It is well known to alter the inlet valve timing and, in particular, the inlet valve closing position of an engine in response to changes in speed and load to improve the torque and power output, improve fuel consumption, and improve exhaust emissions.

Engine tests have also shown that substantial improvements can be made to idle fuel consumption and exhaust emissions by very early closing of the inlet valve especially if the engine idle speed is lowered simultaneously. The inlet valve should close later at high engine speed to give increased power and this later closing can be accompanied by an advancement of the inlet valve opening. A variable event camshaft can also be fitted to the exhaust valves to extend the expansion stroke at low engine speeds to further improve fuel consumption and further improve low engine speed vehicle drivability. A variable event exhaust camshaft can also advantageously alter the position of exhaust valve closing.

Many camshaft drive mechanisms that provide variable valve timing are known as phase changers. This type of mechanism does not change the duration of valve timing which is known as the event, but simply advances or retards the camshaft relative to the crankshaft to provide overlap control, that is, control of the period during which the inlet and exhaust valves are open together at the commencement of the inlet cycle.

Event changers are camshaft drive mechanisms which do change the duration of valve timing, by altering one or both of the valve opening and closing times, so that the total number of degrees of drive shaft rotation during which the valve is open is altered from for example high engine speed to low engine speed. There are differences in the requirements of valve timing between gasoline engines and diesel engines. In a diesel engine, a substantial change in valve position may only be made to the inlet closing position, as it is necessary to avoid advancement of the inlet opening due to the close valve-to-piston clearances. Even though the change is made to the inlet closing position only, the duration of valve opening changes substantially from low engine speed to high engine speed. In the case of the exhaust valve in a diesel engine, a change to the opening may be made provided that the position of valve closing is not retarded.

Several different types of event changer are available. One system of this type is disclosed in US Patent No. 4505235 and consists of a drive shaft running inside a hollow camshaft. An eccentric linkage connects the drive shaft to the camshaft.

When the axis of the drive shaft is moved relative to the axis of the camshaft, the camshaft rotates at variable angular velocity and changes the duration of valve timing.

There are a number of problems with the system shown in US 4505235. One problem is a potential problem associated with shaft alignment through a number of movable shaft supports. This arrangement also means that a special chain or belt tensioner is required, or a special seal in the case of a belt driven apparatus. There are also difficulties with engines that use twin overhead camshafts. Some engines with twin overhead camshafts would require a complete redesign of the camshaft drive system to incorporate a variable valve timer as shown in US 4505235.

Some types of variable valve timing mechanisms use an intermediate member between the camshaft and drive shaft. This type of system requires two eccentric linkages, one from the drive shaft to the intermediate member and the other from the intermediate member to the camshaft. An example of this type of variable valve timer is shown in US Patent No. 3633555. Increased frictional losses occur with the double eccentric linkage system and also with the bearing supporting the intermediate member. There are also significant cost penalties with such a system.

US Patent No. 5787849 describes a variable valve timing phase changer that is suitable for use with an event changer. As shown in Fig. 6 of US 5787849, a hollow camshaft is connected to a driving shaft running through the centre of the camshaft by an eccentric linkage, the eccentricity of the linkage being varied in response to engine speed. At high engine speed, the camshaft and driving shaft are concentric so that the camshaft is driven at the same constant angular velocity as the driving shaft. At low engine speed, the camshaft and driving shaft are eccentric so that the camshaft is driven at variable angular velocity to change the duration of valve timing.

The eccentric linkage described in US Patent No. 5787849 consists of an arm that is integral to the driving shaft and which supports a sliding block that runs in a slot provided m a central part of the camshaft. The machining of a slot is a non preferred machining operation in the automotive industry.

"Block and slot"connections are used in many variable velocity camshaft mechanisms, not just those having hollow camshafts. US Patent No. 5333579 discloses an arrangement where the driving shaft is connected to an intermediate member by a block and slot eccentric linkage. The intermediate member drives the camshaft via another block and slot linkage. US Patent No. 5152262 also uses a double block and slot system, with a block and slot linkage positioned at the front and rear of the engine to drive camshafts at variable angular velocity. However, double block and slot systems produce significant frictional losses.

A further problem with hollow camshafts such as that disclosed in US Patent No. 5787849 is that the two halves of the camshaft cannot be made by a sintering process because there are undercuts between the cam and a large diameter section of the camshaft that is necessary to contain the block and slot linkage. Such undercuts cannot be incorporated in the die because of component removal requirements, and have to be machined separately. Any machining reduces the advantages of using a sintering process, and machining an undercut around a cam profile is very complicated and extremely undesirable.

It is also well known to provide variable valve lift systems in which the magnitude of valve displacement from a closed to a fully open position can be controlled. Part load power requirements of a gasoline engine are normally met by throttling the air supply into the engine, incurring fuel consumption penalties due to engine pumping losses. Known variable valve lift systems reduce such losses by eliminating throttling and relying instead on variable valve lift to provide maximum lift at maximum load and reduced lift at lower engine loads. It is also known to combine variable valve lift with other features, for example delayed inlet valve opening combined with reduced valve lift and reduced valve timing duration.

One known variable valve lift system is described in US patent No. 5373818.

The system illustrated in figure 5 of that patent includes a first valve actuation lever supporting a roller which bears against the cam and a second valve actuation lever supporting a roller which bears against the first actuation lever. The first valve actuation lever is slidably mounted on a pin which extends through an elongate slot in the lever.'1 lie second valve actuation lever is pivotally mounted, an end of that lever remote from its pivotal support bearing against the valve. The first valve actuating lever is displaceable by an eccentric mechanism so that the displacement of the valve is a function of the displacement of the first valve actuation lever by the eccentric mechanism. The roller supported by the first valve actuation lever is maintained in contact by the cam by a spring mechanism. That spring mechanism is relatively large, occupying valuable space and absorbing energy as the spring mechanism is forced to contract and then expand each time the cam rotates.

It can also be seen from Figure 4 of US 5,373,818 that the valve timing duration shown for full (maximum) lift is 210 degrees of crank which is a shorter duration than would normally be the case and this limits the full load performance of the engine. Furthermore, if the valve timing duration is reduced to 180 degrees for maximising volumetric efficiency at low engine speeds for increased torque, which would be desirable especially in small engines, the valve lift is reduced so the system has reduced full load engine performance and reduced low engine speed torque.

Performance can be expected to be reduced because of the substantial mass of the first valve actuation lever together with its extra spring loading. This extra mass and spring loading results in greater frictional losses which will reduce engine performance. In addition, the first valve actuation lever the profiled end of which moves in a direction substantially at right angles to the displacement direction of the valve is bulky and expensive to manufacture.

It is an object of the present invention to obviate some or all of the problems outlined above.

According to a first aspect of the present invention there is provided a variable valve timing system for an internal combustion engine, comprising a drive shaft driven by the engine, a camshaft driven by the drive shaft and a cam supported by the camshaft for bearing against a valve actuating member, wherein the camshaft and the drive shaft have parallel, spaced apart fixed axes of rotation and are interconnected by a drive link one end of which is coupled to the drive shaft and the other end of which is coupled to the camshaft, means being provided for varying the spacing between an end of the link and the axis of rotation of the shaft to which that end is coupled l rius, the variable valve timing system of the present invention eliminates the need to move the axis of the central drive shaft, or to move the camshaft.

Preferably, an end of the link is pivotally connected to the camshaft.

Preferably an end of the link is pivotally connected to a drive member mounted on and moveable with respect to the drive shaft axis.

Preferably, the drive shaft extends through the camshaft and the drive member is engaged by a control shaft provided in and rotatable with the drive shaft, the drive member being coupled to the control shaft such that a displacement of the control shaft relative to the drive shaft causes a radial movement of the drive member. The drive member is preferably engaged with the control shaft such that axial movement of the control shaft causes the radial movement of the drive member, preferably by means of an inclined key defined by the control shaft, the drive member defining a slot engaged with the inclined key such that axial movement of the control shaft causes the drive member to travel along the inclined key.

The link may be coupled to the drive shaft and camshaft by needle bearings.

As the camshafts and drive shaft are located in fixed bearings, there are no movable members supporting the camshafts or drive shaft. There is also no intermediate member between the camshaft and drive shaft, and therefore no movable members to support such an intermediate member. There is further no separate control shaft that would need to be located and supported parallel to the camshafts and drive shaft. The present invention thus provides a mechanism that is simple, less costly to produce, more compact and more robust than prior art systems.

According to a second aspect of the present invention there is provided a variable valve timing system for an internal combustion engine comprising a hollow camshaft through which a drive shaft extends, the camshaft being coupled to the drive shaft by a drive link extending through an aperture in the camshaft, wherein the camshaft is assembled from two subassemblies that define opposite sides of the aperture and the link is coupled to a pivot pin which extends into both of the subassemblies.

The provision of two subassemblies to define the camshaft means that the camshaft may be made by a sintering process, as there is no need to separately bore through the camshaft, nor to machine slots or undercuts in the camshaft. This provides ease of manufacture and greater strength than with prior art devices.

The camshaft subassemblies may be connected together by means of a spigot extending from one subassembly into a socket defined in the other subassembly. The subassemblies may be welded together.

The subassemblies are arranged so that if the two camshaft halves are pressed together and not electron beam welded, the pin which connects both cam halves also prevents relative angular movement between the cams.

According to a third aspect of the present invention, there is provided a variable valve timing system comprising at least one camshaft and a drive shaft extending parallel to the camshaft, wherein the camshaft is coupled to the drive shaft by a link one end of which defines at least two spans through which a pin mounted on the camshaft extends, each span being received between a respective pair of pivot pin supports defined by the camshaft. Given that each span of the link can be relatively short, relatively low bending stresses are applied to the pin, which means a much smaller pin can be used. The link may be a single member defining all the spans, or two or more members each defining one span.

The reduction in pin diameter also enables the camshaft to be used easily where the cams are designed to activate valve levers and rockers. This is because levers or rockers have a lever ratio such that cam lift is much less than valve lift so that adjustment of the lever pivot point can adjust valve clearance. However, because of the reduced cam lift, the cam nose must be smaller so that there is a reduced space available for the placement of a spindle at the cam nose.

According to a fourth aspect of the present invention, there is provided a system for controlling displacement of a valve, comprising a cam shaft, a cam supported by the cam shaft, a first valve actuation lever supporting a first roller which bears against the cam, a second valve actuation lever supporting a second roller which bears against a surface defined by the first valve actuation lever, the second valve actuation lever being located between the first valve actuation lever and the-valve such that displacement of the first valve actuation lever by the cam causes displacement of the second valve actuation lever and displacement of the second valve actuation lever causes displacement of the valve, and means for causing a relative displacement between the first valve actuation lever and the cam such that valve displacement is a function of the said relative displacement, wherein the first valve actuation lever is mounted to pivot about a pivot axis and the relative displacement causing means is operative to displace the pivot axis relative to the cam along a predetermined path, the said surface against which the second roller bears is defined by a portion of the first valve actuation lever which extends between the pivot and the first roller, and the shape of the predetermined path and of the said surface against which the second roller bears are such that valve timing phase and valve displacement are a function of the displacement of the pivot axis.

The first valve actuation lever may be pivotally connected to a crank on a valve lift control shaft, rotation of the control shaft causing displacement of the pivot axis of the first valve actuation lever. The second valve actuation lever may be supported by a valve clearance adjustment mechanism.

The above valve displacement control system has the advantage that it does not require a spring mechanism to maintain the first valve actuating lever in contact with the cam, thereby saving space and reducing energy losses.

It will be appreciated that any two or more of the four aspects of the present invention as set out above could be combined. In particular, the first and fourth aspects of the present invention could be combined so as to provide the capability of varying both valve timing and valve lift.

Embodiments of the invention will now be described with reference to the accompanying drawings, in which: Figure 1 is a partial cross-sectional view of a cylinder head incorporating a variable valve timing system in accordance with the present invention; Figure 2 is a cross-sectional view through one of the camshafts of Figure 1 and a part-sectional view of an inlet cam acting on a bucket tappet in the valve opening position at high engine speed; Figure 3 is a cross-sectional view through one of the camshafts of Figure 1 and a view of an inlet cam acting on a bucket tappet in the valve opening position at low engine speed; figure 4 is a cross-sectional view through one of the camshaits oi Figure 1 and a part-sectional view of an inlet cam acting on a bucket tappet in the valve closing position at high engine speed; Figure 5 is a cross-sectional view through one of the camshafts of Figure 1 and a view of an inlet cam acting on a bucket tappet in the valve closing position at low engine speed; Figure 6 is a graph of valve lift plotted against crankshaft degrees for an inlet valve; Figure 7 illustrates an alternative link arrangement to that shown in Figure 1; Figure 8 is a section through a camshaft shown in Figure 7, also showing a drive shaft extending through the camshaft; Figure 9 illustrates an alternative camshaft structure to that of Figures 7 and 8; Figure 10 illustrates a variable valve lift system in accordance with the present invention, the illustrated valve being shown in the fully closed position; Figure 11 is a view of the assembly of Figure 10 with the valve in the fully open position; Figure 12 is a view of the assembly of Figures 10 and 11 after adjustment to the valve lift control mechanism; Figure 13 is a part sectional end view of components shown in Figure 10 to 12; and Figure 14 is a partial cross section through a cylinder head incorporating the variable valve lift mechanism of Figures 10 tol3 and a duration variable cam shaft as previously described with reference to Figures 1 to 9.

Referring to Figures 1 and 2, a cylinder head is illustrated, and has many conventional features which it is considered do not need detailed description since they are well understood by persons skilled in the art. A pulley 1 is driven by an engine crankshaft by means of a belt or chain (not shown). The pulley 1 is attached to a hollow drive shaft 2 having an axis A1 and a control shaft 3 in its centre. An inclined key 4 is machined in the control shaft 3 and is engaged with a slot in a drive block 5 which carries a pin 6 that is attached to one end of a link 7. The axis A2 of pin 6 is parallel to the axis Al of drive shaft 2. The other end of link 7 is attached between two cams 8 and 9 by means of a pin 10 having an axis A3. One set of cams 8 and 9 (and hence drive blocks 5, links 7 etc) is provided above valves 11 and 12, and another set of cams 8a and 9a is provided above valves 13 and 14. A bucket tappet 15 is provided on each valve to provide a cam actuating surface. The set of cams 8 and 9 above valves 11 and 12 are placed at a different rotational position around the drive shaft 2 to the other set of cams 8a and 9a above valves 13 and 14. Cams 8a and 9a will not be further described, as they are identical to cams 8 and 9.

Drive shaft 2 is supported on fixed bearings 16 and 17, the bearings being split along the horizontal centreline. A control cylinder (not shown) is connected to control shaft 3 via a thrust race (not shown) to enable longitudinal movement of control shaft 3 within the hollow centre of drive shaft 2. The cylinder head and split caps are also machined to provide bearings for the camshafts.

Cams 8 and 9 are spigotted together and form a camshaft that is provided around drive shaft 2, the pin 10 preventing relative movement of the cams. In addition, cams 8 and 9 could have a further screw or rivet to supplement pin 10. The together. The material and method of machining the cams is dependent on the type of cam follower used.

In the case of cams operating rockers or levers with curved pads or bucket tappets with curved or flat tops, the cams are made from a material suitable for nitride hardening. After the cams are spigotted together and electron beam welded, the bearing journals and the cam profiles are ground, holes in the cam sides are machined and then the cams are nitride hardened.

In the case of cams designed to operate valve levers with hardened roller followers, the cams are made from a material suitable for case hardening. Each cam is in this instance initially carburized only, then a machining operation is used to remove the carbon from the sides of the cam. The spigot diameter is then machined, which removes carbon from the area of electron beam welding. The two cams are then spigotted together and electron beam welded. After electron beam welding, the cams are reheated and quenched. Finally, the cam bearing journals and cam profiles are ground and the holes in the sides of the cams are machined.

The link 7 may be a one-piece link member, or may be two individual link members, or a hybrid link as shown. The link could be made from sintered iron or an SI iron precision casting and could run on a pin which was nitride hardened. As oil would be coming out of the camshaft bores, the underside of the link could have oil drillings into the link bores. Such a material combination with oil lubrication would give long life especially having regard to the load and angular movement of the link cycle.

The camshaft defined by cams 8 and 9 is supported in fixed bearings such that its axis A4 is offset or eccentric relative to the longitudinal axis Al of the drive shaft 2. Axial movement of the control shaft 3 causes block 5 to travel along inclined key 4, thus moving the end of link 7 that is attached to block 5 via pin 6 nearer to or further away from the drive shaft axis Al. As the axes A4 of the camshafts and Al of the drive shaft are offset, this causes a variation in angular velocity of the camshaft during each rotation with respect to the angular velocity of the drive shaft. The maximum variation in camshaft angular velocity occurs when the end of link 7 attached to block 5 is nearest to the axis of rotation of the drive shaft, and this occurs at low engine speed. The minimum variation in camshaft angular velocity occurs when the end of link 7 attached to block 5 is furthest away from the centre of the drive shaft axis, which occurs at high engine speed, and when the cam is profiled to give the correct high speed trajectory to the valve when the link is in this position.

Referring now to Figures 2 to 5, cams 8 are shown in a variety of configurations. Cams 9 are mirror images of cams 8 and will not be further described.

Figures 2 and 3 show an inlet cam at its valve opening position at high and low engine speed positions respectively. It can be seen that in the high engine speed position (Figure 2), drive block 5 has been moved radially outwardly by means of the axial movement of control shaft 3, thus causing the end of link 7 attached to drive block 5, and hence axis A2, to be moved further from the axes of rotation of the camshaft and the drive shaft 2. In the low engine speed position (Figure 3), drive block 5 has been moved radially inwardly by means of control shaft 3. Figures 4 and 5 are similar to Figures 2 and 3, and show the inlet cam at its valve closing position at high and low engine speeds.

Figure 6 shows a diagram of valve lift for an inlet valve plotted against crankshaft degrees for a gasoline engine. Curve H shows the high engine speed results and curve L shows the low engine speed results. At high engine speed, with the end of link 7 that is attached to block 5 in the maximum radially outwards position, the inlet valve is opened at 12° BTDC (before top dead centre) and is closed at 60° ABDC (after bottom dead centre). At low engine speed, with the end of link 7 attached to block 5 in the minimum radial position, the inlet valve is opened at 4° BTDC and is closed at 19° ABDC. In the case of an exhaust valve, a greater range of variation in timing will be given to the valve opening.

It should be noted that an increase in cam nose speed is caused during the opening and closing of the valve when the end of link 7 attached to block 5 is furthest away from the centre of the drive shaft. This does not, however, give a reduction in cam life. The reason for this is that as this is an event timing changer, the duration of valve timing is increased to match the highest operating speed of the engine, and if the valve lift is either kept the same or increased slightly compared to a conventional fixed timing camshaft, then the valve accelerations can be decreased. This lowers the maximum loads and lowers the valve spring loads which compensates for the increase in cam speed. As an example, a timing of 10 (inlet opening) 4° BTDC In (inlet closing) 52° ABDC with a fixed timing camshaft is altered to 10 12° BTDC IC 60° ABDC at maximum engine speed with the variable camshaft. Another advantage of the variable speed cam is that the cam nose radius is increased over prior art devices and this lowers the cam nose contact stress.

As the engine speed is increased, and hence the camshaft driving torque is increased, the end of link 7 attached to block 5 is moved to a larger radius position, thus reducing the load acting on the link bearings. The reduction in load means that needle bearings may be used in this end of link 7. The needles in such a bearing would rotate in one direction when under load so that oscillation takes place when the cam is not under load, i. e. when it is operating inlet valves only or exhaust valves only. The driving shaft eccentricity and other link parameters and the range of timing given by these parameters are based on very low pressure x velocity (PV) figures for the link mechanism at all engine speeds up to a maximum during the opening flank and nose periods and lower PV figures during the valve closing flank and nose periods.

It should be noted that the design shown in Fig. 1 is for the invention applied to an engine with the cylinders in-line and with the camshaft drive at the front of the engine. For this reason, the central driveshaft runs through the bores of hollow camshafts. The invention can be applied to different configurations of engines such as a single cylinder engine. In this case the camshaft could be solid, and the drive member, which would be a sprocket, fixed parallel to the camshaft but offset. The sprocket would be connected to the camshaft by an eccentric linkage. With such a design the control shaft could go through the bore of the sprocket assembly to move the link radially. It should also be noted that in such a design the link which is attached to the camshaft could be moved radially to vary the valve timing. In this case the camshaft could be made hollow but it would be hollow to receive the control shaft and not any drive shaft. Variations on this design could be made for multi cylinder engines where the drive arrangement is through the centre of the engine as in motor cycle engine design.

The design shown uses a link to drive the camshafts from a drive member.

The invention also can use a block mounted on a pin with the block running in a slot in the other member. In this case the centre of the block would be moved radially to vary the valve timing. Both the link design and the block design could use a solid shaft moved axially to move the link or block radially. The solid shaft would have an angled surface to move the link or block radially. The drive pulley or sprocket would be located axially in this case.

It should be noted that there are a number of commercial phase changers on the market whereby axial movement of a piston or control rod causes rotational movement of the camshaft sprocket or pulley by a helix mechanism. The axial movement of control rod 3 in Fig. 1 can also actuate one of these mechanisms to combine phase change with the duration change shown in Fig. 6.

In the system illustrated in Figure 1, the link 7 is a one-piece member defining two spaced apart spans through which the pin 10 extends. The pin 10 is received at each of its ends in a pivot pin support bore defined by the camshaft and passes through a central pivot pin support located between the two spans of the link. Such an arrangement makes it possible to use relatively small spans to achieve relatively low pin bending stresses. This in turn makes it possible to use a relatively small diameter pin.

As an alternative to the single piece link 7 of Figure 1, two or more parallel links may be used to couple the drive shaft to one camshaft. Such an arrangement is shown in Figures 7 and 8. Referring to Figures 7 and 8, the illustrated arrangement comprises cams 18 and 19 which may be manufactured by sintering. The interface between the two cams is indicated by line 20. A spindle 21 is coupled to a member 22 which is attached to drive shaft 23. Two links 24,25 are connected to spindle 21.

The links 24,25 are coupled to a spindle 26. The ends of the spindle 26 extend into support bores defined in the camshaft and through a central support 27 which is also defined by the camshaft. Thus the span of each link (represented by arrow 28) is relatively small which makes it possible to use a relatively small pin for spindle 26.

The spindle 26 prevents relative rotational movement between the two cam halves.

In the arrangement of Figures 7,8 and 9, variable valve timing can also be achieved by relative movement between the camshaft and drive shaft, rather than by varying the characteristics of the linkage between the two shafts as in the case of the embodiment of Figures 1 to 6. In the case of Figure 8, the two members could be flanged in the centre and riveted together if the cams were operating lever-type cam followers. A system incorporating a displaceable drive shaft is described in British Patent No. GB2066361.

Figure 9 shows two cams 18 and 19 spigotted together. In contrast to the arrangement of Figures 7 and 8, cam 18 is extended to engage in cam 19 so that the joint 20 is not at the centre as shown in Figure 8.

Referring now to Figures 10 to 14, a further embodiment of the invention is illustrated which incorporates a variable valve timing system as described with reference to Figures 1 to 6 and in addition incorporates a variable valve lift system in accordance with the invention.

Referring to Figure 10, the illustrated cam shaft assembly is identical to that of the embodiment of Figures 1 to 6 and incorporates a driving shaft 29 which is coupled to a cam 30 by a link 31 pivotally connected to the drive shaft 29 by pin 32 and pivotally connected to cam 30 by pin 33. The axis of pin 32 is displaceable radially relative to the axis of drive shaft 29 by the same mechanism as that illustrated in Figures 1 to 6 although that mechanism is not illustrated in Figure 10.

In contrast to the arrangement of Figures 1 to 6 in which the cam 8 bears on a tappet 15 mounted directly on the valve 11, in the arrangement of Figures 10 to 14 a valve 34 is moved as a result of rotation of the cam 30 by a lever system interposed between the cam 30 and the valve 34. That lever system comprises a first valve actuating lever 35 supporting a roller 36 which bears against the cam 30 and a second valve actuating lever 37 supporting a roller 38 which contacts an arcurate surface 39 defined by the first valve actuating lever 35. The first valve actuating lever 35 is pivotally mounted on pin 40 supported on a crank 41 extending from a valve lift control shaft 42, that control shaft being rotatable about axis 43. The second valve actuating lever 37 is supported on trunnions 44 (see figure 13) mounted on a body 45 supported on a hydraulic valve clearance adjuster rod 46, the rod 46 being displaceable in the direction of axis 47 to maintain the roller 36 in contact with the cam 30. It will be appreciated however that the axial position of the rod 46 does not change in response to rotation of the cam 30.

The end of the second valve actuation lever 37 remote from the trunnions 44 bears against a tappet supported on the end of the valve 33. Thus as the cam 30 rotates the lever 35 is caused to rotate about pin 40 and as a result of the contact between roller 38 and surface 39 the lever 37 is caused to rotate about the trunnions 44. Rotational movement of the cam 30 is thus converted into axial displacement of the valve 34. The timing of displacements of the valve 34 in terms of the angular position of the cam 30 can be adjusted by displacing the pin 32 relative to the axis of the drive shaft 29 as in the embodiment described in Figures 1 to 6. In addition however the magnitude of valve displacements can be adjusted by displacement of the lever 35 relative to the cam as a result of rotation of the shaft 42 about axis 43.

Figure 10 shows the system with the valve 34 fully closed and the pin 32 at its maximum radial distance from the axis of drive shaft 29.

Figure 11 shows the same system after rotation of the cam to a valve fully open position and it will be seen that the head of the valve 34 has moved a substantial distance away from the valve seat 48. Figure 12 shows the same system after displacement of the pin 32 towards the axis of the shaft 29 and after rotation of the shaft 42 about axis 43. The system is shown in Figure 12 with the valve as far open as is possible given the position of the shaft 42 and it will be noted that the spacing between the head of the valve 34 and the valve seat 48 in Figure 12 is far less than in the case of Figure 11. Thus not only will the timing duration have been adjusted in a manner similar to that illustrated in Figure 6 although of less duration but in addition the magnitude of valve opening will also have been adjusted. It can be seen that the contact point 49 between roller 38 and surface 39 has been substantially displaced as between the system configuration as shown in Figures 10 and 11 and the configuration shown in Figure 12. The shape of surface 39 is such that the roller 36 remains in contact with the cam 30 despite rotation of the shaft 42.

The described arrangement provides the capability of varying the duration of inlet valve opening in combination with the capability of varying the magnitude of valve displacement and can also be arranged so as to provide a reduced inlet valve opening duration at low engine speed but with full valve lift so that engine torque can be maximised for drivability. The extra torque can also be used with automatic gear changing systems so that a higher gear can be selected in"economy"engine mode so as to provide additional fuel consumption advantages in addition to part load advantages due to a reduction in engine pumping losses. The variable valve lift system in addition does not introduce substantial frictional losses because of the use of rollers between the levers and the surfaces that they contact. Substantial parasitic mass is not added to the valve assembly and therefore substantial increases in valve mechanism loadings and engine wear are not caused. Furthermore the maximum possible valve lift is not reduced so that power at high engine speed is not reduced.

As described above, in the illustrated arrangement valve timing can be adjusted by adjusting the radial distance between the axis of drive shaft 29 and pin 32.

Separately, valve lift can be adjusted by rotating the valve lift control shaft 42 about axis 43, such rotation moving the lever 35 so that the leverage ratio is changed between the pivot pin 40, the roller 36 which acts against the cam 30, and the point 49 on the curved surface 39 which contacts the roller 38. Both the valve timing and valve lift adjustment mechanisms illustrated in Figures 10 to 14 have been adjusted as between the configuration shown in Figure 11 and the configuration in Figure 12 and thus the effects of the two mechanisms are combined. If however only the valve lift mechanism was adjusted as a result of rotation of the shaft 42 without any change in the distance between the axis of the drive shaft 29 and the pin 32, there would still be some timing change associated with the change in valve lift as rotation of the shaft 42 would cause the roller 36 to roll around the cam 30 such that the point of contact between the roller 36 and the cam 30 would be adjusted as a result of adjustment of the valve lift. Thus adjustment of the valve lift produces a valve timing phase change.

The mechanism described with reference to Figures 10 to 14 has certain advantages. The valve actuation lever arrangement is very similar to a conventional valve actuation lever where a roller follower normally contacts the cam. In the described arrangement where valve actuation lever 37 supports a roller 38 in contact with the surface 39, it will be noted that the roller 38 contacts a concave surface (surface 39) which considerably lowers the contact stresses between these components as compared with contact between a roller and the convex face of the cam. In addition, the roller 38 is subject to only small rotations. The roller 36 which is supported on valve actuation lever 35 can have a relatively large radius from the lever pivot axis as compared with conventional arrangements and this means that the load on this roller and on the cam is substantially reduced compared with conventional arrangements where a valve actuation lever bears against both the valve to be controlled and via a roller directly against the cam. Thus the described arrangement has a high mechanical efficiency and does not impose any engine speed limits or problems of increased wear as compared with conventional arrangements.

The advantages of being able to vary the duration of valve opening so as to maximise volumetric efficiency particularly at low speed by adjusting the spacing between the axis of shaft 29 and pin 32 can be substantially retained when used in conjunction with the variable valve lift mechanism. For example, a normal inlet valve timing of 10 5° BTDC and IC 55° ABDC at high engine speed with full valve lift could be altered to 10 5° ATDC and IC 10° ABDC with full valve lift at low engine speed for substantially improved torque and drivability. For part load conditions the inlet valve timing could be reduced to 10 85° ATDC and IC 35° ABDC at minimum valve lift required when the engine is idling. As the valve lift is increased, the timing duration could be retained but the increased valve lift would introduce a phase change. This phase change brings the maximum opening point of the valve nearer to the maximum piston velocity and itself would produce an increase in cylinder filling so this action is complimentary to the associated increase in valve lift. These two parameters are also involved in the velocity of the incoming gas which is important in fuel and air mixing. For example, the above inlet valve timings would be rephased to 10 60° ATDC and IC 10° ABDC at half full, valve lift.

The valve lift could be controlled directly from a vehicle accelerator pedal with full valve lift causing a valve timing duration change so as to maximise volumetric efficiency and there could also be a valve timing duration change from the minimum valve timing duration and above half full valve lift at engine speeds above 3000 rpm for example. The cam profile whether used for duration variable valve timing or used in conjunction with the variable valve lift mechanism may be profiled to give the correct valve trajectory at high engine speed and full valve lift with the cam rotating at the variable angular velocity produced with the block 5 at its furthest distance from the centre of the driveshaft.