pressure, of 400 to 3200 bar, optionally of 1000 to 3000 bar, optionally of 1000 to 1800 bar. Injection pressures of 1000 to 3000 bar have been found to result in advantageous levels of mixing of an air-fuel mixture within the combustion chamber, resulting in efficient combustion. Injection pressures of between 1 100 and 1300 bar have been found to result in particularly high levels of mixing of an air-fuel mixture within the combustion chamber, with an injection pressure of 1 195 bar being found to be particularly beneficial.

Optionally, each fuel injector is configured to inject fuel at an injection pressure of 1000 to 3000 bar when the engine is operating at a speed of above 1750 rpm and/or at a load, or brake mean effective pressure (BMEP) of above 16 bar. Optionally, each fuel injector is configured to inject fuel at an injection pressure of 1000 to 3000 bar when the engine is operating at a speed of above 2500 rpm and/or at a load, or brake mean effective pressure (BMEP) of above 8 bar.

According to another aspect of the present invention, there is provided a control system comprising one or more controllers configured to control the fuel injector of an engine in accordance with a previous aspect of the invention, to inject fuel into the cylinder of the engine according to a split injection pattern. A split injection pattern refers to an injection pattern in which the injected fuel is split into two or more separate injections during a single cycle of the engine.

The split injection pattern may comprise three injections of fuel. Optionally, the split injection pattern may comprise between 5 and 8 injections of fuel. A split injection pattern advantageously results in improved mixing of an air-fuel mixture in the combustion chamber, enabling the generation of high local temperatures responsible for engine noise and NOx production to be controlled, and for engine noise and NOx production to be reduced as a result.

The split injection pattern may comprise a first fuel injection, the first fuel injection optionally being initiated while a crankshaft of the engine is at an angle of between -15 and +5 degrees relative to an orientation of the crankshaft when the respective piston is in a top dead centre position. The first fuel injection may be initiated while a crankshaft of the engine is at an angle of between -15 and +5 degrees when the engine is optionally operating at between l OOOrpm and 1750 rpm and/or at a load, or brake mean effective pressure (BMEP), of between 1 and 5 bar or, optionally, of between 3 to 5 bar. Such a delay in a first injection of the injection pattern relative to conventional injection strategies advantageously means that combustion begins when the piston is closer to the top dead centre position, allowing a greater proportion of energy from the combustion event to be transferred into positive mechanical work for generating engine torque.

Optionally, the first fuel injection is initiated while a crankshaft of the engine is at an angle of between -5 and +5 degrees relative to an orientation of the crankshaft when the respective position is in a top dead centre position.

According to another aspect of the present invention, there is provided a vehicle comprising a piston in accordance with a previous aspect of the invention, an engine according to a previous aspect of the invention, or a control system according to a previous aspect of the invention.

Within the scope of this application it is expressly intended that the various aspects, embodiments, examples and alternatives set out in the preceding paragraphs, in the claims and/or in the following description and drawings, and in particular the individual features thereof, may be taken independently or in any combination. That is, all embodiments and/or features of any embodiment can be combined in any way and/or combination, unless such features are incompatible. The applicant reserves the right to change any originally filed claim or file any new claim accordingly, including the right to amend any originally filed claim to depend from and/or incorporate any feature of any other claim although not originally claimed in that manner.

BRIEF DESCRIPTION OF THE DRAWINGS

One or more embodiments of the invention will now be described, by way of example only, with reference to the accompanying drawings, in which:

Figure 1 shows a section view of a cylinder of an internal combustion engine, the cylinder comprising a piston in accordance with the prior art;

Figure 2 shows a schematic plan view of a vehicle having an internal combustion engine in accordance with an embodiment of the invention; Figure 3 shows a section view of a cylinder of the internal combustion engine of Figure 2, the cylinder comprising an example of a piston in accordance with an embodiment of the invention;

Figure 4 shows a perspective view of a piston and crankshaft arrangement of the internal combustion engine of Figure 2;

Figure 5 shows a perspective view of the piston of Figure 3;

Figure 6a shows a section view of the piston of Figure 3, while Figures 6b and 6c show section views of variations on the piston of Figure 3;

Figure 7 shows a section view of a cylinder of the internal combustion engine of Figure 2, the cylinder comprising another example of a piston in accordance with an embodiment of the invention;

Figure 8a shows a section view of the piston of Figure 7, while Figures 8b and 8c show section views of variations on the piston of Figure 7;

Figure 9a shows a graph depicting a first example of a fuel injection strategy for use with the internal combustion engine of Figure 2;

Figure 9b shows a graph depicting a second example of a fuel injection strategy for use with the internal combustion engine of Figure 2;

Figure 10 shows a side view of a fuel injector of the internal combustion engine of Figure 2; and

Figures 1 1 a and 1 1 b show bottom views of variations on a nozzle arrangement of the fuel injector of Figure 10. DETAILED DESCRIPTION

An engine in accordance with an embodiment of the present invention will now be described with reference to the accompanying figures, along with two examples of a piston for use in such an engine. As will be explained in more detail below, the piston arrangements of embodiments of the invention are advantageously configured to reduce the total surface area of surfaces bounding each combustion chamber of the engine, so as to reduce the surface area available for heat transfer from the combustion chamber.

It will be appreciated that certain fundamental components of the engine and vehicle described herein will be common to both conventional arrangements and to the arrangements of embodiments the invention. As such, for ease of understanding, features of one example of the invention will be provided with reference numerals incorporating a prime, and differing features of another example of the invention will be provided with reference numerals incorporating two primes.

Figure 2 shows a schematic plan view of a vehicle 24’, comprising an internal combustion engine 22’ to provide motive power to the vehicle 24’. A driveline 26’ of the vehicle 24’ is connected to the engine 22’ by way of a transmission 28’, which may comprise one or more disconnect clutches (not shown) for facilitating connection and disconnection between the driveline 26’ and the engine 22’. The driveline 26’ is configured to transmit power from the engine 22’ to a front axle 30’ and/or a rear axle 32’ of the vehicle 24’, in order to drive a pair of front wheels 34’, 36’ and a pair of rear wheels 38’, 40’ of the vehicle 24’, respectively.

As is conventional, the engine 22’ comprises a cylinder block 12’ and a cylinder head 14’ (both shown in Figure 3) which together define a number of cylinders 18’. In the depicted embodiment, the engine 22’ comprises four cylinders 18’ arranged in an inline configuration, although it will be appreciated that any one of a number of cylinder arrangements may be employed. The vehicle 24’ additionally comprises a control system 42’, in the form of an engine control unit (ECU), which comprises a controller 44’ having an input 46’ and an output 48’. The controller 44’ is configured to control a pattern of fuel injection into each cylinder 18’. Figure 2 also depicts an exhaust system 50’ of the vehicle 24’. The engine 22’ is fluidly connected to an exhaust manifold 52’, in which any exhaust gases expelled from the engine 22’ are collected before passing through downstream components of the exhaust system 50’. Such downstream components include one or more aftertreatment devices 54’, which may include a catalytic converter assembly and/or an SCR unit. These aftertreatment devices 54’ ensure that the exhaust gases are treated prior to their discharge into the atmosphere via one or more tail pipes 56’. Typically, the exhaust system 50’ will also comprise one or more mufflers, or silencers (not shown), to reduce noise associated with the exhaust gases as they travel through the system and exit the vehicle 24’.

The engine 22’ comprises an equal number of pistons 10’ to the number of cylinders 18’, with each piston 10’ being slidably received within a respective cylinder 18’. Referring to Figures 3 and 4, each piston 10’ and cylinder 18’ pair together defines a combustion chamber 16’ within which fuel and air are combined and the fuel combusted to force the piston 10’ downwardly in the cylinder 18’. The engine 22’ additionally comprises a crankshaft 58’, driven by movement of each of the pistons 10’ via a respective connecting rod 60’ that extends therebetween. As will become apparent from the following description, each piston 10’ is advantageously configured such that heat transfer from the gas mixture within the combustion chamber 16’ to the piston 10’ during combustion is minimised, permitting this heat to be used for other means within the vehicle 24’.

Figure 3 additionally illustrates the positions of an intake port 62’, an exhaust port 64’ and associated intake and exhaust valves 66’, 68’ for each combustion chamber 16’. Actuation of the intake and exhaust valves 66’, 68’ can be controlled to control the flow of air into the combustion chamber 16’ and the flow of exhaust gases out of the combustion chamber 16’ respectively. Exhaust gases expelled through the exhaust port 64’ are collected in the exhaust manifold 52’ (shown in Figure 2) and are guided through the vehicle 24’ exhaust system 50’ as previously described.

As is well-known, the cylinder head 14’ of the engine 22’ also houses a fuel injector 70’, for injecting fuel into the combustion chamber 16’ of a respective cylinder 18’. The fuel injector 70’ is arranged such that a central longitudinal axis C of the cylinder 18’, a central longitudinal axis P’ (shown in Figure 5) of the piston 10’ and a central longitudinal axis F (shown in Figure 10) of the fuel injector 70’ are aligned and overlapping. While each of these axes is indicated in a separate figure, it will be appreciated that these axes therefore coincide during use of the engine 22’. Alternative arrangements are also envisaged in which the central longitudinal axis F of the fuel injector 70’ and the central longitudinal axis P’ of the piston 10’ are offset from one another, and do not coincide. The vehicle ECU 42’ (shown in Figure 2) is connected to each fuel injector 70’ of the vehicle 24’, transmitting control signals to the injectors 70’ to control both the timing, injection pressure and the length of fuel injections into each of the combustion chambers 16’ during the cycle of the engine 22’.

A single cycle requires two full rotations of the crankshaft 58’. For any one cylinder 18’, the first rotation of the crankshaft 58’ is made up of an intake stroke and a compression stroke. In the intake stroke, the piston 10’ descends within the cylinder 18’ towards a ‘bottom dead centre’ (BDC) position, the intake valve 66’ is open and air is drawn into the combustion chamber 16’. In the compression stroke the piston 10’ ascends towards a‘top dead centre’ (TDC) position with both valves 66’, 68’ closed, and the air in the combustion chamber 16’ is compressed. The second rotation of the crankshaft 58’ then comprises a power stroke and an exhaust stroke. At the end of the compression stroke and continuing into the start of the power stroke, fuel is injected into the combustion chamber 16’ by way of the fuel injector 70’, such that air and fuel mix in the combustion chamber 16’. A component of the air-fuel mixture comprises exhaust gas recirculation gases, which have been recirculated back into the engine from an exhaust system of the vehicle. As a result of a spark ignition this fuel ignites, or alternatively as a result of the high temperature and pressure within the combustion chamber 16’, the fuel vaporises and ignites. The piston 10’ is driven back down to the BDC position, before the exhaust valve 68’ is opened and the piston 10’ ascends to the TDC position again to force the combustion products from the combustion chamber 16’.

Referring to Figures 5 and 6a-c, an example of a piston 10’ according to the present invention will now be described in more detail. The piston 10’ will be described with reference to the central longitudinal axis P’ of the piston 10’. Reference to‘vertically above’ or‘upper’ and‘vertically below’ or‘lower’ throughout the following description should be understood to mean disposed upwardly along the longitudinal axis P’ or disposed downwardly along the longitudinal axis P’ of the piston 10’ respectively, with the longitudinal axis P’ in the orientation shown in Figures 5 and 6a-c. Further, reference to‘radially outwardly’ or‘radially inwardly’ should be understood to mean disposed away from the central longitudinal axis P’ or towards the central longitudinal axis P’ respectively, along a line intersecting and perpendicular to the central longitudinal axis P’.

As is conventional, the piston 10’ comprises a piston head 72’ and a skirt 74’, which is connected to and extends vertically below the piston head 72’. The skirt 74’ is configured to assist with alignment of the piston 10’ within the cylinder 18’ (shown in Figure 3), preventing rotation of the piston 10’ relative to the cylinder 18’ as the piston 10’ reciprocates within the cylinder 18’. The piston head 72’ includes an uppermost portion 76’ in the form of a crown, and a side wall 78’ depending from the crown 76’ and having an outer surface 80’ and an inner surface 82’. This outer surface 80’ and inner surface 82’ face radially outwardly from the head 72’ and radially inwardly from the piston head 72’, respectively. In this way, the side wall 78’ depends from the crown 76’, extending generally vertically between the crown 76’ and the skirt 74’.

To allow relative movement between the piston 10’ and the cylinder 18’ there is a clearance between the two. The piston head 72’ is exposed to the heat of the burning gases during combustion in the combustion chamber 16’ (shown in Figure 3) and, as such, absorbs a significant amount of heat from these gases, tending to expand during operation of the engine 22’. The side wall 78’ may therefore be slightly angled, such that a clearance between the cylinder 18’ and the piston 10’ is larger in the region of the piston head 72’ than the skirt 74’. To guard against any leakage of gas from the combustion chamber 16’ around the outside of the piston 10’ head and skirt 74’ of the piston 10’, the piston 10’ head is additionally provided with ring grooves 84’ spaced vertically apart from one another and extending around a perimeter of the piston head 72’ and configured to receive piston rings (not shown) for sealing purposes.

An upper surface of the crown 76’ forms a compression face 20’, which directly contacts the air-fuel mixture within the combustion chamber 16’ during operation of the engine 22’. Referring briefly to Figure 3, the combustion chamber 16’ is therefore bounded by the compression face 20’, an internal surface 86’ of the cylinder head 14’ and an internal surface 88’ of the cylinder block 12’ extending between the compression face 20’ and the cylinder head 14’. Since the compression face 20’ of the piston 10’ moves relative to the cylinder head 14’, the size of the combustion chamber 16’ varies continuously throughout the compression-ignition cycle of the engine 22’. The central longitudinal axis P’ of the piston intersects the compression face 20’ at a centre X’ of the compression face 20’, shown in Figure 5.

Referring to Figure 1 , a conventional piston 10 for use in an engine operating a compression-ignition cycle is depicted to provide context for the present invention, and as means for comparison. A piston for use in an engine operating a spark ignition cycle may also be considered. This conventional piston 10 similarly comprises a crown 76 having an upper surface in the form of a compression face 20. The compression face 20 is shaped so as to form a‘bowl’ region 90, surrounded by an outer rim 92. This outer rim 92 is substantially perpendicular to a central longitudinal axis L of the piston 10, forming a horizontal annulus at the uppermost part of the piston 10.

The purpose of this bowl region 90 in the conventional piston 10 is to promote mixing between the air and the fuel within the combustion chamber 16, to aid combustion. To form the bowl 90, the compression face 20 extends downwardly from the outer rim 92 and slightly outwardly, such that the outer rim 92 appears to overhang the bowl 90. From this point the bowl 90 extends first radially inwardly and then both radially inwardly and vertically upwardly to a centre of the compression face 20, where the central longitudinal axis L of the conventional piston 10 intersects the compression face 20. In this way, the bowl 90 appears to comprise a central dome, positioned radially inwardly of an annular trough region. An outer diameter of the bowl region 90 at the uppermost part of the piston 10 is approximately 48 mm, while a diameter of the cylinder 18 is 83 mm.

As can be seen from Figure 1 , an annular channel 94 extends through the piston head 14 of the conventional piston 10, positioned slightly radially inwardly of an outer surface 80 of the piston head side wall 78. This annular channel 94 forms a cooling gallery, within which oil is circulated to cool the piston head 14 during operation of the engine 22. In contrast to the conventional piston 10, and as is apparent from Figures 5 and 6a-c, the piston 10’ of this embodiment of the invention has a compression face 20’ that is relatively shallow and dish-shaped, defining a substantially concave formation 96’ across its entirety. Such a design may be deployed in a compression-ignition or a spark- ignition engine. This concave formation 96’ has an outer perimeter 98’ which adjoins the outer surface 80’ of the piston 10’ head side wall, such that the concave formation 96’ of the piston 10’ extends from the outer perimeter 98’ of the concave formation 96’ both radially inwardly towards the central longitudinal axis P’ of the piston 10’ and downwardly towards the skirt 74’ of the piston 10’. In contrast to the conventional piston 10, the piston 10’ of the described embodiment of the invention therefore comprises no outer rim 92, since the concave formation 96’ abuts, or meets, the outer surface 80’ of the piston head side wall 78’ around the entire perimeter of the concave formation 96’. Expressed another way, the outer surface 80’ of the piston head side wall 78’ and the outer perimeter 98’ of the concave formation 96’ share, or meet at, a mutual boundary line or interface.

Due to the relative position and orientation of the outer surface 80’ of the piston head side wall 78’ and the concave formation 96’, it will be appreciated that the compression face 20’ and outer surface 80’ of the side wall 78’ meet at a mutual boundary line to define an acute angle therebetween, marked Ό’ in Figure 6a. The acute angle Q is between 45 and 75 degrees, and may be 67 degrees for example. Defined another way, in a plane positioned along and parallel to the central longitudinal axis P’ of the piston 10’, a tangent T i (shown in Figure 6a) to the concave formation 96’ at the outer perimeter 98’ extends towards and intersects a tangent T ₂ (shown in Figure 6a) to a substantially linear or vertical portion of the outer surface 80’ of the piston head side wall 78’, defining an acute angle Q between the tangents Ti , T ₂ . In addition, the tangent Ti to the concave formation 96’ at the outer perimeter 98’ extends towards and intersects the central longitudinal axis P’ of the piston 10’ to define the acute angle Q therebetween.

The mutual boundary line may comprise a chamfered region 100’, for ease of manufacture, where the outer perimeter 98’ of the concave formation 96’ adjoins this chamfered region 100’. While a‘chamfer’ typically forms an angle of approximately 45 degrees with each of the surfaces it adjoins, it is envisaged that a chamfer of any suitable angle may be employed. As an alternative, the mutual boundary line may comprise a radiused region, where it is the radiused region that adjoins the outer perimeter 98’ of the concave formation 96’. In either case, this chamfered or radiused region 100’ has a width of less than 1 mm along a radial axis R’, perpendicular to the central longitudinal axis P’ of the piston 10’. In particular, the chamfered or radiused region 100’ can have a width of approximately 0.5 mm along this radial axis R’. Alternatively, it is envisaged that the piston 10’ may not comprise either a chamfered region or a radiused region 100’.

In the piston arrangement depicted in Figures 6a-c, it is readily apparent that a height difference between the outer perimeter 98’ and the remaining portion of the compression face 20’ radially inwards from the outer perimeter 98’ is at a maximum at a centre X’ of the compression face 20’. In particular, the height difference at the centre X’ of the compression face 20’ is at a ratio of 1 :8 to a diameter of the outer perimeter 98’, where the height difference is taken in a direction parallel to the central longitudinal axis P of the piston 10’. To define this in another way, the concave formation 96’ can be described as defining a continuous curve across an entire diameter of the outer perimeter 98’, where the length of the continuous curve is approximately 3 to 7% longer than the diameter of the outer perimeter 98’. In particular, the length of the continuous curve may be approximately 85.3 mm.

Referring now to Figures 7 and 8a-c, an alternative piston 10” according to the present invention will now be described. As with the piston arrangement of Figures 5 and 6a-c, the piston arrangement of Figures 7 and 8a-c similarly comprises a compression face 20” that is relatively shallow and dish-shaped, defining a generally concave formation 96” across its entirely. Again, an outer perimeter 98” of this concave formation 96” adjoins an outer surface 80” of a side wall 78” of the depicted piston head 72”, such that the piston 10” comprises no outer rim portion and the concave formation 96” abuts the outer surface 80” of the piston head 72” around the entire perimeter of the concave formation 96”. The concave formation 96” may also adjoin a chamfered or radiused region 100” of the outer surface 80” of the piston head side wall 78”, in the same way as previously described.

The concave formation 96” of the alternative piston arrangement extends from the outer perimeter 98” of the concave formation 96” simultaneously both radially inwardly towards a central longitudinal axis P” of the piston 10” and downwardly towards the skirt 74” of the piston 10”. Unlike the piston arrangement of Figures 5 and 6a-c, from this point, the concave formation 96” extends simultaneously radially inwardly towards the central longitudinal axis P” and upwardly away from the skirt 74” of the piston 10” to a centre X” of the compression face 20’, where the central longitudinal axis P” of the piston 10” intersects the compression face 20”. In this way, the concave formation 96” appears to comprise a central convex region 102”, or central dome, positioned radially inwardly of a shallow annular trough region 104”. When the engine 22’ is in operation, this central dome 102” advantageously promotes mixing of the air-fuel mixture in the combustion chamber 16’, forcing air radially outwardly towards the jets of fuel injected into the cylinder 18’.

In the same way as described in respect of the piston 10’ of Figures 5 and 6a-c, the concave formation 96” and outer surface 80” of the side wall 78” of the piston 10” together define an acute angle Q. In addition, a tangent T to the concave formation 96” at the outer perimeter 98” extends towards and intersects a central longitudinal axis P” of the piston 10” to define an acute angle Q therebetween. The acute angle Q is between 45 and 75 degrees, and may be 67 degrees for example.

As is apparent from Figures 8a-c, a height difference between the outer perimeter 98” and the remaining portion of the compression face 20”, taken in a direction parallel to the central longitudinal axis P” of the piston 10”, is at a maximum within the trough region 104”. This maximum height difference is approximately 5 to 1 1 % of the diameter of the outer perimeter 98” and, in particular, may be approximately 7 mm. Further, a height difference between the outer perimeter 98” and the remaining portion of the compression face 20” is approximately 2 to 7% of the diameter of the outer perimeter 98” in the convex region 102”, at a centre X” of the compression face 20”. In particular, the height difference at the centre X” of the compression face 20” may be 3.8 mm. The remaining portion of the compression face 20” is the compression face 20” radially inward of the outer perimeter 98”.

The concave formation 96” can again be described as defining a continuous curve across an entire diameter of the outer perimeter 98”, where the length of the continuous curve is approximately 3 to 7% longer than the diameter of the outer perimeter 98”. In particular, the length of the continuous curve may be approximately 86.3 mm In contrast to the conventional piston 10, the outer perimeter 98” of the concave formation 96’, 96” of each of the described embodiments of pistons 10’, 10” of the invention has a nominal diameter of approximately 83 mm, for a typical cylinder 18’ diameter of 83 mm, while allowing for a sliding tolerance between the piston 10’, 10” and the cylinder 18’. Therefore, the ratio of the outer diameter 98’, 98” of the concave formations 96’, 96” of the pistons 10’, 10” depicted in Figures 5 to 8c to the diameter of the cylinder 18’ is much larger than the ratio of the outer diameter of the bowl region 90 of the conventional piston 10 to the diameter of the cylinder 18.

The profiles of the compression faces 20’, 20” of the pistons 10’, 10” of the described embodiments have a number of advantages.

Firstly, the relatively large diameter of the concave formation 96’, 96” of the compression faces 20’, 20” requires fuel to travel a greater distance from the fuel injector 70’ to the compression face 20’, 20” relative to the conventional piston 10.

The conventional piston 10 comprises a wall portion around the outer circumference of the bowl 90, the effect of which is to re-direct the fuel as it is sprayed into the combustion chamber 16, interfering with the injection spray plume. As such, the piston arrangements of the described embodiments of the invention each permit fuel to more easily and rapidly disperse along the compression face 20’, 20”, such that the fuel reacts with the air across a greater surface area of the compression face 20’, 20”, promoting faster and more efficient combustion.

In addition, the pistons 10’, 10” of the described embodiments are capable of being positioned much closer to the opposing surface of the cylinder head 14’ when the pistons 10’, 10” are in the TDC position. In more detail, since portions of the intake and exhaust valve assemblies protrude into the combustion chamber 16 during operation of the engine 22, the conventional piston arrangement requires that the piston 10 is vertically spaced from the cylinder head 14 at TDC, to prevent the outer rim 92 of the piston compression face 20 from coming into contact with the valves 66, 68 during the cycle of the engine 22. In the piston arrangements of the described embodiments of the invention, the concave formation 96’, 96” extends all the way to the outer surface 80’, 80” of the piston head side wall 78’, 78”, and so the compression face 20’, 20” is slightly set back in the region of the valves 66’, 68’. The vertical spacing between the cylinder head 14’ and the uppermost part of the pistons 10’, 10” can therefore be reduced at TDC relative to the conventional piston 10.

Since the piston 10’, 10” can be positioned closer to the cylinder head 14’ at TDC, the total surface area bounding the combustion chamber 16’ is reduced. This is because the internal surface of the cylinder block 12’ extending between the compression face 20’, 20” of the piston 10’, 10” and the cylinder head 14’ is minimised when the piston 10’, 10” is at TDC. As can be seen from Figures 1 and 5 to 8c, this surface area is additionally reduced by virtue of the shallow profile of the piston compression faces 20’, 20”. Since heat generated during combustion of the air-fuel mixture within the combustion chamber 16’ is readily transferred to the surfaces bounding the combustion chamber 16’, a reduction in this total surface area results in a corresponding reduction in an amount of heat lost from the combustion chamber 16’.

It follows that a larger portion of the heat generated during each combustion event is advantageously retained in the gases contained in the combustion chamber 16’, to be used for other means. For example, any residual heat in the air-fuel mixture during combustion can be used to increase the rate of reaction in the combustion chamber 16’ and to expand the gases contained therein. In this way, the residual heat can be transferred into mechanical work done by the gas mixture in forcing the piston 10’, 10” downwardly in the cylinder 18’ as combustion occurs, and driving the crankshaft 58’ of the engine 22’. Alternatively, or in addition, this heat can be retained in the exhaust gas mixture as the mixture is expelled from the engine 22’ and is guided through the exhaust system 50’. In this way, the heat is retained in the exhaust gas mixture as it reaches the aftertreatment devices 54’, warming the catalytic converter to promote the chemical reactions taking place therein. The rate of reactions occurring in the catalytic converter can therefore be increased, and the exhaust gases are treated more efficiently and effectively prior to exiting the vehicle 24’ through the tail pipe 56’.

The increased distance travelled by the fuel from the fuel injector 70’ to the compression face 20’, 20” means that undesirable concentrations of heat at the compression face 20’, 20” are reduced. As such, the temperature of the compression face 20’, 20” is more uniform, and the piston 10’, 10” is, as a whole, able to increase further in temperature relative to conventional piston arrangements prior to any part of the piston reaching a threshold operating temperature for the piston material. The average piston temperature can therefore be higher than for the conventional piston 10, resulting in faster burning of fuel in the combustion chamber 16’, more efficient combustion and reduced NOx production. Further, since a higher average piston temperature results in a reduced temperature gradient between the air-fuel mixture of the combustion chamber 16’ and the compression face 20’, 20” of the piston 10’, 10”, heat transfer between the combustion chamber 16’ and the compression face 20’, 20” is reduced.

Since heat distribution is more uniform across the compression face 20’, 20” the cooling requirements of the engine components defining the combustion chamber 16’ are reduced. The oil cooling requirements for each piston 10’, 10” are therefore also reduced. This in turn means that smaller oil pumps can be used, resulting in reduced parasitic losses of energy from the engine. Further, in the event that the more even distribution of heat across each piston 10’, 10” means that the pistons 10’, 10” do not exceed a threshold operating temperature, the oil cooling gallery 94, 94’, 94” of the conventional piston arrangement, and present in the described embodiments of the invention of Figures 6a and 8a may no longer be required at all. In this case, the oil cooling gallery 94, 94’, 94” can be removed and may be substituted with a jet to cool the piston. This advantageously results in greater fuel economy for the vehicle 24’.

Referring to Figures 6b and 6c, slight variations on the piston 10’ of Figure 6a are depicted, while Figures 8b and 8c similarly depict variations on the piston 10” of Figure 8a. In particular, Figures 6b and 8b depict examples of each of the described piston arrangements without the respective oil cooling gallery 94’, 94”. In these arrangements, the piston head 72b’, 72b” of the pistons 10b’, 10b” is a solid part formed without any internal channels therein. Figures 6c and 8c show further modified arrangements in which the oil cooling gallery 94’, 94” has been removed and the cross sectional area of the pistons 10c’, 10c” has been reduced. Specifically, a hollow 106c’, 106c” in the piston 10c’, 10c” extends upwardly into the piston head 72c’, 72c”, leaving a relatively thin layer of material beneath the compression face 20c’, 20c” of the piston 10c’, 10c”. The effect of this is to create pistons 10c’, 10c” having the same compression face profile as described in relation to Figures 6a and 8a but with reduced mass. The mass of the pistons 10c’, 10c” of Figures 6c and 8c is therefore reduced relative to the pistons 10’, 10” of Figures 6a and 8a, reducing the raw materials requirements associated with the manufacture of the pistons 10c’, 10c”. In addition, this reduction in mass of the pistons 10c’, 10c” results in reduced vibration of the pistons 10c’, 10c” during operation of the engine. This means that less mass is required within the engine 22’ to act as a counterbalance to the mass of the pistons 10c’, 10c”, resulting in improved fuel economy of the vehicle 24’.

In addition to permitting greater flexibility in piston design, the reduction in cooling requirements associated with the piston arrangements of the described embodiments of the invention means that manufacture of the pistons 10b’, 10b”, 10c’, 10c” is less complex and is cheaper. Further, there is no need to rely on the thermal coatings of the prior art to mitigate against heat loss to the surfaces of the combustion chamber 16’. Since a well-established problem with these coatings is that they degrade over time, the performance of the engine 22’ is more consistent, and can therefore be predicted more readily, without their use.

Two injection strategies for use with the piston arrangements of the described embodiments will now be described, with reference to Figures 9a and 9b. It will be appreciated that the described injection strategies are two examples of a number of injection strategies that may be used in conjunction with the described piston arrangements, and that the injection strategy employed at any one time may depend on a number of operating conditions of the engine 22’.

The described injection strategies have multiple associated advantages, particularly when combined with the piston of the illustrated embodiments, including an improvement in mixing of the air-fuel mixture within the combustion chamber 16’ resulting in more efficient combustion, and a reduction in NOx production.

The controller 44’ of the ECU 42’ is configured to communicate with each fuel injector 70’ of the engine 22’ to control the pattern of fuel injection into the respective cylinder 18’. In particular, the output of the controller 44’ transmits control signals to any given fuel injector 70’ to activate the injector 70’ at the time at which an injection of fuel is required. Figure 9a is a graph depicting two fuel injection strategies for use when the engine 22’ meets a first set of operating conditions: the first line, indicated by the darker line A, is a conventional fuel injection strategy, used in an engine 22 in which the conventional pistons 10 are employed; and the second, indicated by the lighter line B, is an injection strategy for use with the piston arrangements of the described embodiments of the invention. The x-axis of the graph indicates a crank angle of the crankshaft 58, 58’, where a crank angle of 0 degrees represents the orientation of the crankshaft 58, 58’ when the piston 10, 10’, 10” is in a TDC position. A negative angle on the graph corresponds to the same piston 10, 10’, 10” ascending to the TDC position during the compression stroke of the engine 22, 22’, while a positive angle on the graph corresponds to the piston 10, 10’, 10” descending away from the TDC position during the power stroke of the engine 22, 22’.

As can be seen from Figure 9a, the conventional injection pattern A comprises a pilot fuel injection, delivered while the crank angle is between approximately -15 degrees and -12 degrees. A main fuel injection follows, while the crank angle is between approximately -2 degrees and 12 degrees. This main injection is longer than the pilot injection, and also involves injecting fuel at a higher rate. In particular, in the pilot injection, fuel is injected at a maximum flow rate of approximately 0.0100 kg/s, while in the main injection fuel is injected at a maximum flow rate of approximately 0.0275 kg/s.

In contrast, when executing the injection strategy B of the described embodiment of the invention, depicted by the lighter line, the controller 44’ is configured to delay the pilot injection, transmitting a signal to the fuel injector 70’ to cause the fuel injector 70’ to deliver the pilot fuel when the crank angle is between approximately -2 degrees and 2 degrees. The injection pressure, or injection rail pressure, and therefore the maximum flow rate of this injection is additionally increased relative to the conventional injection pattern A, consequently increasing the overall mass of fuel injected into the cylinder 18’ during the pilot injection. The maximum flow rate during the pilot injection is controlled by the controller to be just below 0.0300 kg/s, at a crank angle of approximately 1 degree.

Advantageously, increasing the injection pressure and the quantity of fuel delivered during this injection has the effect that the initial phase of combustion of the injected fuel is faster when compared to the conventional fuel injection strategy A used with the conventional piston 10. This allows for combustion to occur more readily, which itself permits the pilot fuel injection to be delayed for the injection strategy B of the invention relative to the conventional injection strategy A. Delaying this pilot injection means in turn that combustion begins when the piston 10’, 10” is closer to the TDC position in the compression stroke, such that a greater proportion of energy from the combustion event can be transferred into positive mechanical work for driving the piston 10’, 10” downwardly in the cylinder 18’ during the power stroke.

In addition, delaying a pilot injection may have the effect that peak and overall pressures in the combustion chamber 16’ are reduced for a given quantity of fuel injected, since combustion is delayed and the period of time within which combustion occurs is shortened. As a consequence, the durability of the engine 22’ may be improved.

Following the pilot injection, the controller 44’ is configured to transmit two separate signals to the fuel injector 70’, causing the fuel injector 70’ to deliver two‘main’ injections. The first of these main injections is delivered when the crank angle is between approximately 5 degrees and 12 degrees, with a maximum flow rate of approximately 0.0375 kg/s. The second of these main injections is relatively short, occurring between a crank angle of approximately 16 degrees and 18 degrees, with a relatively low maximum mass flow of approximately 0.0175 kg/s. In this way the controller 44’ is configured to instruct the execution of a split injection pattern comprising three injections: a first pilot injection, and first and second main injections, the second main injection being injected over a smaller crank angle range and at a lower mass flow rate than the first.

The delay in the pilot injection results in more aggressive combustion in the combustion chamber 16’. Introducing a greater number of subsequent main fuel injections results in improved mixing in the combustion chamber 16’ and enables the generation of high local temperatures responsible for engine noise and NOx production to be controlled. With an increase in the number of injection events, the NOx production is controlled to be lower than that for the conventional injection strategy A when this strategy A is employed with the conventional pistons 10.

In addition, as a result of the later injections in the injection strategy B of the embodiment of the invention and the reduced heat loss to the surfaces of the combustion chamber 16’ the temperature of the exhaust gases expelled from the combustion chamber 16’ is higher, increasing the efficiency with which the emissions are processed in the exhaust gas aftertreatment devices. Further, when the engine 22’ is started from an off state and is cold, this increase in exhaust gas temperature allows the aftertreatment devices to warm to operating temperature more quickly.

Figure 9b is a graph depicting two fuel injection strategies for use when the engine 22’ meets a second set of operating conditions. Again, the first line, indicated by the darker line A’, is a conventional fuel injection strategy, used in an engine 22 in which the conventional pistons 10 are employed. The second line, indicated by the lighter line B’, is an injection strategy for use with the pistons 10’, 10” of the described embodiments on the invention.

Similarly to the injection strategy B described above, employed when the engine 22’ meets the first set of operating conditions, the injection strategy B’ of Figure 9b also comprises a delayed pilot injection and a greater number of subsequent main injections relative to the corresponding conventional injection strategy A’. Again, delaying the pilot fuel injection has the effect that combustion begins when the piston 10’, 10” is closer to the TDC position in the compression stroke, allowing a greater proportion of energy from the combustion event to be transferred into positive mechanical work for generating engine torque. Further, delaying the main injections and splitting these into a greater number of injections assists in controlling and reducing engine noise and NOx production during operation of the engine 22’.

The temperature within the combustion chamber 16’ can additionally be controlled by managing an amount of exhaust gas recirculation. Recirculating a portion of the exhaust gases from the engine 22’ back into the intake port 62’ has the effect that the exhaust gases replace a portion of the oxygen drawn into the combustion chamber 16’, and additionally absorb a proportion of the heat generated during combustion. This has the combined effect of reducing the temperature of the combustion chamber 16’, reducing NOx production. The ability to control parameters of the fuel injections such as their timing and duration has been described. In addition, it is envisaged that the angle of these injections may be controlled by modifying characteristics of the fuel injector 70’ itself, as will now be explained.

Referring to Figure 10, a side view of a fuel injector 70’ in accordance with the present invention is shown. As is conventional, the fuel injector 70’ comprises an elongated main body 108’, terminating at a nozzle 1 10’ through which fuel is expelled. The fuel injector 70’ is connected to the controller 44’ of the ECU 42’ by way of one or more electrical wires 1 12’ to enable the ECU 42’ to activate the injector 70’ and control the timing of fuel injections into the respective cylinder 18’. The fuel injector 70’ has a central longitudinal axis, labelled F in Figure 10.

Referring now to Figures 1 1 a and 1 1 b, two alternative nozzle arrangements for the fuel injector 70’ are shown. In particular, these figures show alternative arrangements for a lower end of the nozzle 1 10a’, 1 10b’ furthest from the fuel injector main body 108’. The figures depict this lower end as viewed along the longitudinal axis F of the fuel injector 70’ in a direction towards the main body 108’.

The nozzles 1 10a’, 1 10b’ each comprise a substantially solid shaft, through which a number of elongated channels 1 14a’, 1 14b’ extend for delivering fuel to the cylinder 18’. In the nozzle 1 10a’ of Figure 1 1 a, eight such fuel channels 1 14a’ are provided, defining spray holes 1 16a’ positioned in an annular arrangement in the lower face of the nozzle 1 10a’. The fuel injector 70’ of Figure 1 1 b similarly comprises eight fuel channels 1 14b’ defining spray holes in the lower face of the nozzle 1 10b’. Flowever, in the arrangement of 1 1 b only seven of these fuel channels 1 14b’ are arranged in an annular arrangement, with the eighth fuel channel 1 14b’ extending along the central longitudinal axis F of the fuel injector 70’ to form an eighth spray hole 1 16b’ at the centre of the lower face of the nozzle 1 10b’.

The spread of fuel injected into the cylinder 18’ is determined by the orientation and arrangement of the fuel channels 1 14a’, 1 14b’ within the injector 70’. In particular, the fuel channels 1 14a’, 1 14b’ together define a fuel cone angle, or‘injection angle’: the angle formed between outermost tangents to the fuel spray on opposing sides of the fuel injector 70’. With the spray holes 1 16a’, 1 16b’ arranged as depicted in Figure 1 1 a it has been found that an injection angle of between 135 and 155 degrees is advantageous. For injection angles above 155 degrees mixing of the air-fuel mixture at the centre of the compression face 20’, 20” is reduced, whereas injection angles below 135 degrees result in reduced mixing of the air-fuel mixture towards the outer perimeter 98’, 98”. Injection angles in the range of 135 and 155 degrees optimise mixing in the combustion chamber 16’ by providing a trade-off between these regions of mixing, improving mixing in the combustion chamber 16’ relative to arrangements in which the injection angle is above or below this range. With the spray holes 1 16a’, 1 16b’ arranged as depicted in Figure 1 1 a it has been found that an injection angle of 140 degrees is particularly advantageous for mixing.

An even greater improvement in mixing is experienced when the fuel channels 1 14a’, 1 14b’ are arranged as shown in Figure 1 1 b, and such that the injection angle is approximately 156 degrees. In fact, it has been found that for injection angles of above 155 degrees, the inclusion of a spray hole 1 16a’, 1 16b’ at the centre of the nozzle 1 10a’, 1 10b’ always results in an improvement in a level of mixing of the air-fuel mixture in the combustion chamber 16’.

The injection strategies B, B’ fuel injector nozzle arrangements and piston arrangements described herein therefore together provide improved mixing of the gases within the combustion chamber 16’. The combination of injection angle and a shallow, concave compression face 20’, 20” has the effect that fuel injected into the cylinder 18’ is relatively unimpeded, compared to conventional arrangements, and is free to disperse within a larger region of the combustion chamber 16’ prior to coming into contact with any surfaces. Improvements in mixing of the air-fuel mixture result in faster reaction between the fuel and the gas and therefore more efficient combustion, along with reduced NOx production in the combustion chamber 16’.

Mixing can additionally be enhanced through use of high injection pressures. For example, injection pressures of between approximately 1 100 and 1300 bar have been found to result in high levels of mixing within the combustion chamber 16’, with an injection pressure of approximately 1 195 bar being found to be particularly advantageous. The use of high injection pressures also permits the use of injection pressure as the primary means for controlling and influencing mixing levels, rather than controlling mixing through piston design as is conventional.

As such, the described injection strategy B and fuel injector nozzle arrangements reduce the requirement for a‘bowl’ region 90 in the compression face 20’, 20” of the pistons 10’, 10” as is seen in conventional piston arrangements. The lack of bowl 90 and the shallower, substantially concave profiles of the piston arrangements of the described embodiments of the invention results in reduced surface area of the combustion chamber 16’, and consequently reduced heat loss from the air-fuel mixture to the surfaces of the combustion chamber 16’. This heat can therefore be used either in the engine 22’, being transferred into mechanical work by the expanding gases in the cylinder 18’ during the power stroke, or in the exhaust system 50’, to warm the aftertreatment devices 54’ for more efficient treatment of the exhaust gases. Alternatively, or additionally, this heat can be used in an exhaust-driven turbocharger or exhaust energy recovery system. The described piston arrangements and injection strategy B therefore together result in faster and more efficient combustion, as well as reduced production of NOx gases during the combustion process.

The elimination of the bowl region 90 of the compression face 20’, 20” additionally results in a piston 10’, 10” having a much simpler compression face profile. The compression faces 20’, 20” of the pistons 10’, 10” of the described embodiments are relatively shallow, with no overhang as in the conventional piston arrangement. This both increases the ease of manufacture of the pistons 10’, 10” by casting or machining processes, and results in pistons 10’, 10” of reduced mass compared to the conventional piston 10. Machining process may involve any one or more of a number of machining processes, such as additive manufacturing methods or forging.

It will be appreciated that various changes and modifications can be made to the present invention without departing from the scope of the present application.

For example, while the invention has been described in respect of a cylinder 18’ having a diameter of 83 mm, it will be appreciated that the cylinder could take any suitable from and could have any diameter from a wide range of suitable diameters. Further, while the invention has been described in respect of a vehicle 24’ powered by way of an internal combustion engine 22’ only, it is envisaged that the piston 10’, 10” and fuel injector 70’ arrangements and the injection strategy B described herein may equally be applicable to a hybrid vehicle. In this case, the driveline of the vehicle is connected to two prime mover devices in the form of a cranks haft- in teg rated motor generator (CIMG) and an internal combustion engine, with connection and disconnection between the engine and the CIMG being facilitated by way of a disconnect clutch. The CIMG is further coupled to the input of a transmission of the hybrid vehicle, the output of the output of the transmission being coupled to the driveline. In the hybrid vehicle, the driveline is configured to transmit power from one or both of the internal combustion engine and the CIMG to front and/or rear axles of the vehicle. As is conventional, the CIMG of the hybrid vehicle is linked to a battery of the vehicle via a high voltage power invertor.

It is envisaged that the pistons 10’, 10”, fuel injectors 70’ and control system 42’ described herein could be applied to the hybrid vehicle in the same way as described above, without modification. The vehicle may alternatively take the form of an electric vehicle, in which energy for propulsion is provided by a battery. In this case it is envisaged that the piston 10’, 10” and fuel injector 70’ arrangements may form part of an auxiliary power unit, or range extender of the vehicle, which is used to drive an electric generator that charges the battery.

In addition, while the internal combustion engine 22’ and piston arrangements 10’, 10” described herein have been described as forming part of a vehicle, it will be appreciated that these may be applied to any one of a number of mechanical systems. For example, the internal combustion engine 22’ may be used to drive an electric generator, for powering one or more electrical grids for example, or to operate a pump or propeller.