The efficiency of such compressors is largely influenced by the clearance volume (also known as squish volume) when the piston is at the inner or upper most point in its reciprocating movement and it is typically desired to keep this volume to a minimum. One factor which is particularly relevant to the size and nature of the clearance volume are manufacturing tolerances. Typically a number of components are assembled together to form a compressor and the tolerance on each component can affect the clearance volume of the compressor and the head of the pumping chamber in which the piston reciprocates. Providing close tolerances so that the clearance volume is small can cause a compressor to be relatively expensive to manufacture.

It is also desirable, from a manufacturing cost point of view, to keep the construction of the compressor simple and to involve only a minimum number of components and requirements for machining thereof. Ideally, it is typically sought to keep the overall size and weight of the fully assembled compressor to a minimum. Such compressors should also provide effectiveness and reliability in service.

Moreover, where such compressors are used in an air assisted direct injection fuel systems, packaging constraints can make integration of the compressor onto the engine difficult. In low capacity applications, such as single cylinder scooters, the parasitic effect of the compressor can have greater relative significance than for higher capacity engines.

Accordingly, one aspect of the present invention provides a compressor comprising a working chamber having one end open, a piston located in said chamber and arranged to reciprocate therein in sealed rotation to the chamber internal wall, a plate valve member adapted and located to span said open end of

the chamber, said plate valve member being constructed and arranged so that at least the perimetal portion thereof will oscillate in the axial direction of the chamber between a closed position sealing said open end of the chamber, and an open position wherein at least said perimetal portion is axially spaced from said open end of the chamber.

In one form of the compressor, a central portion of the plate valve is restrained against movement relative to the cylinder, with an annular perimetal portion extending about the central portion being deflectable to oscillate in the axial direction. The degree of movement of the perimetal portion progressively increases across the width of the perimetal portion from the central portion to the outer perimeter of the plate valve.

In another form, the plate valve oscillates in the axial direction as a unit, so the movement in the axial direction is substantially uniform across the whole of the area thereof.

Each of these constructions will provide substantially the same cross sectional area of the flow path between the chamber and plate valve when in the open position.

Conveniently, the working chamber is a cylinder and the compressor is configured as a gas or air compressor.

Conveniently, the cylinder is of uniform diameter over the full length of the stroke of the piston and up to the level of the plate valve member when the plate valve member is in the closed position. It is also possible for the piston to move in the cylinder up to a point substantially flush with the upper extremity of the cylinder internal wall. The plate valve member may simply be arranged to seat on an extension of the cylinder internal wall.

Conveniently, around the upper perimeter of the cylinder, there is provided a narrow annular seat surface upon which a corresponding portion of the plate valve member seats to isolate the interior of the cylinder from a compressed gas outlet.

The plate valve member or the perimetal portion thereof is lifted in response to the rise of pressure in the cylinder to provide an annular passage between the annular seat surface formed on the upper perimeter of the cylinder, and the corresponding

peripheral edge portion of the plate valve member, to provide communication between the interior of the cylinder and the gas outlet, as the piston moves upwardly in the cylinder.

The compressor conveniently is provided with a closure portion, such as a cylinder head or other equivalent means above the plate valve member to close off the compressor.

This cylinder head can be configured to determine the maximum stroke of the plate valve member or the perimetal portion thereof in that when displaced off its annular seat surface, the upward movement of the plate valve member is restricted by contact thereof with the cylinder head or a part thereof. Accordingly, by adjusting the location of the cylinder head, the stroke of the plate valve member may be varied.

Further, the internal surface of the cylinder head may be provided with small dimples to prevent the plate valve member from sticking to a flat surface thereof.

The presence of oil on the cylinder head or valve member can result in sticking of the valve to the cylinder head. However, when the compressor is constructed so that only the perimetal portion thereof moves relative to the annular seat surface provided by the cylinder, then the cylinder head has a generally centrally positioned projection that the plate valve contacts to prevent axial movement of the central portion thereof. Thus the perimetal portion of the plate valve is resiliently deflected when open and will thus be urged to return to the closed position. The plate valve is not required to be fixed or attached in any way to the cylinder head, and can be so arranged to provide a pre-load on the plate valve in the direction to close the valve.

It is possible, with the construction of the compressor where the whole of the plate valve moves, to achieve a negative squish. This is primarily possible due to the"free"unrestrained nature of the plate valve member.

Negative squish is where the piston may extend marginally beyond the upper extremities of the cylinder bore. The piston would not hit the valve member as it would have been displaced off its annular seat prior to the piston extending

beyond the cylinder bore upper extremity, by an extent that is of the order of millimetres and fractions thereof.

This aspect is useful as it means that tolerances, such as manufacturing tolerances, do not have to be exact and that the compressor will still successfully operate with the piston extending past the cylinder bore extremities as stated. This is not possible in a traditional reed valve type compressor.

Accordingly, a further aspect of the present invention provides a compressor comprising a working chamber having one end open, a piston located in said chamber and arranged to reciprocate therein in sealed relation to the chamber internal wall, a plate valve member adapted and located to span said open end of the chamber, wherein said plate valve member in use, in response to reciprocation of said piston in said chamber, oscillates in the axial direction of the chamber under pressure of fluid intermediate said plate valve member and an upper portion of said piston such that said fluid pressure permits said upper portion of said piston to reciprocate into a position at least co-extensive with the open end of said chamber and to simultaneously locate said plate valve member in spaced apart relation to said upper end of said piston.

According to another aspect of the present invention there is further provided an internal combustion engine comprising a compressor with a reciprocating piston co-operatively associated with a rotating member of said engine ; said co-operative association adapted to impart operational motion to said compressor and wherein said compressor has a reciprocating piston for compressing gas and a plate valve member for exhausting compressed gas.

Preferably said plate valve member adapts said engine to operate with manufacturing tolerances sufficient to, in use, permit negative squish of said piston of said compressor.

Preferably said rotating member is a crankshaft of said engine though a cam shaft for operation of valves may also be used.

Preferably said crankshaft has a cam located thereon for imparting reciprocating motion to said piston in a direction radial to the axis of said crankshaft.

Preferably said cam is eccentrically located relative to said axis of crankshaft and preferably said cam is circular so as to impart sinusoidal reciprocating motion to said piston over a single revolution of said crankshaft.

One or a number of ports may be provided in the wall of the cylinder of the compressor and are arranged to be open as the piston of the compressor approaches its lower most point in the cylinder, so that air or another gas can pass into the cylinder at a level above the piston. Other means of providing for the entry of gas to the cylinder may also be used. As the piston moves upwardly in the cylinder, the port or ports are closed by the piston in the known manner. Thereafter the air or gas trapped in the cylinder above the piston is compressed to a level to lift the plate valve member off the annular seat to provide an annular passage therebetween and thereby allow discharge of the air or gas from the cylinder through an outlet. As the plate valve member is displaced from the annular seat under pressure of compressed air, the piston can reciprocate such that it has zero or negative squish without fouling or contacting the plate valve member. In certain arrangements negative squish of at least 1. 7% of stroke length has been achieved whilst maintaining a satisfactory compromise between cost and operation of the compressor, however testing indicates that negative squish of up to 2. 5% of stroke length could be utilized whilst maintaining satisfactory operation of the compressor.

In particular manufacturing tolerances sufficient for squish to vary between positive squish of 3. 0% of stroke length and negative squish of 1. 7% of stroke length have been found to provide a satisfactory compromise between cost and performance of the compressor. Such cost and performance being a crucial tradeoff in the design and development of low cost applications, such as scooters and particularly single cylinder scooter applications.

The compressor is configured such that the valve member seats on the annular seat surface when the piston is on a downward stroke due to the delivered gas pressure creating a sufficient differential to the pressure in the cylinder.

Also it is possible, but not essential for a spring to be provided on either side of the valve member (depending upon whether it is a tension or compression spring) to aid in the return of the valve member onto the annular seat surface and

in the sealing thereof thereagainst. When the valve member is supported so only the perimetal portion thereof moves to open the valve, no spring is required as the resilience of the valve member will effect return thereof.

In an alternative form, the upper end of the cylinder is closed by a plate like member with a plurality of apertures therethrough and the plate valve is seated thereon. The plate valve also has a plurality of apertures therein, arranged to not align with or partly overlap with the apertures in the plate closing the upper end of the cylinder. As the pressure rises in the cylinder, the gas pressure is applied to the plate valve through the holes in the plate to thereby raise or deflect the plate valve whereby the gas can pass through the apertures in the plate valve and/or about the periphery of the plate valve. This construction reduces the mass of the plate valve.

The compressor as described above has the advantage that the valve controlling the delivery of the fluid at the required pressure from the cylinder is of a particularly simple construction with a minimum of components and is arranged to provide a comparatively large exhaust or outlet area for the compressed fluid. The use of a simple plate member seated on a continuous annulus seat, as the delivery valve, is a particularly economical construction compared with more conventional reed valves that require an appropriate housing to support the valve arrangement and a form of anchorage for at least one edge of the actual valve member to the housing. Also, by suitable selection of the thickness of the plate or varying the stroke of the plate, the delivery rate of the compressor can be changed.

In comparison, a compressor having a reed valve type exhaust or outlet valve would require as"extra"components a stop for reed valve, a reed plate upon which the reed valve would act, a seal between the reed plate and the upper end of the compressor cylinder bore and some means to fix or attach the reed valve to the reed plate. All of this is replaced by a simple plate valve which preferably is a disk and which preferably acts on the upper end of the compressor cylinder bore in the present system. Hence, only one part instead of five parts are used. Hence, the compressor is cheaper to make and easier to assemble. Importantly, the high pressure joint, at which point the seal is located in a traditional reed valve system,

is eliminated. Further, unlike the traditional reed valve arrangement, there is no clearance volume, such as the exhaust plate port squish volume in a traditional reed valve arrangement.

In the currently proposed construction, the valve member is not deflected or distorted cyclically to open and close the valve as is normally the case with a reed valve. Rather, the plate valve member is merely physically displaced by the pressure difference existing thereacross. Also, the use of the plate valve member enables a relatively large flow area to be achieved by a comparatively small movement of the valve member, particularly as compared with conventional reed valve systems. Further, the plate valve has freedom to flex along its whole perimeter when seated which reduces the risk of breakage and defective seating.

The nature and configuration of the delivery valve results in there being an even loading on the delivery valve which is desirable from durability and long-life point of view. This can be compared to a traditional reed valve arrangement where there is clearly uneven loading on the reed itself which is caused to flex repeatably about one part thereof.

The plate valve member is however able to flex due to pressure thereon whilst it is seated, this ensures it does not break or create problems at the seat.

The construction, configuration and the flex would be even across the valve, operation of the compressor herein disclosed renders the compressor particularly suitable where a compressor of a capacity of the order of 25 to 100 cc is required though smaller capacities such as 3cc can also be achieved.

The piston may be driven by any suitable means and in one preferred embodiment, is driven by a cam mounted on a rotating shaft. This shaft may be incorporated in an internal combustion engine, and take form of a crankshaft or camshaft. Preferably, the connection between the piston and the cam on the rotating shaft is not via a conventional piston rod, but is by a follower and preferably a roller follower that is rotatably mounted in association with the compressor piston and that is co-operatively associated with a cam surface formed with or attached to the cam on the rotating shaft. A spring may be provided to promote the return movement of piston and roller and to maintain operational contact between the cam

and rotating shaft so as to maintain the compressor in synchronization with the cam.

According to a further aspect of the present invention there is provided a two stroke internal combustion engine being crankcase scavenged and comprising a reciprocating piston compressor co-operatively associated with a rotating member of said engine located in said crankcase ; said co-operative association adapted to impart operational motion to said compressor in timed relation with said engine and wherein said compressor is in fluid communication with an internal cavity of said crankcase to thereby increase scavenging volume of said crankcase.

In one preferred form, the increase in scavenging volume is provided by an internal wall of a cylinder that houses the reciprocating piston being in fluid communication with the crankcase. This arrangement has the additional advantage that oil mist in the crankcase will lubricate at least some of the moving parts of the compressor.

It is also preferable that the inlet valve of the compressor is in fluid communication with the crankcase and that it is arranged so as in use to inlet air from the crankcase when said air from said crankcase is at maximum pressure.

This may be achieved by arranging the cycle of the piston of the compressor to be in phase, or nearly in phase with the piston of the engine so that both pistons are at top dead centre at nearly the same point in time and so that both of the pistons are at bottom dead centre at nearly the same points in time. This helps to maximise scavenging volume and inlet pressure. Preferably the cycle of the piston of the compressor operates in timed relation to the piston of the engine so as to reach top dead centre in a range of between 30 degrees before the piston of the engine reaches top dead centre to 60 degrees after the piston of the engine reaches top dead centre.

One application of the compressor is in dual fluid fuel injection systems that entrain fuel in air when injecting the fuel into a combustion chamber. Typically, a metered quantity of fuel is delivered into a holding chamber of delivery injector.

Compressed gas, such as air, is in fluid communication with the holding chamber of the delivery injector. Upon opening of the delivery injector the compressed gas

operates as a propellant that delivers the fuel into the combustion chamber.

Typically such a system is used in a direct injection gasoline application for both two stroke and four stroke application.

To operate such a system a supply of compressed air is required. It is convenient to use the internal combustion engine as a power source for operating the compressor that supplies the compressed gas to the fuel injection system.

Examples of dual fluid injection systems are detailed in the applicants Untied States patents No. 4, 934, 329 and 5, 622, 155 which are incorporated herein by reference. US 5, 622, 155 specifically shows application of dual fluid systems to small engine applications.

The invention will be more readily understood from the following description of preferred embodiments of the present invention as applied to an internal combustion engine and as illustrated in the accompanying drawings.

In the drawings, Figure 1 is a longitudinal sectional view of a two stroke cycle engine incorporating a compressor of the construction discussed above ; Figure 2 is a longitudinal sectional view of the compressor as shown in Figure 1 ; Figure 3 is a detained longitudinal sectional view of the cylinder head portion of the compressor shown in Figure 2 with the construction of the valve modified ; Figure 4 is a partial sectional view showing the compressor in spatial relationship to a crank shaft and associated web on the crankshaft of the engine of Figure 1 ; Figure 5 is a partial sectional view through the web of crankshaft in Figure 4 showing a roller follower of compressor in spatial relationship to an eccentric cam on the web of the crankshaft ; Figure 6 is a longitudinal sectional view of an alternate compressor construction ; and Figure 7 is a plan view of the compressor in Figure 6.

Referring now to the drawings, Figure 1 shows a single cylinder two stroke cycle internal combustion engine 10 of conventional construction having an engine

piston 12 connected by a connecting rod 13 to a crankshaft 14 housed in a crankcase 16. The two stroke cycle engine 10 is of the common crankcase scavenged construction and will only be further described herein where necessary in relation to describing a compressor unit 15 incorporated therein.

The compressor unit 15 is mounted on an external wall 11 of the crankcase 16 of the engine 10 in a position and manner so that a lower portion 17 of the compressor unit 15 projects into a cavity 18 in the crankcase wall 11. The compressor 15 is attached to the external wall 11 so as to provide an air tight seal.

The lower portion 17 of the compressor unit 15 is located adjacent an aperture that extends through the full thickness of the crankcase wall 11 so that a corresponding aperture in the lower portion 17 of the compressor 15 permits a cam follower 20 of the compressor unit 15 to engage a cam surface 21 formed integral with the crankshaft 14. The cam 21 may be circular in shape and located eccentric to the axis of the crankshaft. In addition to allowing contact between the a cam follower 20 and the cam surface 21, alignment of corresponding apertures in the crankcase 16 and the lower portion 17 of the compressor 15 permits fluid communication between a cavity defined by the crankcase 16 and a cavity defined by the compressor 15.

As shown in Figure 2, the compressor unit 15 has a cylindrical compressor bore 24 in which a compressor piston 25 can axially reciprocate. The compressor piston 25 having mounted in a lower portion thereof the cam follower 20 which engages the cam surface 21. The cam follower 20 is preferably cylindrical and rotatably mounted on a follower bearing 22 located on a follower mounting shaft 23 in the lower portion of the compressor piston 25. It will thus be appreciated that as the crankshaft 14 rotates, the cam surface 21 will cause the cam follower 20 to reciprocate relative to the housing of the compressor and in turn cause the compressor piston 25 to reciprocate within the compressor bore 24.

It will also be appreciated that to achieve contact between the cam 21 and the follower 20 the may need to be of a relatively larger diameter than the follower so the distance between the crankshaft and the aperture in the crankcase adjacent the compressor can be bridged. Accordingly the roller follower may tend to rotate

at a higher speed than the crankshaft as the follower under such a construction is likely to have a smaller diameter than the cam surface on the crankshaft.

Accordingly it has been found that a roller follower will preferably have a diameter that is at least one quarter the diameter of the cam. Best results have been obtained with the roller follower having a diameter one third the diameter of the cam. Preferably needle bearings are utilized for the roller follower.

The cam 21 is preferably circular and is located so that its center is laterally displaced from the axis of the crankshaft, though the axis of the cam 21 and the axis of the crankshaft are preferably parallel. Location of a circular cam in this manner allows sinusoidal reciprocation to be imparted to the compressor piston during a single rotation of the crankshaft. This is believed to promote even wear on the cam surface and the follower bearings.

The cam 21 is preferably arranged on the crankshaft 14 so that compressor piston 25 reaches top dead centre in the compressor bore 24 in timed relation to the engine piston 12 over a range of 30 degrees before the engine piston 12 reaches top dead centre to 60 degrees after the engine piston has reached top dead centre.

Further details of the construction and operation of the compressor unit 15 will now be described.

Lubrication of the moving components of the compressor unit 15 is conveniently provided by oil mist in the engine crankcase 16. This oil serves to lubricate the bearing 22 of the follower mounting shaft 23 and is communicated to the compressor via corresponding apertures in the base of the compressor 15 and the wall of the crankcase 16.

A compression spring 26 located in an annular cavity 27, extending about the compressor bore 24, engages an annular carrier plate 28 suspended from the respective ends of the follower mounting shaft 22, which support the cam follower 20, via a bearing 22, to an upper internal face of the cavity 27, and hence effectively returns the compressor piston 25 during the downward stroke of the piston. Equally, the plate 28 and the spring 26 prevent undesirable lateral movement of the shaft 23. The spring 26 may alternatively be in the form of a gas

spring. It is also preferable that the spring have a pre-load force when the piston is at bottom dead center as this assists the compressor to operate at frequencies exceeding the natural frequency of the spring.

The cylindrical compressor bore 24 of the compressor unit 15 extends through the full axial length of the portion thereof in which the compressor piston 25 reciprocates. At the upper extremity of the compressor bore 24, is a peripheral shoulder 55, the internal face of which is coextensive with the internal face of the compressor bore 24. An annular groove 29 forms the outer wall of the shoulder 55 so that the shoulder 55 defines a continuous annular upper face 30 which is coaxial with the axes of the compressor bore 24. A closure portion 32 is located at an axially spaced location above the peripheral shoulder 55 to provide a control face 33 which defines a nominated axial distance between the shoulder 55 and the closure portion 32, and also defines the maximum stroke of a delivery valve member 35. A delivery port 31 is in fluid communication with the annular groove 29. The closure portion 32 may typically be a cylinder head.

The delivery valve member 35 is preferably in the form of a flat planar member often referred to as a plate and may preferably be in the form of a disc having a diameter at least equal to or greater than the outer diameter of the peripheral shoulder 55 and which is typically sufficiently rigid to in use remain substantially flat. The spacing in the axial direction between the shoulder 55 and a control face 33 of the closure portion 32 is sufficient to allow the valve member 35 to move therebetween in the axial direction so as to provide fluid communication between a pumping chamber 34 and a compressor delivery port 31. The extent of axial movement of the valve member 35 from the shoulder 30 is not required to be substantial, in view of the circumferential extent of the resultant opening, in order to provide the required flow area to deliver the fluid from the pumping chamber 34 to the delivery port 31.

A compressor inlet port 36 through which the fluid enters the pumping chamber 34 is preferably located in the wall of the compressor bore 24 at a point such that it is exposed when the compressor piston 25 is at or near to the end of the return or intake stroke thereof. The inlet port 36 is open and closed cyclically

by the compressor piston 25 as it reciprocates in the compressor bore 24 as is commonly known. Accordingly such a compressor may be referred to as being of piston ported construction.

The above described compressor construction and operation is specifically unique in relation to the construction and operation of the plate valve or delivery valve 35 and the operation thereof. A delivery valve 35 in the form of a flat disc is very cheap to produce such as by a simple stamping operation.

The piston clearance relative to the flat disc delivery valve 35 can be maintained very small thus improving delivery efficiency.

The compressor bore 24 is of uniform diameter and extends substantially the total length of the compressor unit 15 thus making machining, and honing if desired, of the compressor bore 24 simple and less time consuming. In particular, a through honed bore is possible in the case where, for example, the closure portion 32 is formed as a cylinder head and is subsequently fitted to the end of the compressor bore 24. Although a good surface finish on the piston and bore are preferred, a reasonable surface finish achieved by simple machining techniques can be adequate. In such a case, a simple piston ring 39 could be provided on the piston shirt as shown in Figure 2. Delivery valves of differing thickness can be used to alter the displacement of the valve and hence the amount of gas or air which the compressor can supply.

The flat disc form of the delivery valve 35 and the fully floating mode of operation thereof give a long life quality and the stroke of the valve 35 can be varied readily such as by varying the axial position of the annular shoulder 30 in the compressor body. The simple nature of the parts and the limited number of parts results in a fairly easy and straightforward assembly of the compressor.

Further, because of the total area of the delivery flow path, the risk of blockage in the delivery flow path when the disc valve is in the open position is reduced. This is due to the large flow of air/gas through the delivery flow path which serves to enable discharging gas to blow any residual matter or debris from the area of the valve member and annular seat surface. Also due to the high frequency movement of the valve member onto and off of the seat, this will serve to

physically dislodge any residual matter from the aforementioned area. Of course, alternative means for inletting air to the compressor may be used instead of the ported inlet 36 as described. For instance the inlet port 36 may communicate with a cavity, such as cavity 40, surrounding the compressor bore 24 and housing the spring 26 whereby the compressor draws air from the engine crankcase 14 rather than directly from the surrounding atmosphere.

By appropriate selection of the timing of the movement of the cam 21 with respect to the stroke of the engine piston 12, the engine crankcase 14 can take in more air than it normally would. That is, if air is being taken into the pumping chamber 34 whilst the conventional reed valve of the crankcase 14 is open, more air can be taken into the engine crankcase 14 by the amount of air drawn in by virtue of the movement of the compressor piston 25. Also, depending upon the timing and phasing, it is conceivable to take air from atmosphere into the pumping chamber 34, while the engine piston 12 and or compressor piston 25 are contributing to compression of the air in the crankcase 16. By appropriate timing, such compressions may serve to supercharge the compressor. Alternatively, locating the inlet port 36 of the compressor in fluid communication with the crankcase 16, allows such compressions to effect delivery of air to the delivery port at higher pressures, which represents an increase in compressor efficiency.

Further, it is also possible for the compressor to be arranged to draw fuel with air into the pumping chamber 34 such as from a fuel collection device or from a fuel supply source via a fuel metering device along with drawing air or gas on during the intake stroke of the piston. The compressor unit 15 can then be used to deliver a quantity of fuel and air or gas wherein the fuel is entrained in the air or gas. This may be applicable to low cost simple engine applications.

There is shown in figure 3 a modification to the support of the delivery valve 35 wherein the cylinder head 32 has a central projecting block 32a that engages the delivery valve member 35 centrally of the side thereof which is opposite to that which engages the shoulder 30 adjacent the cylinder bore 24. The axial length of the projecting block 32a is such that when the delivery valve 35 is in the normally closed position and seated on the shoulder 30 the block 32a is also in contact with

the delivery valve on the opposite side thereof to the shoulder 30. The effective axial length of the block 32a may be such as to apply a pressure to the delivery valve 35 which in turn will be in pressure contact with the shoulder 30.

In this arrangement the opening of the delivery valve is effected by the pressure of air or gas in the pumping chamber 34 during the compression stroke of piston causing the peripheral edge of the valve member 35 to deflect upward whilst the central portion remains seated on the block 32a. If desired the axial length of the block may be such as to effect a limited degree of deflection of the valve member 35 into a dished shape.

Referring now to Figure 4 which is a partial sectional view of the engine of Figure 1 showing the spatial relationship of the cam follower 20 to the eccentric cam surface 21 on the crankshaft 14. As the crankshaft 14 rotates the eccentric cam surface rotates in an eccentric manner about the axis of the crankshaft. This causes the cam follower to be laterally displaced by an amount corresponding to the lateral eccentricity of the cam. This lateral displacement of the cam follower 20 corresponds to the stroke length of the compressor piston 25. In an alternate form the cam surface 21 is circular and has its center laterally displaced from the axis of the crankshaft 14. In this way the circular cam surface 21 has eccentric movement relative to the axis of the crankshaft 14. This eccentric movement of a circular surface translates into sinusoidal oscillation of the cam follower 20 over a singly revolution of the crankshaft 14.

Referring now to Figure 5 which is a sectional side view of Figure 4. It shows in further detail the cam follower 20 and its spatial location relative to the cam 21. It also shows the spatial relationship of the cam 21 relative to the crankshaft 14 and specifically a crank web 37.

The centre line of the compressor piston 25 is offset by 130 degrees from the centre line of engine piston 12. This arrangement is not essential and serves to ensure that oil in the crankcase does not drain into the portion of the compressor 15 that is in fluid communication with the cavity internal to the crankcase wall 11.

Referring now to Figures 6 & 7 which detail a preferred form of the compressor for application to single cylinder two stroke crankcase scavenged

internal combustion engine having an air assisted direct injection fuel delivery system. The compressor is a 3cc unit typically used on small single cylinder two stroke engines of capacity such as 50cc.

In particular Figure 6 shows a sectional side view of a compressor having a disk valve member 35 that operates substantially as for the compressor detailed in Figure 2, except for a return spring 61 that operates between the upper surface of the disk valve member 35 and the internal surface of closure member 63. This ensures that disk valve returns to it resting position on annular shoulder 30 after compressed gas, under the influence of the compressor piston 25 has forced the disk valve member 35 free from the annular shoulder 30. The return spring 61 is located relative to the closure member 63 by projection 60. The projection 60 prevents lateral movement of the return spring 61 relative to the surface of the disk valve member 30 and closure member 63. The return spring 61 is particularly advantageous where there is a possibility of the disk valve member 35 being wetted with oil or other liquid and hence where there is a possibility of the disk valve member 35 sticking to the internal surface of the closure member 63.

An O-ring 64 is located between the closure member 63 and the body 66 that defines the compressor bore 24. An inlet port 65 is provided in the body 66 in such a position that the compressor is a piston ported compressor. In particular, as the compressor's preferred form of operation is as a roller follower that is in contact with circular cam eccentrically located on a crankshaft web, the inlet port 65 is in fluid communication with a crankcase that houses the crankshaft containing the eccentric cam.

Figure 7 is a bottom view of the compressor of Figure 6 showing relative diameters of the compressor bore 24 and the roller follower 22.

Specific details of one embodiment of the compressor according to figures 6 && 7 are as follows : Cam Inputs Cam Outputs Eccentricity 3 mm Cam angular 1047. 198 rad/sec velocity Cam Speed 10000 RPM Total 46. 90 gram reciprocating mass Piston mass 20 gram Max peak 527 MPa contact stress Pin mass 3. 6 gram Equivalent 46 mm conrod length Follower/bearing 11. 8 gram Follower 28333 RPM mass average speed Spring seat 5 gram mass Spring mass 19. 5 gram Spring Pre-Load 80 N Spring Rate 16. 6 N/mm Follower Radius 12 mm Follower effect. 5 mm Length Cam Diameter 68 mm Nominal Offset 0 degree s Compressor Compressor Inputs Outputs Swept Volume 3. 04 Cc Piston 506. 7 mm2 Area

Comp. Ratio 20 Clearanc 0. 1600 cc e Volume Inlet Pressure 0 KPa Bore 25. 399 mm gauge Diameter Delivery 550 KPa Pressure gauge Crankcase 20 KPa Pressure gauge The 3mm eccentricity of the cam provides a stroke length of 6mm to the compressor piston. Small stroke length of this nature is preferred for single cylinder engines as an eccentricity of approximately 3mm on the crankshaft web 37 enables the engine to be balanced, or balanced satisfactorily.

The spring pre-load force (i. e. the compression force in the spring when the crankshaft 14 is at rest and the compressor piston is at bottom dead centre) allows the follower to maintain contact with the cam at speeds exceeding the natural frequency of the spring. Where a pre-load force is utlised, the spring will tend to exert a substantially constant force on the eccentric web throughout a single rotation of the crankshaft when the crankshaft is rotating at a speed that corresponds with the natural frequency of the spring. As the speed of the crankshaft increases beyond the natural frequency of the spring, the force exerted on the cam at the top dead center postion will tend to reduce. The maximum speed at which the crankshaft can rotate without the follower losing contact will be the speed near to that at which the spring exerts no force on the cam at some point during a rotation of the crankshaft and which will typically be near top dead center.

In this way it can be seen that the pre-load increases the speed of rotation of the crankshaft at which the follower will lose contact with the cam surface compared with a system that does not utilise a pre-load force on the spring. Utilisation of a pre-load therefore provides greater flexibility in spring selection as the natural frequency of the spring has reduced influence on the maximum speed of rotation of the crankshaft.

As detailed above, the compressor preferably utilises a disk valve plate that floats within a cavity volume between a closed position that seals the compressor outlet so that the intake is operational and an open position so that the outlet is operational. The disk valve floats on the compressed gas that is intermediate its underside and the top surface of the compressor piston. As the disk floats the piston can achieve negative squish without fouling with, or contacting, the disk valve plate.

The ability to operate with negative squish allows the manufacturing tolerances of the engine, such as the tolerances of the crankshaft, crankshaft web, eccentric cam, crankcase housing and the compressor housing and possibly other aspects of the compressor to be less stringent than otherwise might be the case were negative squish not possible. It has been found that allowing for manufacturing tolerances so that compressor squish may vary in the range between a positive squish of 5% of stroke length and a negative squish of 1. 7% of stoke length provides an adequate compromise between performance and cost. It is believed that negative squish of up to 2. 5% of stroke length could be supported through use of a disk valve member in a compressor.

To promote an effective seal between the plate valve 35 and shoulder 30, it is preferable that the plate valve be flat to within 50 microns and that the shoulder have a surface finish rating 0. 8.

The description of the invention made above is not intended to be limiting to the invention and other variations may be made by those skilled in the art without departing from the scope of the invention as defined in the appended claims.