GB2110791A - Variable compression ratio pistons - Google Patents

Abstract

A piston for an internal combustion engine, has a crown portion or upper end and parts (1 and 2) capable of limited relative movement defining a first or upper chamber (3) and a second or lower chamber (11). Relative movement of the parts in one direction will increase the volume of the upper chamber and decrease the volume of the lower chamber. This will increase the compression ratio. Relative movement in the other direction will have the opposite effect. A connecting rod (5) has therein an oil supply passage (4) connected to the first and second chambers by non-return valves (10 and 21 respectively) for permitting oil flow to these chambers. There is also discharge valve means for controlling oil flow from the first chamber at a first flow rate when the oil pressure in the first chamber is at a first level and a second higher discharge rate when the oil pressure in the first chamber is at a second higher level. The discharge valve means may comprises two separate discharge valves (14 and 25) or a single discharge valve (Fig. 4). The second chamber has a restricted discharge passage (12) for controlling the flow of oil therefrom. <IMAGE>

Description

SPECIFICATION Internal combustion engines and pistons therefor The invention relates to internal combustion engines and pistons therefor and in particular to means for varying the compression ratio of an internal combustion engine to suit engine loads.

Variable compression improves efficiency by increasing the compression ratio at part loads.

With highly pressure charged engines, having a fixed compression ratio, it is necessary to use a relatively low ratio to avoid too high stresses at full load.

It is known, see for example British Patent No. 762074, to provide a piston in which an inner member is coupled to the gudgeon pin and said inner member is surrounded by an outer member and separated from the outer member by an upper and a lower oil chamber.

The upper oil chamber supports the outer member against cylinder gas pressure, and is controlled by a valve which opens at a predetermined level of oil chamber pressure to control the level of cylinder gas pressure. The upper oil chamber is supplied with oil via a non-return valve which opens when inertia of the outer member of the piston tends to increase the volume of the upper oil chamber.

The lower oil chamber has a restricted outlet oil passage which limits the rate of upward movement of the outer member with respect to the inner member in order that the volume increase of the upper chamber does not exceed the rate of oil supply. The lower oil chamber is supplied via a non-return valve in a manner similar to the upper chamber.

A possibility during operation of the above described mechanism is that sudden application of engine load, when the engine is at high compression ratio, may require a rate of decrease of volume of the upper chamber which is faster than permitted by the oil flow capacity of a discharge valve of the upper chamber. The provision of much greater flow capacity, through the existing discharge valve of the upper chamber, could lead to excessive cyclic movement of the outer member with respect to the inner member during normal engine operation.

The invention provides a variable compression ratio piston for an internal combustion engine and having means which provides an increased rate of reduction of compression ratio with increased rate of application of engine load.

The invention also provides a piston for an internal combustion engine, having a crown portion and parts capable of limited relative movement defining a first chamber and a second chamber, such relative movement in one direction increasing the distance between the crown portion of the piston and the means for converting reciprocating movement of the piston into rotary motion of the engine output shaft and also increasing the internal volume of the first chamber and decreasing the internal volume of the second chamber whilst such relative movement in the other direction decreases the said distance and also decreases the internal volume of the first chamber and increases the internal volume of the second chamber, a supply passage adapted to receive substantially incompressible fluid, i.e. liquid, under pressure from a source external to the piston, a first inlet passage communicating between the supply passage and the first chamber, a second inlet passage communicating between the supply passage and the second chamber, a first non-return valve permitting flow of fluid through the first passage towards the first chamber, a second nonreturn valve permitting flow through the second passage towards the second chamber, discharge valve means for controlling flow of fluid from the first chamber at at least a first discharge rate, when the pressure of fluid in the first chamber is at a first level and at a second higher discharge rate when the pressure of fluid in the first chamber is at a second higher level, and a discharge passage leading from the second chamber and for controlling the flow of fluid from the second chamber.

The invention still further provides an internal combustion engine having a plurality of pistons as set forth in one or other of the two preceding paragraphs.

Preferably, the discharge valve means comprises at least two discharge valves, one of the discharge valves being arranged to open when the pressure in the first chamber is at a higher level than that at which the other discharge valve opens, and, advantageously, the discharge rate of said one discharge valve is greater than the discharge rate of said other discharge valve.

Hence, improved control of pressure in the upper chamber can be achieved by means of two separate rates of discharge of liquid e.g.

oil from the upper chamber, each rate starting and finishing at separate pressure levels in said upper chamber. The first rate only is used for normal control. This discharge rate starts at a preset level of oil pressure in the first or upper chamber and proceeds for a limited time as it terminates at a pressure slightly bellow the starting pressure. This ensures that the cyclic movement, of the outer member relative to the inner member, is limited, typically to 1 /50th of the total possible move ment. The second discharge rate commences at a pressure level slightly higher than that at which the first discharge rate starts, so it does not operate if the first discharge rate is sufficient for adequate control.Furthermore the sec ond discharge rate proceeds at a much greater rate than the first discharge rate and contin ues until a lower pressure level is reached in the upper chamber, hence a greatly increased cyclic movement of the outer member occurs, and the complete range of reduction of com pression ratio can be accommodated in a few engine cycles, say less than five cycles.

The rapid reduction of compression ratio could lead to problems in filling the second or lower oil chamber which undergoes a corresponding increase of volume, and requires a greater rate of oil supply than can be delivered in the normal manner via the supply passage, conveniently provided in the con necting rod. The increased rate of oil supply to the lower chamber during rapid decrease of compression ratio is, preferably, supplied by arranging for the second, high rate discharge from the upper chamber to be delivered to the lower chamber.

In a preferred embodiment the two discharge rates are controlled by two similar discharge valves, the differing opening and closing pressures being set by adjustment of the seat diameter and spring load. Flow rate through the valves can be controlled by adjustment of valve lift. The valves are designed so that once open they always have sufficient seat clearance that they cannot act as particle traps, that is greater than 0.1 mm. Maximum seat clearance will be dependent on piston size, but will be controlled to reduce impact wear at the seat. Because of the reduced number of operating cycles which the second valve will undergo relative to the first valve, the second valve can be given a greater seat clearance to increase its flow capability. If necesssary, for extreme conditions, a third discharge valve may be provided.

A feature of the preferred discharge valves is that by opening towards the upper oil chamber, inertia, of the valve and spring, tends to compensate for the inertia of the outer member and of the oil contained therein so reducing the variation with engine speed, of the controlled level of cylinder gas pressure. The discharge valves may be set at an angle to the vertical either to provide more accurate compensation for the inertia or to facilitate fitting in the inner member.

Another method of adjusting inertia compensation which avoids side thrust of the valve, is to direct oil from the supply passage to the spring chamber of the discharge valve and to control the area of the valve subject to this pressure.

The invention will now be more particularly described, by way of example, with reference to the accompanying drawings, wherein: Figure 1 is a cross section of one embodiment of a piston according to the invention; Figure 2 is a cross section of part of another embodiment of a piston according to the invention; Figure 3 is an enlarged view of the seat arrangement of one of the discharge valves of Figs. 1 and 2; Figure 4 shows an alternative discharge valve with full inertia compensation; Figure 5 is a cross-section of another em bodiment of a piston according to the invention, in a condition which provides a low compression ratio;; Figure 6 is a section along the line A-A of Fig. 5, but with the piston in a condition which provides a high compression ratio, and Figure 7 shows two alternative discharge valves in cross-section, one on the left hand side and the other on the right hand side of a central plane normal to the paper, Figure 8 is a cross-section of yet another discharge valve, Figure 9 is an alternative non-return valve in cross-section, and Figures 10 and ii are cross-sections, taken at right angles to one another, of yet another embodiment of a piston according to the invention.

Referring to Fig. 1, the piston shown therein comprises outer and inner cup shaped piston members 1 and 2 respectively with upper and lower oil chambers 3 and 11 respectively defined therebetween. The upper end or crown of the outer member 1 forms the head of the piston and the outer member 1 is provided with piston rings. The inner member 2 is connected to the small end of a connecting rod 5 in normal manner by a gudgeon pin.

During the late part of the exhaust stroke and the early part of the inlet stroke, inertia of the piston outer member 1 causes it to rise relative to the inner member 2, so increasing the volume of the upper oil chamber 3. At the same time inertia of an oil column 4 in the connecting rod 5 augments the engine oil pressure. Oil then flows through a pick-up cap 6, chamber 7 and passages 8 and 9, to an inlet (non-return) valve 10 to the upper oil chamber 3. To prevent an uncontrolled rise of the outer member this action compresses oil in the lower chamber 11 and the rate of escape of oil from the lower chamber is controlled by a restrictor 1 2.

The lifting of the outer member reduces the clearance volume 1 3 above the piston so increasing the peak cylinder gas pressure until it reaches a level corresponding to the oil pressure at which a first discharge valve 14 opens. The valve 14 opens when the upper chamber pressure, which is transmitted via passage 1 5 to annular volume 16, to act on the annular area between valve seat 1 7 and guide 18, overcomes the force of spring 1 9.

Once the first valve opens the upper chamber pressure acts also on the area 20 inside the seat circle (see Fig. 3) so increasing the lifting force on the valve to ensure substantial opening. The ratio of the annular area, subject to opening pressure, to guide area, determines the closing to opening pressure ratio of the valve. Arranging for the valve to close at a lower than opening pressure has two-fold benefits. In the first case it ensures that the valve is either shut or substantially open and avoids conditions which could cause cavitation and erosion of the seat 1 7 or particle trapping. Secondly it prolongs the opening period so that size reduction of the valve is possible for a given discharge quantity.Oil passed by the first discharge valve goes directly out of the piston to drain i.e. to the crankcase, As the upper chamber decreases in volume due to action of the discharge valve, the lower chamber 11 increases in volume and is maintained full of oil by flow up the connecting rod passage 3, cap 6, chamber 7, passages 8 and 9 and inlet (non-return) valve 21.

For consistent operation of the discharge valve 14 it is essential that chamber 22, housing the spring 19, does not become pressurised by leakage along the valve guide 1 8. Venting of the chamber 22 may be achieved by drillings through the inner member 2 to its lower surface but, in the embodiments shown, venting is achieved by means of axial passages 23, 24 through valve members 14, 25. These passages connect spring chambers 22, 26 to the low pressure ends of the respective valves. When a valve is open the oil flow 27 (see Fig. 3) past the lower end of the vent passage, by means of its direction and kinetic energy, maintains low pressure in the vent passage 23.The vent passsage may be enlarged for part of its length 28, in order to lighten the valve 14, so as to reduce impact effects at the seat 1 7. All features mentioned relative to the first discharge valve 14 are equally applicable to second discharge valve 25 to be referred to later.

Peak gas pressure in the space 13, hence peak oil pressure in the upper oil chamber 3, occurs for a very short period in each cycle so the amount of oil which is discharged in one cycle is limited. This is desirable for normal operation because a large loss of oil from the upper chamber requires a corresponding large make-up of oil to the lower chamber 11 and this is limited by the capacity of the engine oil pump and supply passages terminating in passage 4. Furthermore excessive relative movement of the outer member to the inner member leads to reduced efficiency of the engine and to excessive wear of the piston parts.

For the case of a suddenly applied load an equally sudden transition from high to low compression is required, necesitating an immediate movement of the outer member from its uppermost to its lowermost position. A second discharge valve 25 is provided and adjusted to accommodate this. The two valves 14 and 25 may be substantially similar for manufacturing economy, however the seat 1 7a of the second valve is of larger diameter than the diameter of the seat 1 7 of the first valve. The annular area 1 6a is made smaller than the annular area 1 6 so the second valve 25 requires a higher oil pressure than the first valve 14 to cause it to open, even with spring 29 set to a lower load than the spring 1 9.

Because of the lower spring load the closing pressure of the second valve is lower than the closing pressure of the first valve hence the second valve 25 stays open for a longer period than the first valve 1 4. As the second valve operates only for a limited proportion of the total number of engine cycles less than the first valve-it may be adjusted to have larger valve lift. The second valve thus has the greater lift, the greater opening period and the greater seat diameter so it has a threefold increase in discharge quantity per cycle over the first valve. This enables a load to be applied suddenly to the engine without causing excessive cylinder pressure with consequent possiblility of damage to the engine.

One difficulty which arises with sudden decrease of compression ratio is that the lower chamber 11 increases in volume at a greater rate than can be filled with oil in the normal manner via passage 4, chamber 7 and passages 8 and 9. To accommodate this the discharge passage 30 from the second valve may be replaced by passage 31 so that the lower chamber may be filled by oil dicharged from the upper chamber. The difference in volume decrease of the upper chamber and the volume increase of the lower chamber is accommodated by concurrent flow through the first valve 14 and discharge through the orifice 1 2.

An alternative arrangement is shown in Fig.

2, in which the second discharge valve 25A is located so as to discharge into passage 8 so that oil flowing out of the upper chamber may enter the lower chamber via the inlet valve 21, or it may be fed back into the engine oil system via chamber 7, cap 6 and passage 4.

The discharge valves 1 4 and 25 are shown in Fig. 1 in a vertical position and arranged so that the inertia, of the valve members and their respective springs 1 9 and 29, at the moment of valve opening, tends to reduce the required oil pressure in the upper chamber 3 to cause the valves to open. This compensates for the effect of inertia of the outer member 1, and of the oil in the chamber 3, which tends to lessen the level of oil pressure in the annulii 1 6 and 28, with respect to the prevailing pressure in the combustion chamber 1 3.

This arrangement of the discharge valves reduces the variation with speed, of peak gas pressure in the combustion chamber 1 3.

As shown in Fig. 2 both the first discharge valve 1 4a and the second discharge valve 25a are inclined to the vertical. This lessens the effect of inertia, of the valves and springs, due to piston motion. The degree of inclination may be adjusted for precise compensa tion of the effect of inertia of the outer member 1 and oil in the upper chamber 3.

The discharge valves 14 and 25 may also be varied in position to suit the space available in different applications.

In a high speed engine it may be possible that the mass of one of the discharge valves causes it to over compensate for inertia of the piston components. An alternative design of discharge valve is shown in Fig. 4 in which oil pressure, which is influenced by engine speed, is ducted to the valve spring chamber and the area of valve influenced by this oil pressure is adjustable by design.

In Fig. 4 a central passage 32 of valve member 33 is in direct communication with the upper chamber 3. Spring chamber 34 is in communication via passage 35 with the oil passage 4 in Fig. 1, through the connecting rod, either directly or via the lower chamber 11 in Fig. 1. Annular chamber 36 is in communication via passage 37 with the engine crankcase so annulus 36 is always at a low pressure when valve 33 is seated.

When oil pressure in upper chamber 3 and passage 32 rises this pressure acts upwards on the valve member 33, over the annular area between the inner seat diameter 38 and the upper guide diameter 39. Because the whole piston assembly is decelerating in the upward direction at the moment of valve opening, oil pressure in passage 32, tending to lift the valve 33, is reduced relative to gas pressure acting on the piston. For the same reason inertia of the valve 33 and a portion of the spring 40 tends to reduce the level of oil pressure in passage 32 required to open the valve. Inertia of the valve and spring has the greater effect so there is a tendency for the valve to over-compensate for inertia. The difference is countered by oil pressure in the spring chamber 34 which tends to keep the valve closed. The pressure in chamber 34 is created by inertia of oil in the connecting rod.

All three varying effects have the same dependence on engine speed so once in balance they are in balance at all engine speeds. The effect of oil pressure in chamber 34 can be controlled by adjustment of the diameter 41, which determines the annular area between this and the guide diameter 39, hence the area subjected to the compensating pressure.

Once the valve 33 lifts off its seat 42, oil under pressure flows into the annular space 36 to create additional lifting force. Some oil escapes via passage 37 but as this is of restricted area the pressure effect of this is negligible. After a short valve lift determined by the overlap 43, oil passes to the spring chamber 34 and passage 35 to the lower chamber of the piston, to prevent voids therein when rapid reduction of compression ratio takes place.

Figs. 5 and 6 show an arrangement suitable for a light weight piston and which embodies additional features to facilitate manufacture and assembly of the piston. All valves are housed in capsules which can be assembled, pre-set and tested outside the piston. Furthermore the valves can be removed and replaced without dis-assembly of the piston. This assembly also eliminates the need for an oil pick-up cap 6 (Fig. 1) to be mated to the external surface of the connecting rod 4.

In Fig. 5 the inlet valve 1 0a to the upper chamber, and the inlet valve 21 a to the lower chamber, are both mounted in the gudgeon pin 44. Both valves 1 0a and 21 a receive oil from the connecting rod passage 4 via radial passages 45 and axial passage 46. Valve 10a is connected to the upper chamber 3 via passages 47, 48. Valve 21 a is connected to the lower chamber 11 via passages 49, 50, 51.

In Fig. 6, a first discharge valve 14b receives oil from the upper chamber 3 via passage 52, 53, and discharges through the lower end 54 direct to the engine crankcase.

A second discharge valve 25b receives oil from the upper chamber via passages 55, 56, and discharges through annulus 57, and passage 58 to annulus 59, round the gudgeon pin 44, in the plane of the passages 49, 50, which communicate with the lower chamber 11.

Because of the accessibility of all of the valves the fixing of piston crown 60 to plate 61, and plate 62 to piston body 63, may be permanent, e.g. by some form of welding, so as to avoid the weight and space required for removable fixing and locking means.

A further simplification is possible by combining the first and second discharge means in a single discharge valve of which Fig. 7 shows two possible configurations. This is similar to the high rate second discharge valve of Fig. 4, but it also provides for an initial lower rate of discharge.

In the combined valve of Fig. 7, oil pressure from the upper chamber 3, passes down through the central passage 32 then bears upward on the annular area between the guide diameter 39, and the seat diameter 38, to overcome the force of the spring 40 and lift the valve 64, to allow oil to flow into small annular space 65 from which oil can escape direct to the crankcase via orifice 66. A limited clearance 67 has substantially smaller flow area than the orifice 68, until the valve lift exceeds L1, and this prevents pressurisation of the annulus 69 for small lifting of the valve 64. For a large increase of engine load, pressure is maintained in the space 32 and valve lift exceeds L1 and the flow into annulus 69 is greater than can be accommodated by the orifice 68. The subsequent rise of pressure in annulus 69 causes further valve lift and when this exceeds L2 oil can flow along the multiple grooves 70, 71 to the spring chamber 34 and discharge passage 35 to the lower chamber as previously described with reference to the second discharge valve of Fig. 4.

The left hand side of Fig. 7 shows grooves 70 to be of uniform large cross-sectional area and the valve to have a limited total lift Ls, to give a substantially constant flow area through the grooves 70 during the fast discharge period. The right hand side of Fig. 7 shows the grooves to be tapered in width and/or depth and the total possible valve lift L4 is substantially unlimited. With this right hand arrangement the flow area through the grooves is proportional to lift and to the rate of application of engine load.

An alternative construction of discharge valve is shown in Fig. 8. This is functionally the same as that shown in Fig. 4, but it could also be built in the two phase mode as that in Fig. 7. Features of the valves shown in Fig. 8 are that the use of disc springs 72 instead of a coil spring enables a reduction in length, hence in mass, of the valve; also the parts of the valve are so designed that they can easily be made by simple sintering and require no subsequent machining. Spring adjustment is achieved by variation of the depth of the chamber 73 either by selection of the cap 74, or by shimming at the face 75.

The inlet valves, 10 and 21, Fig. 1 may also be housed in capsules which can be fitted from the bottom of the piston. The form shown in Fig. 9 consists of a moving plate valve 76, contained within upper and lower body parts 77 and 78, the whole assembly being held in the piston inner member 2 by the retaining ring 79 which may be wedged shaped in section in order to accommodate manufacturing tolerances. Because of the limited lift of the valve 76, and the relatively large diameter of the seat 81, the valve moves rapidly in response to pressure difference and no spring is required. Passage 82 of Fig. 9 corresponds to passage 9 of Fig. 1 and is connected to the engine oil supply via the connecting rod. Passage 83 leads to either the upper or the lower chambers of the piston.When the pressure in either chamber is lower than pressure in the connecting rod, the pressure difference causes the valve 76 to move downward to allow oil to flow from passage 82 via space 84 and seat 81, passages 85 and annulus 86 to passage 83 then to the upper or lower chamber dependent which of these valves is in operation.

Figs. 10 and 11 show an example of a piston with all valves encapsulated as in Figs.

7 and 8, and accessible from the bottom of the piston. To facilitate accommodation of the four valves in the piston inner member, a spherical small end 87 is used on the connecting rod.

In Fig. 10, engine oil passes up to the passage 4 in the connecting rod to the space 7 it then passes through horizontal slots 88, 89, then vertical passages 90, 91, to the two inlet valve assemblies 92, 93. From valve assembly 92 oil passes out of the apertures 94 to annulus 95 and vertical passage 96, shown by broken lines, to the upper chamber.

Similarly oil from valve assembly 93 exits from apertures 97 to annulus 98 then passes through vertical passage 99 and horizontal passage 100, both shown by broken lines, to the lower chamber 101.

Fig. 11, which is a section transverse to Fig. 10, shows the location of the first and second discharge valves 102, 103. When pressure in the upper chamber 3 reaches the first predetermined level the valve 102 opens to allow oil to pass down passage 104, thriugh the valve, and exit through the bottom of the valve 185, direct to the crankcase below the piston. The annulus 106 is connected to the lower chamber 101 via passages 107, 108 shown by broken lines, solely for the purpose of compensation for inertia.

When pressure in the upper chamber reaches the second, higher preset level, the valve 103 opens and allows oil to pass down passage 109. In the first stage of operation of valve 103 oil exits from the lower face 110, and in the second stage it exits also via the apertures 111 to annulus 112, then via passages 113, 114, shown by broken lines, to the lower chamber 101.

Claims (21)

1. A variable compression ratio piston for an internal combustion engine and having means which provides an increased rate of reduction of compression ratio with increased rate of application of engine load.

2. A piston for an internal combustion engine having a crown portion and parts capable of limited relative movement defining a first chamber and a second, chamber such relative movement in one direction increasing the distance between the crown portion of the piston and the means for converting reciprocating movement of the piston into rotary motion of the engine output shaft and also increasing the internal volume of the first chamber and decreasing the internal volume of the second chamber whilst such relative movement in the other direction decreases the said distance and also decreases the internal volume of the first chamber and increases the internal volume of the second chamber, a supply passage adapted to receive substantially incompressible fluid, i.e. liquid, under pressure from a source external to the piston, a first inlet passage communicating between the supply passage and the first chamber, a second inlet passage communicating between the supply passage and the second chamber, a first non-return valve permitting flow of fluid through the first passage towards the first chamber, a second non-return valve permitting flow through the second passage towards the second chamber, discharge valve means for controlling flow of fluid from the first chamber at least a first discharge rate, when the pressure of fluid in the first chamber is at a first level and at a second higher discharge rate when the pressure of fluid in the first chamber is at a second higher level, and a discharge passage leading from the second chamber and for controlling the flow of fluid from the second chamber.

3. The piston of Claim 2, wherein the discharge valve means comprises at least two discharge valves, one of the discharge valves being arranged to open when the pressure in the first chamber is at a higher level than that at which the other discharge valve opens.

4. The piston of Claim 3, wherein the discharge rate of said one discharge valve is greater than the discharge rate of said other discharge valve.

5. The piston of Claim 3 or Claim 4, wherein the said one discharge valve, when open, communicates said first chamber with said second chamber, and said other discharge valve when open, communicates said first chamber with drain.

6. The piston of anyone of Claims 3 to 5, wherein each discharge valve is arranged so as to close at a pressure lower than that at which it opens.

7. The piston of Claim 2, wherein the discharge valve means comprises a single discharge valve movable from a closed to a first open position at a first pressure to discharge fluid therethrough at a first rate and to a second open position at a second pressure higher than said first pressure to discharge fluid therethrough at a second rate higher than the first rate.

8. The piston of Claim 7, wherein said single discharge valve, when in its first open position, communicates said first chamber with drain and when in its second open position, communicates said first chamber at least with said second chamber.

9. The piston of Claim 7 or 8, wherein said single discharge valve is arranged so as to close at a pressure lower than that at which it opens.

1 0. The piston of anyone of Claims 3 to 9, wherein the discharge valve or valves open(s) towards the first chamber.

11. The piston of Claim 10, wherein the discharge valve or valves is/are set at an angle to the vertical.

1 2. The piston of anyone of Claims 2 to 11, wherein one or more of the valves is/are encapsulated.

1 3. The piston of anyone of Claims 2 to 12, wherein one or more of the valves is/are mounted so as to be removable from the end of the piston remote from the crown portion without disturbing the remainder of the piston.

14. The piston of anyone of Claims 2 to 13, wherein each non-return valve includes one or more springs for urging a valve member against a valve seat.

1 5. The piston of anyone of Claims 2 to 1 3 wherein the valve member of each nonreturn valve is freely movable (excepting fluid pressure applied thereto)

1 6. The piston of anyone of Claims 2 to 15, wherein the or each discharge valve has one or more springs for urging a valve member against a valve seat.

1 7. The piston of Claim 16, wherein the discharge valve spring(s) is/are located in a spring chamber which communicates which said supply passage.

1 8. The piston of Claim 1 6 or Claim 1 7 wherein the or each discharge valve spring is a disc spring.

1 9. The piston of anyone of Claims 2 to 18, wherein the piston is coupled to a connecting rod by a part spherical joint.

20. A variable compression ratio piston for an internal combustion engine, substantially as hereinbefore described with reference to anyone of the embodiments shown in the accompanying drawings.

21. An internal combustion engine having a plurality of pistons as set forth in anyone of the preceding claims.