TECHNICAL FIELD

The invention relates to a suspension damping system, in particular, a suspension damping systems for vehicles, including for application to fork type suspension assemblies and spring-damper units for vehicles for both on-road and off-road use. More specifically, the invention is concerned with a damping control arrangement in such a suspension system.

BACKGROUND

Suspension systems for vehicles, such as of the hydraulically damped type, characterized by a tubular body containing oil which is acted upon by a piston arrangement, typically employ valves and one or more orifices through which oil flows between chambers within the body. Damping of suspension elements, whether during initial compression or during subsequent rebound due to operation of associated spring elements of the suspension system, may be selected by choice of oil viscosity and/or orifice size. Sometimes hydraulic damping is combined with another fluid in gaseous form, such as nitrogen, which acts in unison with a cooperating hydraulic oil circuit. For example, in mono-tube dampers, gas is retained in a damper tube by a secondary floating piston upon which opposed faces the oil and gas charges respectively act. In motorcycle suspensions, a widely employed technique for controlling oil flow through such valves is the use of resilient obturators associated with valve passages which flex in response to impinging oil, particularly where the suspension is required to cope with a sudden, large excursion of suspension elements as can be encountered during off-road riding. These valves are typically arranged in parallel with a fixed size orifice and provide a pressure relief function. The relief valve may include a plurality of such obturators, for example in the form of discs (sometimes referred to as "shims"), which are arranged in a stack to provide varying resistance to flow as pressure of impinging oil increases. The shims can be of differing thickness and/or of differing diameter within the stack, to provide a degree of (fixed) damping rate variation.

It will be appreciated, that relief valves within the suspension system may be individually tuned by selecting the thickness, diameter and number of shims within a given "stack." For example, and considering off-road riding applications, in motocross style riding the terrain encountered and motorcycle speed generally involves regular large suspension excursions necessitating pressure relief at relatively higher pressures thus calling for stiffer shim stacks. In comparison, enduro or trail style riding typically involves relatively smaller suspension excursions and/or adverse terrain taken at lower vehicle speeds calling for pressure relief at relatively lower pressures, particularly in front suspensions so as to preserve steering accuracy for the motorcycle on narrow trails or tracks and in rear suspensions to maintain traction over the relatively small excursions. The present Applicant has observed, particularly in regard to compression damping pressure relief, that the damping force exerted by prior art pressure relief valves (including of shim stack construction) is proportional to the suspension element excursion rate. Put another way, the damping force exerted by prior art pressure relief valves is typically constant over suspension excursion distance. The major influences on suspension excursion rate, other than the nature of terrain and available total suspension extension or "travel", are the mass of the motorcycle including the rider and the speed at which the terrain is approached.

This results in the suspension requiring tuning of damping with respect to the anticipated application, with consideration for both a rider's skill level, which is typically reflected in approach speed to adverse terrain, and his/her weight, which is often a significant proportion of the overall mass of a motorcycle when in use. Consider for example a leading endure-style motorcycle, such as Yamaha Motor Company's WR450F model, which has a wet mass of the order of 124kg, compared with a 50th percentile weight for adult male riders aged 20- 45 of 65-85kg. This results in a motorcycle gross vehicle mass having a typical range of approximately 190kg to 210kg. In addition, for optimal suspension performance, some suspension systems demand a relatively high terrain approach speed, which may present difficulty for non-expert riders.

Accordingly, it would be desirable to provide a suspension system wherein damping may be controlled independently of suspension excursion rate for example as caused by vehicle speed when traversing rough terrain, whilst being relatively insensitive to variations in rider weight.

Any references to methods, apparatus or documents of the prior art are not to be taken as constituting any evidence or admission that they formed, or form part of the common general knowledge.

It is an object of the present invention to provide a suspension damping system that exhibits improved damping performance, especially compression damping, over a wide variety of terrain or to provide a suspension damping system that addresses or at least ameliorates the problems of the prior art. SUMMARY OF THE INVENTION

In one aspect, the invention provides a suspension damping system for damping a sprung mass, the system including: at least one elastic support member coupled to the sprung mass; a first damping chamber containing hydraulic fluid; a reservoir chamber for hydraulic fluid urged from the first damping chamber; a piston assembly which extends into the damping chamber during a compression stroke and applies work on the hydraulic fluid as a compressive damping force; at least one compression damping valve which when open provides fluid communication between the first damping chamber and the reservoir chamber, said valve coupled to the at least one elastic support member; wherein, in use, the valve operates in accordance with the displacement of the elastic support member relative to the sprung mass.

Suitably, the work done on the hydraulic fluid by the piston assembly is a function of displacement of the piston assembly within the damping chamber; compression damping force is a function of the force provided by the elastic support member. Suitably, the compression damping force is a function of the ratio of the cross-sectional area of the piston assembly to the ratio of the cross- sectional area of the compression damping valve. Suitably, the fluid flow through the compression valve is regulated by the potential energy of the elastic support member and the pressure within the damping chamber.

Preferably, the compression damping force increases with a decrease in length of the elastic support member.

Preferably, the energy dissipated by the compression stroke is of a similar order of magnitude or greater than the energy dissipated during rebound.

Preferably, the piston assembly includes a sliding guide, slidably engaged with an inner wall of the damper chamber, said guide channeled to minimize flow restriction. Preferably the piston assembly includes a rebound check valve which is slidably engaged with an inner wall of the damper chamber, said check valve channeled to minimize hydraulic fluid flow restriction during a compression stroke of the piston assembly.

Suitably, the elastic support member is selected from one of more of: coil spring, leaf spring, air piston, or airbag.

In some embodiments, the distal end of the piston assembly is rigidly attached to a sprung mass. In other embodiments, a distal end of the piston assembly is rigidly attached to an un-sprung mass.

The sprung mass may be a vehicle chassis. Suitably, the vehicle chassis is selected from one of: motorcycle, automobile, truck and off-road vehicle.

In another aspect of the invention, there is provided a vehicle including a suspension system as set out above. The vehicle may comprise a motorcycle having a telescopic fork front suspension and/or a spring damper unit rear suspension incorporating the suspension system. The vehicle may comprise an automobile having a spring damper unit rear suspension incorporating the suspension system.

BRIEF DESCRIPTION OF THE DRAWINGS

Preferred features, embodiments and variations of the invention may be discerned from the following Detailed Description which provides sufficient information for those skilled in the art to perform the invention. The Detailed Description is not to be regarded as limiting the scope of the preceding Summary of the Invention in any way. The Detailed Description will make reference to a number of drawings as follows:

Figure 1 is a partial cross-sectional view of a fork type suspension assembly of an embodiment of the present invention;

Figure 1A is a full cross-sectional view of the fork type suspension assembly of Figure 1 on line D-D with some of the components omitted from the drawing for clarity;

Figure 2 illustrates a cross-sectional view of an exemplary embodiment of the present invention with break lines;

Figure 2a is an enlarged view of Figure 2 showing details of a preferred compression damping valve arrangement;

Figure 2b is an enlarged view of Figure 2 showing details of a preferred piston assembly;

Figure 2c is an enlarged view of Figure 2 showing details of a compression bypass valve subassembly;

Figure 3 is a cross-sectional illustration of the suspension damping system 10 with hydraulic fluid flows during compression indicated;

Figure 3a is an enlarged view of the compression damping valve, shown in Figure 3, in an open position;

Figure 3b is an enlarged view of the rebound check valve, shown, in Figure 3, in an open position;

Figure 4 is a cross-sectional illustration of the suspension damping system 10 with hydraulic fluid flows during rebound indicated;

Figure 4a is an enlarged view of the compression damping valve, shown in Figure 3, in a closed position; Figure 4b is an enlarged view of the rebound check valve, shown in Figure 3, in a closed position;

Figure 5 is a cross-sectional illustration of an alternative exemplary embodiment of the present invention with an external hydraulic fluid reservoir;

Figure 6 is a cross-sectional illustration of another alternative exemplary embodiment of the present invention with remote suspension spring arrangement;

Figures 7a - 7d are cross-sectional illustrations of varying valve shapes; Figure 8a and 8c are orthogonal illustrations of sliding guide 80;

Figure 8c is a three-dimensional illustration of sliding guide 80;

Figure 9a is a three-dimensional illustration of the second collar of the compression damping valve system;

Figure 9b is an orthogonal illustration from above of the second collar of the compression damping valve system;

Figure 9c is a three-dimensional illustration of the first collar viewed from above;

Figure 9d is a side illustration of the first collar;

Figure 9e is a three-dimensional illustration of the first collar showing the underside of the collar;

Figure 10a is shows a plot of force versus displacement for a conventional shock compared to a shock of the present invention, with compression and rebound bypass valves open and adjusted; and

Figure 10b is shows a plot of force versus displacement for a conventional shock compared to a shock of the present invention, with compression and rebound bypass valves closed.

DETAILED DESCRIPTION OF PREFERRED EMBODIMENTS

Referring now to Figure 1 , there is depicted a partial sectional view of a telescopic fork type suspension assembly 10 for a vehicle, such as may be employed, for example, in a front suspension of a motor cycle. In the drawing, a first suspension section in the form of an outer fork tube 12 has been cut away to expose the internal components, including a conventional resilient or elastic support member in the form of coil spring 14 and a damper sub-assembly 16. A second suspension section in the form of a cooperating fork leg, within which the outer fork tube 12 is telescopically engaged, is omitted from the drawing figures for clarity. The omitted fork leg, sometimes referred to as a "slider" in conventionally arranged telescopic fork suspensions, usually allows attachment at a distal end thereof, a front wheel and associated brake. In the less common inverted or upside-down (USD) fork suspension, paired slider portions are fixed to a steering head or yoke of a motorcycle chassis, and the fork tubes or stanchions include a compact bracket provided at a distal end of the tubes for carrying the front wheel.

The telescoping outer fork tube or stanchion and fork leg or slider together provide an enclosure for a damping fluid 18, typically in the form of oil of a selected viscosity or "weight". This fluid reservoir of the suspension damping system typically includes a headspace 17 containing air or other inert gas above the hydraulic fluid 18. If desired, the suspension system may be modified to include a piston or seal (not shown) between the hydraulic fluid and the gas in the headspace. This would, for example, allow for inversion, in use, of the embodiments shown in Figures 1 , 1A and 2. Such a modification would require moving the spring 14 further down the suspension assembly towards the damper distal end.

Turning to Figure 1A, the suspension assembly 10 has been fully sectioned in order to illustrate internal components of the damper sub-assembly 16, which has a tubular body 16a, a distal damper end 16b and a proximal damper end, or damper head 16c. In the illustrated orientation, the distal end 16b is located at the bottom of the damper sub-assembly 16 and the proximal end 16c at the top of the damper sub-assembly. The suspension assembly contains a charge of hydraulic fluid, or oil 18 that is driven by a piston assembly which comprises a piston sub-assembly 22 mounted on a distal end of piston rod 20. A distal end of the piston rod may be attached to, for example, a ground engaging member such as a wheel, or suitably, inverted and attached to a vehicle chassis. The piston assembly reciprocates within damping chamber 26 and applies a compression damping force to the hydraulic fluid in chamber 26 during a compression stroke. In the current embodiment, the compression stroke is the piston assembly extending into the damping chamber 26 towards the distal end 16b. A portion the piston subassembly 22 (see Figure2) is in slidable engagement with an internal wall of the damper sub-assembly 16. A sealing bearing 24 is provided where the damper rod 20 enters the proximal damper end 16c. It will be appreciated that the piston rod 20 and piston sub-assembly 22 are movable relative to the damper sub-assembly 16 as a result of, for example, movement of a wheel, carried by the suspension assembly, relative to a vehicle chassis.

Within the suspension assembly 10 are a number of chambers including a first chamber, or damping chamber 26 contained within the damper body 16a, and a second chamber, or reservoir 28 external of the damper but, in the present embodiment, contained within the outer fork tube 12. The damper subassembly 16 further includes an internal piston guide tube 36 co-axial with the damper body 16a. An inner wall of the piston guide tube 36 provides a guide surface that is slidably engaged by a portion of the piston sub-assembly 22. In the present embodiment, the guide tube 36 is secured within the distal damper end 16b and receives the proximal damper end 16c at an opposed end. The proximal damper end 16c includes a reduced diameter externally threaded portion 16d adjacent to its reception by the guide tube 36, also providing an internal bearing surface for the piston rod 20.

Turning to damping chamber 26 of the damper sub-assembly 16, a plurality of radially extending ports 36a are provided in the piston guide tube 36 adjacent to the distal damper end 16b. The ports 36a allow oil 18 to flow, upon urging by piston 22, from the damping chamber 26 within the guide tube 36 into an annular intermediate chamber section 26b that lies between the guide tube 36 and the damper body 16a. The annular intermediate chamber 26b forms part of the structure of the primary compression damping action of the suspension assembly and chamber 26b is pressure equalized with at least a portion of the damping chamber 26. A distal end of the tubular damper body 16a is received and supported by an internal circumferential channel 16e of the distal damper end 16b, which internal channel is located adjacent to the radial ports 36a in the piston guide tube 36.

The reservoir chamber 28 is bounded by the outer fork tube 12 and cooperating fork body or stanchion (not shown), outwardly of the damper sub-assembly 16. A compression damping valve or pressure relief valve 30 is provided between the first chamber 26 and the second chamber 28 of the suspension assembly 10. In the embodiment, the damping valve 30 is located on the proximal end of the damper sub-assembly 16, and includes a first collar 32 slidably mounted on an exterior surface of the proximal damper end 16c and a second collar 34 fixed to an adjacent proximal end of the damper body 16a. The first collar 32 of valve 30 is retained by an elastic supporting member of the suspension, here in the form of the coil spring 14, said spring for supporting a sprung mass or body, such as a vehicle chassis. In the embodiment, the coil spring 14 is coupled to the valve 30 by provision of a circumferential shoulder portion 32b of the slidable collar 32. In the present embodiment, the first collar 32 and second collar 34 are provided with cooperating frusto-conical surfaces 32a, 32b which, when moved apart by relative axial movement of the collars, are able to vary the size of an annular valve orifice or throat provided therebetween.

Turning to Figures 2, 2a, 2b and 2c, an exemplary embodiment of the present invention is further illustrated, including the enlarged sections illustrated by Figures 2a, 2b and 2c describing various sub-assemblies of the embodiment. In a similar manner to Figure 1 a, the suspension assembly 10, in Figure 2, has been sectioned in order to show internal components. To curtail the overall length of the part of the suspension assembly illustrated in Figure 2, break lines have been added to the view. In operation, the suspension damping system contains a working, or hydraulic fluid, such as a hydraulic oil 18. In the embodiment shown in Figure 2, the damper sub-assembly 16 has a damper head 16c, a damper body 16a and a damper distal end 16b. The damper piston rod 20 passes through the head 16c of the damper sub-assembly 16. The outer wall of rod 20 is sealingly engaged with bearing 24 located at top of the damper head 16c. The piston assembly includes the piston rod 20 and the piston subassembly 22. The piston subassembly 22, located at a distal end of piston rod 20, comprises a high fluid flow sliding guide 80, rebound check valve 94, shimstack 99, and piston head 95 in its sub-assembly, the significance of which will be discussed with reference to Figure 2b. In the illustrated embodiment, the damper sub-assembly 16, main suspension spring 14 and a portion of the hydraulic fluid 18 are contained within the outer fork tube 12. The compression damping valve or pressure relief valve 30 is located at the damper head end 16c of the damper assembly 16. The compression valve system 30 will be discussed in greater detail with reference to the enlarged section illustrated by Figure 2a. The piston guide tube 36, which bounds damper chamber 26, encapsulates: piston 22, top-out spring 46 and the optional spring retainer 49. In the currently described embodiment, the piston tube 36 has been internally machined to increase the cross-sectional area of the tube in order to house a top out spring 46 of higher spring rate and greater diameter. Optionally, and depending on piston tube internal diameter, a typical floating top out spring may be used without the need for internal machining of tube 36 to increase the diameter, or for that matter, without the need for use of a spring retainer 49. In the currently described embodiment, the annular intermediate chamber, or plenum section 26b, is formed between the outer wall of the piston guide tube 36 and inner wall of the outer fork tube 12. The plenum section 26b is in fluid communication with damping chamber 26 via ports 36a located proximal to the distal end 16b of the damper, towards maximum possible or "top out" travel of piston 22 during a compression stroke. The plenum section 26b is also in fluid communication with the compression valve 30 (which as mentioned the particularities of which will be discussed with reference to Figure 2a), as well as a compression bypass needle valve 50 (also known as a slow speed bypass), which is located within the distal damper end 16b, and in the present embodiment, is located within an external housing 52, which is secured by threaded fastening to outer fork tube 12. The operation of the compression damping valve 30 will now be discussed in relation to the enlarged view of the damper head 16c and compression damping valve 30 as shown in Fig. 2a. The piston rod 20, which reciprocates within chamber 26, has a diameter indicated by the double headed arrow 20c and an external surface 20a, which sealingly engages with the inner surface 24a of the bearing 24, which itself is secured to the damper head 16c. The reduced diameter portion 16d is also attached to the damper head 16c, and a surface of the portion 16d is exposed within chamber 26 and forms a retainer 16f for spring 46. The piston rod 20 may optionally be a hollow tube to reduce its mass. Primary components of the pressure relief valve 30 system are first and second mating collars 32 and 34 respectively. The first collar 32 is retained by an elastic body, such as a suspension spring 14, which applies a force against the first collar 32 onto second collar 34.

The force applied by the coil spring 14 to collar 32 is a function of the elastic potential energy stored within the spring due to partial compression of the spring by the sprung mass of, for example, the vehicle to which it is attached, which may include the chassis, drivers, passengers, riders (as the case maybe), or other mass associated with the sprung part of the vehicle. Typically, the sprung mass includes a chassis and everything rigidly connected to it. In the example shown, collar 32 acts as a movable valve plug part of the compression valve system 30, and is biased in the closed position, at least in part, by the above- discussed sprung-mass of the vehicle. The first collar 32 may include first and second shoulders, 32b and 32c respectively, which act as a guide for spring 14 to engage with collar 32. In the presently discussed embodiment, the disc-like collar 32 is slidably engaged around the damper head 16c. An o-ring 54 provides for sealing engagement. The second collar 34, which may be considered the complementary seat for collar 32 of the compression valve system 30, is attached to the un-sprung part of the vehicle, which in the presently described embodiment includes the damper head 16c, the damper body 16a and the damper distal end 16b. Typically, at least in relation to vehicles, the un-sprung mass refers to the ground engaging members, for example, wheels, and everything rigidly connected thereto. Suspension damping systems together with elastic bodies, such as springs, are typically used to couple the sprung portion to the un-sprung portion. Briefly returning to Figure 2, in the presently discussed embodiment, a pressure increase within damping chamber 26 is usually caused by the piston assembly, that is the rod 20 and associated piston 22, traversing towards the distal damper end in a direction indicated by the arrow 40. The length of travel, or displacement, of the piston assembly within chamber 26, is associated with a compression stroke, and concomitantly displacement, of the primary suspension spring 14.

A compression stroke extension of the piston, and concomitant spring compression may be caused, for example, by vehicle braking or from the vehicle traversing uneven terrain. In particular, the energy generated from changes in the wheel position of a vehicle relative to a hypothetical ground plane (e.g. dips and bumps), or changes in the centre of mass of the vehicle system resulting from changes in vehicle velocity (which can result in diving or squatting) is at least partially transferred into the suspension damper system (damping) and partially transferred into elastic potential energy in suspension springs (through spring compression). The energy transferred into the suspension damper system is preferably smoothly dissipated by transfer as heat in the working fluid, which in this case is the hydraulic fluid.

Staying with Figure 2, and by way of example, during a compressive stroke of sufficient intensity, at least some of the working fluid in damping chamber 26 is urged from the first, or damping chamber, 26, via apertures 36a, into a plenum section 26b, with a concomitant increase in fluid pressure in chamber 26b due to work done on the fluid by the piston assembly. This is the compressive damping force. Regarding now Figure 2a, in the currently discussed embodiment, the system working fluid in plenum 26b is in contact with a portion of the lower surface 32a of first collar 32 and the one or more channels 60a formed between an inner wall of the valve seat 34 and the outer wall of the piston guide tube 36. Together this arrangement is at equal pressure on the damping chamber 26 side of the relief valve system 30. Under sufficient pressure in chamber 26b, the working fluid works against the lower surface 32a of the valve plug 32 operating the valve and forcing the plug to move against spring 14, thus, if only slightly, and sometimes only momentarily, compressing spring 14 and opening an orifice between mating surfaces 32a and 34a. When the pressure within the first chamber 26 in insufficient to keep the valve open in light of the force currently applied to the first collar by the elastic body 14, the orifice closes and the valve operation ends. The energy required by the working fluid to actuate valve 30 is largely dissipated into the suspension damping system as heat.

The valve 30 acts as a type of pressure relief valve and closes when the force from the spring 14 is sufficient to overcome the force against collar 32 derived from the pressure in plenum 26b. The force from spring 14 against collar 32 increases as the piston extends into chamber 26. As such, the compression valve system 30 may be considered interactive inasmuch as with increasing compression of spring 14 there will be greater force applied to collar 32, and concomitantly with increasing spring compression there will be increasing extension of piston 22 in chamber 26 accompanied by increasing work applied against valve system 30 from the increasing pressure in chamber 30. Figure 2a indicates a solid, or blocked-out, portion at 60b whilst showing a circumferential channel portion at 60a. The blocked-out portion indicated at section 60b from section of the illustration of Figure 2, passing through an optional mounting guide 60b, which will be further discussed with reference to Figures 9a and 9b. The piston 22 will now be discussed in relation to the enlarged view shown in Figure 2b. An upper portion 64 of the piston sub-assembly 22 forms part of the piston shaft also partially houses the needle valve 70. Needle valve 70 may be adjusted to regulate hydraulic fluid flow through rebound bypass channel 86, said channel passing through the piston head 95. Channel 86 is terminated by a hollow screw 97, which during rebound provides a fluid flow path from subspace 78, through bypass aperture 74, bypass channel 86 and conduit 96, which is gated by a shim stack, into damping chamber 26. The screw 97 also holds in place the rebound shim stack 99. In the shown embodiment, the outer wall of the piston rod 20 is slidingly engaged with spring retainer 49, the upper end 66 of which is a seat for top out spring 46. Turning to the needle valve 70, spring 72 pushes the needle 70 back during adjustment. There is a shaft (not shown) that pushes onto the top of the needle valve with a screw at the other end of the shaft for adjustment. The sliding guide 80, which is slidingly engaged with the inner surface of piston tube 36, has a relatively high cross-sectional area channel 84 to minimize fluid flow restriction, which is formed between the sliding guide 80 and the piston head shaft 79 (see Figure 4b), to provide, during compression, for unrestricted around the piston head 95 and through the guide 80. Continuous with the channel 84, the sliding guide has a plurality of radially extending apertures 82 which direct fluid flow back out of the sliding guide 80. During rebound the flow direction with respect to guide 80 is reversed and fluid is directed into guide 80 through apertures 82 and out through channel 84. A seal 81 is also disposed on the sliding guide 80 for sealing engagement with the inner wall of piston tube 36. The sliding guide 80 is retained in place as part of the piston assembly between shoulder 75 of piston 22 and the piston shaft 20. Adjacent sliding guide 80 is rebound check valve 94 which also slides within chamber 36 and has an o-ring 92 for sealing engagement with the inner wall of piston chamber 36. During a compression stroke, the valve 94, which may be lightly biased by, for example, a spring, is disposed apart from a rear portion 91 of the piston head 95 providing a large cross-sectional area to maximise unrestricted fluid flow through channel 90 of the valve. During rebound, the valve 94 is disposed against the rear portion 91 of the piston head 95, blocking the flow of fluid around the piston head with channel 90 now aligning with, and directing fluid to flow through, rebound shim stack port 96 and the shim stack valve 99. Because of the unrestricted flow around the piston head 95 provided by the high fluid cross-sectional area check valve 94 and the high fluid flow through the guide/slider 80, it is estimated that the effective cross-sectional area of the piston assembly is approximately the cross-sectional area of the piston shaft 20.

Figure 2c shows an enlarged view of the distal damper end 16b of the embodiment illustrated in Figure 2. The damper chamber ports 36a provide fluid communication between the damping chamber 26 and the plenum 26b, the significance of which has been discussed elsewhere. The distal internal housing 102; which is screw fitted to both the piston tube 36 and the damper body 16a, and together with the piston tube 36 and damper body 16a forms a sealed end to plenum 26b, houses at least part of the compression bypass needle valve 128. The bypass needle valve 128 provides adjustable regulated flow between chamber 100, which is a distal part of the damper chamber 26, and the second outer chamber 28, by means of fluid connectivity between needle valve bypass passage 104, return channels 126 and 124, and apertures 122. The needle valve 128 is disposed within a valve housing 130, which is screw fastened within the distal internal housing 102. The location of the tip of the needle valve 120 may be adjusted to regulate fluid flow through passage 104. Return channel 129, formed between an inner wall of the distal external housing 52, and an outer wall of the distal internal housing 102, provides fluid communication between the second outer chamber 28 and apertures 122. An upper portion of the distal external housing 52 is secured to the outer walls of the outer fork tube 12. The compression check valve 108, is secured by the valve housing 130 and sealingly engaged within the internal distal housing 102 by an o-ring 118. The compression check valve comprises a shim 112 which is lightly biased shut by a spring 1 10 to block flow through channel 1 14 during compression. Channel 1 14 is in fluid communication with return channel 124. During a compression stroke the shims 112 are closed such that working fluid from chamber 26 is urged to flow through bypass passage 104 and additionally as necessitated, apertures 36a. During rebound, a pressure-drop across chamber 28 to chamber 106, and consequently chambers 104 and 26, which have low restriction fluid communication with chamber 106, opens shim 112 allowing fluid to return to damping chamber 26 from chamber 28.

Figure 3 is a simple illustration of hydraulic fluid flow, within the suspension damping system 10, during compression. The direction of travel, during compression, of piston rod 20 and piston sub-assembly 22 within piston tube 36 is indicated by arrow 40. The piston 22 drives a portion of the hydraulic fluid in the direction indicated by arrows 150. During compression, lightly biased rebound check valve 94 opens due to increasing fluid pressure within the damping chamber 26, allowing a high fluid flow rate around the piston 22. Check valve 94 is shown in the open position. The relatively high cross-sectional flow area of the sliding guide 80, similarly, minimises restriction of fluid flow around the piston 22. As mentioned, as a result of the light biasing of check valve 94, and the high cross-sectional flow area of guide 80, the piston cross-sectional area is approximately equal to the cross-section area of the piston shaft 20 (see double headed arrow 20c, Figure 2a). The path and direction of hydraulic fluid flow around the piston 22, through check valve 94, through slider 80 and into reservoir 78 is indicated by arrows 156. The reservoir 78 volume increases as the damping piston 22 extends into damping chamber 26. Some of the hydraulic fluid driven by piston 22 is urged into bypass passage 104 of needle valve 50 and through channels 126 and 124 into the hydraulic fluid reservoir 28. The cross-sectional area of this fluid flow path is regulated by needle 128.

The direction of the above-described fluid flow from damper chamber 26, through needle valve 50 into reservoir chamber 28, is indicated by the bifurcated arrow 166 with the arrow stem 164 providing an indication of fluid origin. From time to time, during light compression events, bypass needle valve 50 may provide adequate damping. During such events, the cross-sectional area for fluid flow provided by the needle valve 50 is sufficient, by itself, to accommodate the hydraulic fluid flow from piston compression. When greater force is applied to the working fluid 18 by the piston rod 20 and piston 22, the cross-sectional area of the needle valve is typically insufficient to accommodate fluid flow and fluid may be urged through ports 36a, directionally indicated by arrows 162, into plenum 26b and up towards, directionally indicated by arrows 154, the compression damping valve system 30, with which plenum 26b is in fluid communication. When sufficient force is applied by fluid 18, in the annular chamber 26b, the compression damping valve 30 is forced open, relieving pressure, and allowing fluid to flow through into reservoir 28, as indicated directionally by arrows 152. As discussed, in operation, compression valve 30 opens by way of valve collar 32 being lifted off valve seat 34 by the working fluid 18 in response to the compression damping force generated by the piston assembly in the damping chamber 26. Valve plug 32 is movable because it is slidably retained by a compressible elastic body, in this case, spring 14 around damper head 16c. The compression valve 30 is shown in an open position in Figure 3. Figure 3a is an enlarged illustration of valve 30 in an open position. The face of valve seat 34a and the profile of the face of valve plug 32a are shown in a separated position. The gap between faces 32a and 34a is indicated by arrow 168, the head of which is positioned within the gap. An enlarged view of the rebound check valve 94 in an open position is shown in Figure 3b. Due to fluid pressure, the check valve 94 has moved away from piston head 95 creating the space 160 to allow fluid to flow around the piston head 95 and through the sliding guide 80. The channeling 90 of the valve 94 aligns with the channeling 84 of the guide 80.

Figure 4 is a simplified illustration of hydraulic fluid flow within the suspension damping system 10 during rebound. The direction of travel during rebound of piston rod 20 and piston sub-assembly 22 within piston tube 36, is indicated by arrow 172. As the piston rod 20 and piston 22 withdraw from the chamber, a pressure drop is created between the subspace reservoir 78 and damping chamber 26 in the region below the piston head 95. The subspace reservoir 78 is a subspace of damping chamber 26. The reservoir 78 volume decreases as the piston rod 20 and piston 22 are withdrawn. During rebound the check valve 94 abuts a rear portion 91 (see Figure 4b) of the piston head 95 partially restricting fluid flow. With respect to the valving system alone, and not taking into account system leaks and the like, hydraulic fluid in chamber 78 is forced to escape either: (1 ) through flow around bypass needle tip 76 through channel 86 (see Figure 4b), as directionally indicated by arrow 174; or, (2) through channels in guide 80, valve 94 and the shim stack 99 (see Figure 4b.), as directionally indicated by arrows 176. Broadly, the arrow tails are indicative of origin and the arrow heads indicative of destination of fluid, at least with respect to the portion of the damping assembly under discussion. Due to the piston rod 20 withdrawing from chamber 26, a pressure-drop forms across reservoir 28 and chamber 26, resulting from removal of the volume of the piston rod 20 during withdrawal from chamber 26. Check valve 1 12 is lightly biased shut by spring 1 10. During withdrawal, the pressure drop opens check valve 122 allowing hydraulic fluid to return to damping chamber 26 from reservoir 28 via channels 1 14 and 124. The origin, flow path and destination of hydraulic fluid during rebound, with respect to the piston 22, are approximately indicated by arrows 180. Figure 4a is an enlargement of Figure 4 and shows the compression valve 30 in a closed position. The complementary faces 32a and 34a, of the valve plug 32 and valve seat 34 respectively, mate to form a valve seal 190 between reservoir 28 and plenum 26b, said closed position seal defined by the surface area of the face 32a of collar 32. Figure 4b is also an enlargement of Figure 4 and shows, during rebound, the location of rebound check valve 94 now abutting a rear portion 91 of the piston head 95 and creating a space 88 behind check valve 94. The rebound check valve channel 90 aligns with the shim stack channels 96 in the piston head 95, which are in fluid communications with shim stack 99.

In some suspension designs, the damping oil may be supplemented by a pressurized gas, typically either an inert gas such as nitrogen or air. The gas component may be retained in a separate enclosure, such as a flexible bag or behind a floating piston within a damper body. Figure 5 illustrates an alternative exemplary embodiment of the present invention with an external hydraulic fluid reservoir. The embodiment shown in Figure 5 is sometimes known as a coil over-type suspension system, with the suspension spring 272 coiled around the damper external housing 235. The suspension assembly 200 has been sectioned in order to illustrate the internal components. The damper head 213 is comprised of a top mounting bearing 201 for attachment to, for example, a vehicle chassis (not shown), a refill port 216 may be used to fill or top-up hydraulic fluid (e.g. hydraulic oil) in the damper. The refill port is sealable with a screw 215. The piston guide tube 237, which bounds the damper chamber 233 is secured within an inner portion 229 of the damper head 213.

A compression bypass needle 205 is also partially housed within the damper head 213, along with compression check valve 220. The compression bypass needle, which as mentioned is also known as the slow-speed bypass, is comprised of a tip 212 which may be moved within the aperture 203 to adjust the cross-sectional area of the aperture 203, and hence regulate the flow rate through the aperture 203. The aperture 203 is located in the damper head 213 and allows for fluid communication between channels 214 and 210 via intermediary space 204. The position of the tip 212 of the needle valve 205 within aperture 203 may be adjusted at the rear 208 of the needle valve 205. The needle valve 205 is threadingly engaged within housing 207, said housing secured 206 to the damper head 213. The needle valve may be sealed with an o-ring 21 1 so as to minimize leakage. The compression check valve 220 comprises a check valve nut 226 retaining a spring 225 which is biased to maintain check valve shims 228 in a closed position to occlude flow through channel 224 during compression. The direction of travel of the piston rod 270 and piston assembly 257 during a compression stroke is indicated by arrow 287. The compression check valve bypass passage 285 provides fluid communication between the damper chamber 233 and the needle valve 205 via chamber 214. A compression bypass valve may reduce stiffness of compression at low speed. Stiffness is associated with a harder response and typically a higher spring rate. Typically, when setting up, for example, a motorcycle, if the compression bypass valve is completely closed the ride height is higher than desirable. The ride height is preferably at about one third of wheel travel. The ride height may be too high when using the correct spring and spring preload without the compression bypass correctly set. Setting of the bypass valve is mostly used to tune to the preference of the rider. During light compression events, where suspension spring displacements are relatively small, the bypass needle valve aperture 203, depending on settings, may allow for sufficient hydraulic fluid flow for suspension damping without actuating the compression damping valve 30. When the spring 225 is compressed by a pressure drop across chamber 223 to damper chamber 233 from rebound of the piston rod 270 and piston 257 in the damper chamber 233, fluid flows through check valve channel 224 from high pressure to low pressure. The direction of travel of the piston rod 270 and piston assembly 257 during rebound is indicated by arrow 286.

An arrangement of washers 263 and o-rings 265 (not all indicated) is used as a top out spring for maximum piston extension. Optionally a traditional top-out spring or top out bump-stop may be used for the same purpose.

A side portion 209 of the damper head 213 is sealingly and slidably engaged with an inner wall 217 of damper head housing 219. An o-ring 218 provides a seal between the damper side portion 209 and the housing inner wall 217. The compression valve 30 is located within housing 219. The compression valve plug 232, or first collar, is secured to an inner surface 236 of housing 219 but if desired could be formed integrally. The damper body external casing 235 is also secured 241 to the damper head external housing. As one piece, the collar 232, damper housing 219 and damper body external casing 235 are in sliding engagement with an external wall of the piston guide tube 237 and are attached 241 to the damper body external casing 235. Seals between the guide tube 237 and the contiguous piece comprised of external housings 219 and 235, and first collar 232, are provided by o-rings for, example, 231 and 243. The compression valve seat 234 is secured to an outer portion 229 of the damper head 213, although still within the damper head housing 219. The space 230 is an annular space which allows fluid flowing from damper chamber 233 through ports 227 (only one port indicated) to apply pressure onto compression valve 30. The compression valve plug 232 remains in sealing engagement with the compression valve seat 234 unless the pressure from hydraulic fluid, generated during a compression stroke, urges the compression valve plug to move away from its mating partner 234 thereby opening a pass area between the compression valve plug 232 and seat 234 for fluid to flow through. Ports 222 provide fluid connectivity between the reservoir and the compression check valve as well as the reservoir and the compression valve. The annulus 223, ports 222, channel 275 and reservoir space 276, are at the same pressure and together comprise the available volume for fluid displaced by the piston during compression.

The damper body external casing 235 comprises an externally threaded portion to allow for spring preload adjustment through a spring preload nut 240 the setting of which may be further secured by a lock nut 238. The spring preload nut 240 acts as a seat for the external suspension spring 272. As discussed, the internal side-wall of the damper body casing 235 is slidably engaged with the outer side-wall of the damper piston chamber 233. The retainer 244 prevents the main damper body and head falling out of the assembly when the spring 272 is removed, and provides a limit the extent of movement of the external casing against the damper body, such limit prescribed by the annular space 239. The embodiment shown in Figure 5 may be inverted for some applications, for example, automobiles and off-road vehicles. As intimated, when compression valve 30 is opened by fluid pressure in damper chamber 233 generated by incursion 287 of piston shaft 270 and associated assembly 257 towards the damper head, hydraulic fluid is forced from damper chamber 233, through damper chamber ports 227, through compression valve 30 into chamber 223. The post-compression valve chamber 223 is in fluid communication via channel 275 with an ancillary fluid reservoir 274. The reservoir typically includes an internal bladder 277 the internal bladder space 279 of which may pressurized with air or inert gases such as nitrogen or argon. Different gases may have different compressibility properties. Optionally, the bladder may be replaced by a piston arrangement (not shown). When the pressure of the fluid flowing via channel 275 into space 276 exceeds the bladder pressure, then the bladder 277 will begin to compress and the volume of the bladder internal space 279 will reduce. Fluid will be urged into the reservoir space 276 around bladder 277, which is formed between the bladder and the external wall 278 of the reservoir. When the fluid pressure drops below the bladder pressure, fluid will be urged from reservoir space 276 at least in part by the expanding bladder and flow back through channel 275 into compression valve chamber 223. Aperture 282 allows for gas pressure to be increased (through gas added) or decreased (through gas removed) as required. Typically, the pressure of gas contained within bladder 277 should be sufficient pressure to return the hydraulic fluid from the reservoir space 276 into the damping system as the piston expands. The pressure employed may depend in part on tuning of the vehicle in order to optimize performance. The bladder 277 is secured 280 to a housing 283, which is further secured to the reservoir housing 278. Pressure as low as 10psi may be used with suspension damping systems of the present invention.

The relative surface areas of gaskets 218 and 243 may have an effect on the operability of the compression valve. For example, and considering gaskets 281 and 243, where the contact surface area of gasket 243 is greater than that of gasket 281 , then as you increase the pressure within the bladder the force generated by the area of 218 is greater than the force of the area generated by 243, which has the effect of opposing the force from the spring. This in turn leads to less force by the spring on the valve, which makes it softer. If this is undesirable, then in order to avoid this effect, the diameters of o-rings 218 and 243 should be equivalent.

A damping piston assembly 257 reciprocates within the damper chamber 233. The piston 257 is comprised of a piston body 259, a piston head 246, a sliding guide 255, a piston shaft 270, a rebound shim stack valve 250, a rebound check valve 251 and a rebound needle valve 268.

As discussed, a shim stack valve is comprised of flexible stacks of disc valves covering flow ports through the damping piston. The rebound shim stack 250, located at the piston head, occludes bypass passage 252 during piston compression stroke. The shim stack valve comprises the shim stack 250, the shim stack top washer 247, which may be used to increase the area of the head of the apertured shim stack screw 249, said shim stack screw 249 also holding the rebound shim stack in place 250. The aperture 248 through the shim stack screw allows fluid to flow from the damping chamber 233 through the rebound bypass passage 261. One end 262 of the bypass passage 261 cooperates with a needle valve 268 to regulate rebound bypass flow through channels 264. A rod 273 extends through the piston shaft 270, said rod enabling adjustment of the flow rate of the needle valve 268.

The piston assembly also includes a rebound check valve 251 , which is slidably engaged with the inner side-wall of the piston tube 237. The rebound check valve may include an o-ring 253 for sealing engagement and one or more channels 252 to allow for high fluid flow through the valve. In the present embodiment, the disc shaped check valve 251 is lightly biased by spring 256, said spring urging the disc 251 to abut against the underside of head of the piston 246 forming a seal. The channels 252 are configured to align with one or more shim stack bypass passages 252 in the piston body. The piston shaft 270 is slidably engaged with the bottom nut 269, which also holds the dust seal 271 a and the hydraulic shaft seal 271 b.

Figure 6 illustrates a section of an alternative exemplary embodiment 300 of the present invention with a remote suspension spring arrangement. In the embodiment illustrated in Figure 6, the compression damping valve first collar 332 is retained by spring 330 against second collar 334, by the force of the spring acting on a piston 331. The gasketed 333 piston 331 applies force to a hydraulic fluid sealed within compressed fluid space 344 formed between the piston 331 and a base 342. The compressed fluid space is connected 340 by a tube 328, which connects 324 to a second fluid space 318. Said space is formed by housing 314 and piston 310 and sealed by gaskets 312 and 316. The piston 310 is slidably engaged within the housing 314. The piston 310 is secured to the external damper housing 306, which is slidably engaged with damper head 302, and which contains the compression damping system 30. The first collar 332 is secured to an inner wall of the damper housing 306. Force applied to annular space 318 by the remote spring 330 urges the first collar 332 against second collar 334 to close valve 30. An increase in pressure in damping chamber 350 drives fluid from chamber 350, through apertures 352 into plenum space 351 . When the pressure in chamber 350, and consequently plenum 351 , is greater than the force indirectly applied by the spring 330 to annular space 318, the compression damping valve 30 opens as first collar 332 is forced away from second collar 334, with a throat formed between the collars for fluid to flow.

The compression damping valve seal contact area, that is, the area of contact between the first and second collars of the compression valve, may be varied by modifying the shape or contours of the faces of the compression valve parts. For example, it may be possible to make the profile of the valve such that the force produced by the damper will decrease for very high velocities. Turning to Figures 7a to 7d, which are cross-sectional illustrations of varying valve shapes, in Figure 7a, the standard valve first collar 32 has a valve area as indicated by the numeral 701 . This area will be the effective area that controls the force of the damper. Turning to valve 32a shown in Figure 7b, the force for low velocities will be as for the standard valve however as the velocity increases the fluid flow will be restricted by the gap 703 between the stepped surfaces 702 and 704, which effectively increases the area of the valve thus reducing the force produced by the damper. In the alternative single stepped valve 32β shown in Figure 7c, the velocity needed to start applying pressure in the secondary area is higher because the gap 705 between faces 706 and 707 is larger. Accordingly, valve 32β will start to reduce the force applied by the damper at a higher velocity than 32a. The curved shaped valve 32γ shown in Figure 7d produces a different profile of force change vs velocity. The velocity and how fast the change of force profile changes will depend on the profile of the valve and the viscosity of the oil. This feature may be useful to reduce the force from the damper in cases of hitting large bumps at very high speed without having to reduce the stiffness of the damper at normal speeds, which may increase dive resistance. Figure 8a and 8c are orthogonal views of sliding guide 80. Figure 8c is a three- dimensional illustration of sliding guide 80. Turning to Figure 8a, illustrated is a plan view of the sliding guide showing the high surface area channel 84 and the plurality of radially disposed openings 84. Also displayed are the base 802 and an aperture 801 for piston head shaft 79. Figure 8b is a three-dimensional illustration of the sliding guide 80 and shows the gasket seal 81 exterior to the sliding guide, high flow throughput apertures 82 and the high flow throughput channel indicated by 84. Also indicated is the slider base 802. Turning to Figure 8c, a side view of the sliding guide 80 is provided with gasket 81 , radial apertures 82 and base 802 indicated. Figure 9a is a three-dimensional illustration of the second collar 34 of the compression damping valve system. Indicated is the fluid channel 60a which forms between the damper head 16c and the second collar. Also indicated are the frusto-conical face 34b and the optional guide 60b. Figure 9b is an orthogonal illustration from above of the second collar of the compression damping valve system; Similarly indicated are the channel 60a, which forms between the damper ahead 16c and the second collar 34, and the optional collar guide 60b. Figure 9c is a three-dimensional illustration of the first collar 32 viewed from above. Shoulders 32b and 32c for retaining spring 14 are indicated in the figure. Figure 9d is a side illustration of the first collar 32 with first frusto-conical surface 32a indicated. Figure 9e is a three-dimensional illustration of the first collar 32 viewed from below. Frusto-conical surface 32a is indicated in the figure. Figure 10a shows a dynamometer (dyno) plot of force versus displacement for a conventional shock (plot lines A1 , B1 and C1 ) compared to a shock of the present invention (plot lines A, B and C). For the test, compression and rebound bypass valves were open and adjusted. For the conventional shock:

A1 = dyno speed 1 m/s, B1 = dyno speed 0.5m/s and C1 = dyno speed 0.1 m/s.

For the shock of the present invention:

A = dyno speed 1 m/s, B = dyno speed 0.5m/s and C = dyno speed 0.1 m/s, spring rate 0.45kg/mm.

The dynamometer measures the force but software is used to remove the spring rate and estimates the force of the damper alone. Before dynamometer testing of the shocks, compression and rebound bypass valves were adjusted by an experienced rider under enduro-riding conditions, in order to approach optimal rider comfort and performance.

For each example (A, B, C, A1 , B1 , and C1 ), the area enclosed by the loop of the curve is proportional to the energy dissipated by the damper. On conventional suspension, most energy is dissipated on the rebound which is the part of the area of the loop below the x-axis. For examples A, B and C most of the energy is dissipated on the compression stroke (area above the x-axis).

The plot indicates that for the conventional suspension (A1 , B1 , C1 ), during rebound, the force from the spring is going into the damper instead of pushing the wheel into the ground. This impacts on the grip of the wheel on the ground.

For the present invention example, the impact of the compression bypass valve can be seen from slow speed to high speed. The faster the dyno travels the greater the valve has an effect on the system indicated by the higher force. The same effect can be seen with the conventional shock although the difference is far less. The distance of -74.9 (0mm compression) is when the forks are almost completely extended (uncompressed - not stretched), and 74.8 is when the forks are compressed by about 150mm. The total compression displacement available for the forks is 300mm. The difference in force between the conventional shock and the shock of the present invention, with reference to the part left of zero on the x-axis is not large. As compression increases the difference becomes more and more marked. It is believed that this is because the damper force increases with spring compression and concomitantly, with damper piston extension into the compression chamber. As such, it is expected to be much more difficult to bottom out the damping system of the present invention compared to conventional shocks. At the same time, at the top of the stroke (corresponding to light spring compression - small bumps), the difference between the conventional shocks and the shocks of the present invention is far less marked. It would be expected, from the plot, that the shock of the present invention is less likely to dive at the top of the stroke because of high compression.

Figure 10b is shows a plot of force versus displacement for a conventional shock compared to a shock of the present invention, with compression and rebound bypass valves closed. For the conventional shock:

X1 = dyno speed 2m/s, Y1 = dyno speed 1 m/s and Z1 = dyno speed 0.1 m/s.

For the shock of the present invention:

X = dyno speed 2m/s, Y = dyno speed 1 m/s and Z = dyno speed 0.1 m/s, spring rate 0.45kg/mm.

Regarding the plots for the conventional forks from speed 0.1 m/s to 2m/s, the force more than doubles, whereas with the shock of the present invention the force remains almost the same across all speeds. For the present invention, the plot indicates that as the damper compresses, the force increases, just as you would expect from a spring. It is concluded from the plot that the dyno is indicating that the force of the damper approximately reflects the spring rate. Whereas for the conventional shock the damping effect is approximately the same irrespective of the extent of spring compression, and damping is more a function shaft speed.

Without wishing to be bound by theory, and assuming that the compression bypass is completely closed, it is estimated that the force produced by the piston is the cross-sectional area of the piston shaft times the pressure in the damping chamber, for example chamber 26 Figure 2 (equation 1 below). The pressure in the damping chamber is the force from the spring divided by the area of the valve (equation 2). Replacing the pressure term in equation 1 with equation 2 we get the force from the damper equals the area of the piston shaft times the force from the spring divided by the area of the valve (equations 3 and 4). Thus, for a given spring and given valve cross-sectional area a constant KD is identifiable (equation 5). This is the ratio of the cross-sectional area of the piston shaft 20 and the valve plug contact surface area 32a. The constant KD is an indicator of valve stiffness.

1. Fd = Ap*P

2. P = Fs/Av

3. Fd = Ap ^* Fs/Av

4. Fd=Fs*(Ap/Av)

5. KD = Ap/Av

Where:

Fd = Damper force

P = compression chamber pressure.

Fs = Force from Spring

Ap = Area of Piston (shaft)

Av = Area of Valve

With the bypass closed, the damper follows the behaviour of the spring. For example, if a rising rate spring is used the damper will also have rising rate behavior. Typically, in the present invention, the area of the ports for fluid flow during the compression stroke are in the same order of magnitude as the area of the piston displacing the fluid. This ensures that any forces due to restriction of oil flow, due to insufficient flow cross-sectional area, are minimised.

The total estimated force of the spring and compression damping valve may be expressed as follows:

6. F _T = F _s + F _D

Where: FT = Total Force

Fs = Force of the Spring

FD = Force of the Damper This can be expanded as follows:

7. F _T = d ₀ x K _s + d ₀ x K _s x K _D

Where:

Do = displacement

Ks = Spring Constant or Spring Rate

KD = Valve Ratio

Where displacement (do) is define as follows:

8. d _Q = d _t — d ₂

Where di refers to spring length before a compression event and d ₂ refers to spring length during or at the top of a compression event. Following from the Fs term above, the potential energy (Us) stored by the compressed spring may be expressed as follows:

₌ Ksxd^

2

The energy dissipated by the compression damping valve into the suspension damping system is broadly described as follows:

10. ₌ K _s xK _Dx d ₀ ² The damper force (FD) in equations 6 and 7 above is largely transferred to the working fluid and suspension system as heat.

Example 1 : On-road tests of the suspension damping system of the present invention: Observations were made in comparison to conventional shocks. With a road race motorcycle, the dive under braking or acceleration was reduced even when compared to a conventional suspension that had been set up to minimise diving and squatting at the expense of smoothness (diving happens during braking and squatting during acceleration). The benefit of less diving and squatting is that the vehicle geometry remains closer to optimum during braking or acceleration. This allows better turning behaviour and better stability. The smoothness during small high-speed bumps was considered superior to conventional suspension setup specifically for comfort at the expense of dive. Race motorcycles are typically not set this soft because they would be too unstable during braking. The result was the rider had considerably less fatigue and was able to ride more aggressively for longer. The rebound was set faster than normal allowing the tyre to achieve better grip. The transition between grip and loss of grip was found to be more predictable than on conventional suspension. Shorter settling time during braking was observed, which allowed the rider to continue braking for longer before turning. A lower tyre wear was noted (in cases over 30% lower) compared with conventional suspension. This would allow the use of softer tyre compounds resulting in more grip.

Example 2: Off-road tests of the suspension damping system of the present invention: Observations were made in comparison to conventional shocks. Lower dive and sqatting, which resulted in less geometry change during braking and acceleration. Shorter settling time after landing jumps or braking into corners. This allowed the rider to start turning much sooner than with conventional suspension. Less deflection due to bumps during corners and negligible deflection when hitting bumps at an angle. Increased grip and directional accuracy over rough terrain. Reduced rider fatigue because of reduced harshness, (reduced arm pump). Increased stability, which allowed riders to use paths that were not possible with conventional suspension. A motorcycle using the suspension of the present invention was able to perform better than conventional suspension in both motocross and enduro-tracks with minimal adjustment. Typically, a conventional suspension setup for enduro does not work on motocross tracks and vice versa. It was generally observed during testing that setup and tuning time was reduced. During testing, it was also found that valve changes when changing terrain or for different riders, was not required. During the test period, it was found to be sufficient to adjust the bypass settings. The same valve setting was used for different riders with different riding styles and only a spring change and bypass settings were changed between different riders.

It was found during testing that a valve ratio (KD) of 0.7 produced the best feel for all conditions tested (onroad and offroad) for the embodiment of the present invention tested. Valves with ratios above 0.7 led to increasing stiffness and below 0.7 to increasing softness. The dampers used in the on-road and off-road tests were exactly the same except for the length of travel and the spring used. The rear shock used on the motorbikes was also tested in a car (limited testing) without any changes to the valving. The only difference being shock travel and spring used. The observation from the car test were very similar to the on-road bike tests. Suitably, a very high valve ratio and a very light spring may be used to control the compression damping valve instead of using the main suspension spring. For example, a valve ratio of 7, an auxiliary spring of 0.1 kg/mm and a main spring of 0.9kg/mm. The shape of the compression damping valve surface may be modified so that the flow is not restricted by the smaller valve area. In some embodiments, a linkage, such as lever on one end of a leaf spring may be used to apply the force onto the compression damping valve. Suitably, the valve ratio would be smaller, for example 0.07, and a linkage ratio of 10:1 could be used to increase the force of the spring. Typically, and within limits, the compression force of the damper is not dependent on the viscosity of the hydraulic fluid. As such, the performance of the compression system should not change as the damper heats up.

If the diameter of the passage 60a (see Figure 2a) is small then flow may be restricted which leads to stiffer response. Typically, the area of the passage 60a is at least 70% of the size of the effective cross-sectional area of the piston. In respect of other fluid channels on the damping chamber side of the compression valve (for example, ports 36a and plenum 26b) the cross-sectional area may be greater than over 100% of the effective cross-sectional area of the piston.

Without wishing to be bound by theory, the inventor believes the damping valve to operate in the following manner. The pressure relief valve and coil spring rate are preferably arranged such that, during relatively small suspension excursions di and thus small initial spring compression from a static or rest state, a lower force Fi acts to oppose piston 22, thereby providing a softer ride by damping small bumps in the terrain. When the suspension encounters larger bumps or landings from drop-away terrain features, resulting in a significantly larger suspension excursion d3 and large spring compression - the spring rate (i.e. restoring or reaction force) augments the closing force on valve 30 (compared to the static spring state) thus providing a force F3 to thereby more effectively damping large bumps or landings, including to prevent full suspension excursion or bottoming. In summary, the damping force provided by the suspension fluid flowing from the damping chamber through the compression damping valve into the reservoir increases with increasing suspension excursion. The force of the compression valve is independent of the fluid flow because the suspension spring will allow the valve to open or close to maintain the force being applied by the spring. As such, the force applied by the damping valve is independent of damping piston assembly (or shaft) velocity within the damping chamber, and is dependent on the extension, length of travel or displacement of the damping piston assembly within the chamber. As the damper compresses the suspension spring shortens, which in turn applies a higher force on the valve. The inventor has observed that during compression, with the suspension damping systems of the present invention, the vehicle chassis effectively sees a stiffer spring than just the coil spring, and as such, behaviour of the compression damping valve in compression is likeable to that of a stiffer spring. In rebound, with the suspension damping systems of the present invention, the vehicle chassis predominantly sees the coil spring being controlled by the rebound system. In the suspension damping systems of the present invention, more energy may be dissipated in during compression stroke. This results in much less energy needing to be dissipated during rebound. The result is that the rebound system can be allowed to extend much faster than with conventional suspension without having oscillation.

The suspension damping systems of the present invention are anticipated to be largely velocity independent (referring to the piston damper shaft velocity in the damping chamber during a compression stroke), so the force remains comparably the same regardless of the speed the damper moves but increases as the damper and suspension spring compresses. The result is that small highspeed bumps at the top of the stroke are absorbed resulting in a smooth ride. At the same time, during the slow compression of a braking event, the damping is not reduced due to the low speed, and also increases as the suspension is loaded due to the braking. This results in a smoother ride with less diving, and, at least in part, obviates the previously mentioned compromise with slower and faster rebound systems.

In some suspension applications rebound damping control may be obviated or at least substantially reduced by the present invention. In summary, the pressure relief valve 30 is operated by dynamic feedback from displacement of the elastic support member of the suspension caused by movement of a suspended wheel relative to the chassis, rather than solely in response to the rate of travel of the damper rod and piston within the damping chamber. The suspension system of the invention can also take account of varying weights of riders, since in at least some operating modes the rider's weight adds to the loading of the elastic supporting member, being a coil spring in the embodiment. In an alternative embodiment, the elastic or resilient supporting member may comprise a leaf spring, torsion rod or pneumatic bladder, air piston or airbag.

It will be appreciated that other pressure relief valve arrangements are possible, including for example a plurality of fixed-sized orifices in an alternative collar member of the damper body 16a that could be selectively covered and uncovered by a shutter member of a cooperating collar member. However, Applicant considers that the annular valve orifice with frusto-conical, or similarly shaped engaging surfaces of the illustrated embodiments provides an excellent compromise between positive sealing at closure and precise metering of oil flow upon periodic orifice opening. Broadly, the suspension damping system may be used for any application where damping is required, for example; front and rear shocks for wheeled vehicles such as motorcycles, automobiles (including off-road vehicles), trucks; anti-roll bars, vehicle cabins; and, machinery requiring a damping system.

It is envisaged that the suspension damping system of the invention may also find application in bicycles on one hand and in the landing gear of aircraft on the other.

Optionally, the first and second collars of the compression valve 30 may be removed and replaced to modify and/or tune the properties of the compression valve in operation. The orientation of the suspension damping systems may be inverted for different applications.

In compliance with the statute, the invention has been described in language more or less specific to structural or methodical features. The term "comprises" and its variations, such as "comprising" and "comprised of" is used throughout in an inclusive sense and not to the exclusion of any additional features.

It is to be understood that the invention is not limited to specific features shown or described since the means herein described comprises preferred forms of putting the invention into effect. The invention is, therefore, claimed in any of its forms or modifications within the proper scope of the appended claims appropriately interpreted by those skilled in the art.