Inertia valve shock absorber

A damper including a valve movable between an open position and a closed position to selectively alter the compression damping rate of the shock absorber. The valve may include a self-centering feature that operates to keep the valve body centered about the valve shaft. The damper may also include a timer feature, which retains the valve in an open position for a predetermined period of time after it is initially opened.

BACKGROUND OF THE INVENTION

1. Field of the Invention

The present invention relates to vehicle suspensions systems. More particularly, the present invention relates to acceleration sensitive damping arrangements suitable for use in vehicle dampers (e.g., shock absorbers, struts, front forks).

2. Description of the Related Art

Inertia valves are utilized in vehicle shock absorbers in an attempt to sense instantaneous accelerations originating from a particular portion of the vehicle, or acting in a particular direction, and to alter the rate of damping accordingly. For example, the inertia valve may be configured to sense vertical accelerations originating at the sprung mass (e.g., the body of the vehicle) or at the unsprung mass (e.g., a wheel and associated linkage of the vehicle). Alternatively, the inertia valve may be configured to sense lateral accelerations of the vehicle.

Despite the apparent potential, and a long history of numerous attempts to utilize inertia valves in vehicle suspension, commercial inertia valve shock absorbers have enjoyed only limited success. Most attempted inertia valve shock absorbers have suffered from unresponsive or inconsistent operation due to undesired extraneous forces acting on the inertia valve. These extraneous forces may result from manufacturing limitations and/or external sources and often inhibit, or even prevent, operation of the inertia valve.

Further, there are currently no commercially available inertia valve shock absorbers for off-road bicycle, or mountain bike, applications. The problems associated with the use of inertia valves, mentioned above in relation to other vehicles, are magnified in the environment of lightweight vehicles and the relatively small size of mountain bike shock absorbers. Therefore, a need exists for an inertia valve shock absorber that can be commercially produced, and provides responsive, consistent performance without the problems associated with prior inertia valve designs.

SUMMARY OF THE INVENTION

A preferred embodiment is a shock absorber comprising a first fluid chamber, a second fluid chamber and a fluid circuit connecting the first fluid chamber and the second fluid chamber. An inertia valve includes an inertia mass movable between a first position and a second position. The inertia valve permits a first rate of fluid flow through the fluid circuit in the first position permits a second rate of fluid flow through the fluid circuit in the second position of the inertia mass. The second rate of fluid flow is non-equal to the first rate. A leading surface of the inertia mass when moving in a direction from the first position to the second position defines a leading surface area. A ratio of a mass of the inertia mass to the leading surface area is greater than about 130 grams per square inch.

A preferred embodiment is a shock absorber including a first fluid chamber, a second fluid chamber and a fluid circuit connecting the first fluid chamber and the second fluid chamber. An inertia valve includes an inertia mass movable between a first position and a second position. The inertia valve permits a first rate of fluid flow through the fluid circuit in the first position and permits a second rate of fluid flow in the second position. The second rate of fluid flow is non-equal to the first rate. A ratio of a mass of the inertia mass to a volume of the inertia mass is greater than about 148 grams per cubic inch.

A preferred embodiment is a shock absorber including a first fluid chamber, a second fluid chamber, and a fluid circuit connecting the first fluid chamber and the second fluid chamber. An inertia valve includes an inertia mass movable between a first position and a second position. The inertia valve permits a first rate of fluid flow through the fluid circuit in the first position of the inertia mass and a second rate of fluid flow in the second position of the inertia mass. The second rate of fluid flow is non-equal to the first rate. At least a portion of the inertia mass comprises tungsten.

A preferred embodiment is a shock absorber including a first fluid chamber, a second fluid chamber, and a fluid circuit connecting the first fluid chamber and the second fluid chamber. An inertia valve includes an inertia mass movable between a first position and a second position. The inertia valve permits a first rate of fluid flow through the fluid circuit in the first position of the inertia mass and a second rate of fluid flow through the fluid circuit in the second position. The second rate of fluid flow is non-equal to the first rate. The inertia mass comprises a first portion and a second portion. The first portion is constructed from a first material having a first density and the second portion being constructed from a second material having a second density, the second density being greater than the first density.

A preferred embodiment is a shock absorber including a first fluid chamber, a second fluid chamber, and a fluid circuit connecting the first fluid chamber and the second fluid chamber. An inertia valve includes an inertia mass moveable between a first position and a second position. The inertia valve permits a first rate of fluid flow through the fluid circuit in the first position of the inertia mass and a second rate of fluid flow in the second position. The second rate of fluid flow is non-equal to the first rate. The inertia mass includes a collapsible section defining at least a portion of an external surface of the inertia mass. The collapsible section has a first orientation when the inertia mass is moving in a first direction from the first position to the second position and a second orientation when the inertia mass is moving in a second direction from the second position to the first position. The inertia mass has a first flow resistance when the collapsible section is in the first orientation and a second flow resistance when the collapsible section is in the second orientation. The second flow resistance is greater than the first flow resistance.

A preferred embodiment is a shock absorber including a first fluid chamber, a second fluid chamber, and a fluid circuit connecting the first fluid chamber and the second fluid chamber. An inertia valve includes an inertia mass moveable between a first position and a second position. The inertia valve permits a first rate of fluid flow through the fluid circuit in the first position of the inertia mass and a second rate of fluid flow in the second position. The second rate of fluid flow is non-equal to the first rate. The inertia mass includes first and second opposing end surfaces oriented generally normal to a direction of motion of the inertia mass and a side wall extending between the first and second end surfaces. The inertia mass additionally includes at least one movable, annular skirt extending from the side wall. At least an outer portion of the at least one skirt moves toward the side wall when the inertia mass moves in a first direction and moves away from the side wall when the inertia mass moves in a second direction opposite the first direction. The at least one skirt increases a fluid flow drag coefficient of the inertia mass when moving in the second direction compared to the drag coefficient of movement of the inertia mass in the first direction.

A preferred embodiment is a method of delaying an inertia valve within a shock absorber from returning to a closed position after an acceleration force acting on the inertia valve diminishes. The method includes providing an inertia mass movable in a first direction from a closed position toward an open position of the inertia valve in response to an acceleration force above a predetermined threshold and movable in a second direction from the open position toward the closed position of the inertia valve when the acceleration force is below the threshold. The method further includes configuring the inertia mass to have a first fluid flow drag coefficient when moving in the first direction. The method also includes providing the inertia mass with a drag member configured to increase the fluid flow drag coefficient when the inertia mass moves in the second direction to delay the inertia valve from returning to the closed position until a period of time after the acceleration force is reduced to, and remains, below the threshold.

A preferred embodiment is a shock absorber including a first fluid chamber, a second fluid chamber, and a fluid circuit connecting the first fluid chamber and the second fluid chamber. An inertia valve includes an inertia mass and a stop. The inertia mass is movable between a first position and a second position. The inertia valve permits a first rate of fluid flow through the fluid circuit in the first position of the inertia mass and a second rate of fluid flow through the fluid circuit in the second position of the inertia mass. The second rate of fluid flow is non-equal to the first rate. One of the inertia mass and the stop defines a pocket for receiving the other of the inertia mass and the stop in the second position of the inertia mass. A first refill passage connects the second fluid chamber and the pocket and restricts fluid flow therethrough from the second fluid chamber to the pocket to provide a delay in movement of the inertia mass toward the first position. A second refill passage connects the second fluid chamber and the pocket and a pressure actuated valve substantially prevents fluid flow between the second fluid chamber and the pocket through the second refill passage below a predetermined threshold pressure differential between the second fluid chamber and the first fluid chamber. The pressure actuated valve permits fluid flow between the second fluid chamber and the pocket through the second refill passage at, or above, a predetermined threshold pressure differential between the second fluid chamber and the first fluid chamber, thereby reducing or eliminating the delay.

A preferred embodiment is a method of delaying an inertia valve within a shock absorber from returning to a closed position after an acceleration force acting on the inertia valve diminishes. The method includes providing an inertia mass movable in a first direction from a closed position toward an open position of the inertia valve in response to an acceleration force above a predetermined threshold and movable in a second direction from the open position toward the closed position of the inertia valve when the acceleration force is below the threshold. The method further includes providing a first delay force tending to resist movement of the inertia mass in the second direction when a fluid pressure differential between a first chamber and a second chamber within the shock absorber is below a predetermined threshold. The method also includes providing a second delay force, less than the first delay force, when the fluid pressure differential is at, or above, the predetermined threshold.

A preferred embodiment is a shock absorber including a first fluid chamber, a second fluid chamber and a fluid circuit connecting the first fluid chamber and the second fluid chamber. An inertia valve includes an inertia mass and a moveable stop. The inertia mass is movable between an open position and a closed position. The moveable stop is movable between a first position and a second position. The inertia mass is biased to move toward the closed position at substantially a first rate. The moveable stop and the inertia mass cooperate to define a pocket configured to receive the other of the moveable stop and the inertia mass in the open position of the inertia mass and the first position of the moveable stop. The movement of the inertia mass toward the closed position is restrained to a second rate less than the first rate. The moveable stop moves from the first position to the second position in response to a pressure within the second fluid chamber being greater than a pressure within the first fluid chamber by at least a predetermined pressure differential threshold, thereby permitting the inertia mass to return to the closed position at substantially the first rate.

A preferred embodiment is a damper including a first fluid chamber and a second fluid chamber. A fluid circuit connects the first fluid chamber and the second fluid chamber. An acceleration sensor is configured to produce a control signal in response to an acceleration force above a first predetermined threshold. The damper also has an inertia valve including an inertia mass that at least partially comprises a magnetic material and is movable between a first position and a second position. The inertia valve permits a first rate of fluid flow through the fluid circuit in the first position of the inertia mass and a second rate of fluid flow through the fluid circuit in the second position of the inertia mass. The second rate of fluid flow is non-equal to the first rate. The inertia mass moves in a direction from the first position to the second position in response to an acceleration force above a second predetermined threshold. An electromagnetic force generator is capable of retaining the inertia mass in the second position. A control system is configured to receive the control signal from the sensor and selectively activate the electromagnetic element in response to the control signal to retain the inertia mass in the second position for a predetermined period of time after the acceleration force diminishes below the first predetermined threshold.

A preferred embodiment is a bicycle including a front wheel defining a hub axis, a rear wheel, and a main frame. An acceleration sensor is mounted for movement with the hub axis of the front wheel and is configured to produce a control signal in response to sensing an acceleration above a predetermined threshold. A shock absorber is operably positioned between the rear wheel and the frame. The shock absorber includes a valve arrangement configured to receive the control signal from the sensor and to selectively alter a damping rate of the shock absorber in response to the control signal.

A preferred embodiment is a bicycle including a front wheel defining a hub axis, a rear wheel, and a main frame. An acceleration sensor is mounted for movement with the hub axis of the front wheel and is configured to produce a control signal in response to sensing an acceleration above a predetermined threshold. A shock absorber is operably positioned between the front wheel and the frame and includes a valve arrangement configured to receive the control signal from the sensor. The valve arrangement is configured to selectively alter a damping rate of the shock absorber in response to the control signal.

A preferred embodiment is a method of altering a rate of damping of a bicycle rear wheel shock absorber including sensing an acceleration force above a predetermined threshold acting on a hub axis of a front wheel of said bicycle. The method further includes providing a valve assembly within said rear wheel shock absorber configured to selectively alter a damping rate of said rear wheel shock absorber, and altering said damping rate of said rear wheel shock absorber in response to an acceleration force above said predetermined threshold.

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENT

FIG. 1illustrates an off-road bicycle, or mountain bike,20including a frame22which is comprised of a main frame portion24and a swing arm portion26. The swing arm portion26is pivotally attached to the main frame portion24. The bicycle20includes front and rear wheels28,30connected to the main frame24. A seat32is connected to the main frame24and provides support for a rider of the bicycle20.

The front wheel28is supported by a preferred embodiment of a suspension fork34which, in turn, is secured to the main frame24by a handlebar assembly36. The rear wheel30is connected to the swing arm portion26of the frame22. A preferred embodiment of a rear shock38is operably positioned between the swing arm26and the main frame24to provide resistance to the pivoting motion of the swing arm26. Thus, the illustrated bicycle20includes suspension members34,38between the front and rear wheels28,30and the frame22, which operate to substantially reduce wheel impact forces from being transmitted to the rider of the bicycle20. The rear shock absorber38desirably includes a fluid reservoir44hydraulically connected to the main shock body by a hydraulic hose46. Preferably, the reservoir44is connected to the swingarm portion26of the bicycle20above the hub axis of the rear wheel30.

The suspension fork34and the rear shock38preferably include an acceleration-sensitive valve, commonly referred to as an inertia valve, which allows the damping rate to be varied depending upon the direction of an acceleration input. The inertia valve permits the suspension fork34and rear shock38to distinguish between accelerations originating at the sprung mass, or main frame24and rider of the bicycle20, from accelerations originating at the unsprung mass, or front wheel28and rear wheel30, and alter the damping rate accordingly. It is generally preferred to have a firm damping rate when accelerations originate at the sprung mass and a softer damping rate when the accelerations originate at the unsprung mass. On an automobile or other four-wheel vehicle, this helps to stabilize the body by reducing fore and aft pitching motions during acceleration and braking, as well as by reducing body roll during cornering.

In a similar manner, on two-wheel vehicles such as motorcycles and bicycles, vehicle stability is improved by reduction of fore and aft pitching motions. In addition, in the case of bicycles and other pedal-driven vehicles, this reduces or prevents suspension movement in response to rider-induced forces, such as pedaling forces, while allowing the suspension to absorb forces induced by the terrain on which the bicycle20is being ridden. As will be described in detail below, the inertia valving within the suspension fork34and rear shock38include features which permit responsive, consistent performance and allow such inertia valves to be manufactured in a cost effective manner. Preferably, the inertia valve is located within the reservoir44, which may be rotated relative to the swingarm portion26of the bicycle20. Rotating the reservoir44alters the component of an upward acceleration of the rear wheel30which acts along the axis of motion of the inertia valve and thereby influences the responsiveness of the inertia valve.

FIGS. 2–7illustrate a preferred embodiment of the rear shock absorber38. A shock absorber38operates as both a suspension spring and as a damper. Preferably, the spring is an air spring arrangement, but coil springs and other suitable arrangements may also be used. The shock38is primarily comprised of an air sleeve40, a shock body42and a reservoir44. In the illustrated embodiment, a hydraulic hose46physically connects the main body of the shock38(air sleeve40and shock body42) to the reservoir44. However, the reservoir44may also be directly connected to the main body of the shock absorber38, such as being integrally connected to, or monolithically formed with, the air sleeve40.

The air sleeve40is cylindrical in shape and includes an open end48and an end closed by a cap50. The cap50of the air sleeve40defines an eyelet52which is used for connection to the main frame24of the bicycle20ofFIG. 1. The open end48of the air sleeve40slidingly receives the shock body42.

The shock body42is also cylindrical in shape and includes an open end54and a closed end56. The closed end56defines an eyelet58for connecting the shock38to the swing arm portion26of the bicycle20ofFIG. 1. Thus, the air sleeve40and the shock body42are configured for telescopic movement between the main frame portion24and the swing arm portion26of the bicycle20. If desired, this arrangement may be reversed and the shock body42may be connected to the main frame24while the air sleeve40is connected to the swing arm26.

A seal assembly60is positioned at the open end48of the air sleeve40to provide a substantially airtight seal between the air sleeve40and the shock body42. The seal assembly60comprises a body seal62positioned between a pair of body bearings64. The illustrated body seal62is an annular seal having a substantially square cross-section. However, other suitable types of seals may also be used. A wiper66is positioned adjacent the open end48of the air sleeve40to remove foreign material from the outer surface of the shock body42as it moves into the air sleeve40.

A damper piston68is positioned in sliding engagement with the inner surface of the shock body42. A shock shaft70connects the piston68to the cap50of the air sleeve40. Thus, the damper piston68is fixed for motion with the air sleeve40.

A piston cap72is fixed to the open end54of the shock body42and is in sliding engagement with both the shock shaft70and the inner surface of the air sleeve40. The piston cap72supports a seal assembly74comprised of a seal member76positioned between a pair of bearings78. The seal assembly74is in a sealed, sliding engagement with the inner surface of the air sleeve40. A shaft seal arrangement80is positioned to create a seal between the cap72and the shock shaft70. The shaft seal arrangement80comprises a seal member82and a bushing84. The seal member82is an annular seal with a substantially square cross-section, similar to the body seal62. The shaft seal arrangement80creates a substantially airtight seal between the cap72and the shock shaft70while allowing relative sliding motion therebetween.

A positive air chamber86is defined between the closed end50of the air sleeve40in the cap72. Air held within the positive air chamber86exerts a biasing force to resist compression motion of the shock absorber38. Compression motion of the shock absorber38occurs when the closed ends56and50of the shock body42and air sleeve40(and thus the eyelets52,58) move closer to one another.

A negative air chamber88is defined between the cap72and the seal assembly60, which in combination with the shock body42closes the open end48of the air sleeve40. Air trapped within the negative air chamber88exerts a force which resists expansion, or rebound, motion of the shock absorber38. Rebound motion of the shock absorber38occurs when the closed ends56and50of the shock body42and air sleeve40(and thus the eyelets52,58) move farther apart from each other. Together, the positive air chamber86and the negative air chamber88function as the suspension spring portion of the shock absorber38.

An air valve90communicates with the positive air chamber86to allow the air pressure therein to be adjusted. In this manner, the spring rate of the shock absorber38may be easily adjusted.

A bypass valve92is provided to allow the pressure between the positive air chamber86and the negative air chamber88to be equalized. The bypass valve92is configured to allow brief communication between the positive air chamber86and the negative air chamber88when the air sleeve seal assembly74passes thereby. A bottom out bumper94is positioned near the closed end50of the air sleeve40to prevent direct metal to metal contact between the closed end50and the cap72of the shock body42upon full compression of the shock absorber38.

The shock absorber38also includes a damper assembly, which is arranged to provide a resistive force to both compression and rebound motion of the shock absorber38. Preferably, the shock absorber38provides modal response compression damping. That is, the shock absorber38preferably operates at a first damping rate until an appropriate acceleration input is sensed, then the shock absorber38operates at a second damping rate for a predetermined period thereafter, before returning the first damping rate. This is in opposition to a system that attempts to continually respond to instantaneous input. Such a modal system avoids the inherent delay associated with responding separately to each input event.

The piston68divides the interior chamber of the shock body42into a compression chamber96and a rebound chamber98. The compression chamber96is defined between the piston68and the closed end56of the shock body42and decreases in volume during compression motion of the shock absorber38. The rebound chamber98is defined between the piston68and the piston cap72, which is fixed to the open end54of the shock body42. The rebound chamber98decreases in volume upon rebound motion of the shock absorber38.

The piston68is fixed to the shock shaft70by a hollow threaded fastener100. A seal102is fixed for movement with the piston68and creates a seal with the inner surface of the shock body42. The illustrated seal102is of an annular type having a rectangular cross-section. However, other suitable types of seals may also be used.

The piston68includes one or more axial compression passages104that are covered on the rebound chamber98side by a shim stack106. As is known, the shim stack106is made up of one or more flexible shims and deflects to allow flow through the compression passages104during compression motion of the shock absorber38but prevents flow through the compression passages104upon rebound motion of the shock absorber38. Similarly, the piston68includes one or more rebound passages108extending axially therethrough. A rebound shim stack110is made up of one or more flexible shims, and deflects to allow flow through the rebound passages108upon rebound motion of the shock absorber38while preventing flow through the rebound passages108during compression motion of the shock absorber38.

A central passage112of the shock shaft70communicates with the compression chamber96through the hollow fastener100. The passage112also communicates with the interior chamber of the reservoir44through a passage114defined by the hydraulic hose46. Thus, the flow of hydraulic fluid is selectively permitted between the compression chamber96and the reservoir44.

A rebound adjustment rod116extends from the closed end50of the air sleeve40and is positioned concentrically within the passage112of the shock shaft70. The rebound adjustment rod116is configured to alter the amount of fluid flow upon rebound motion thereby altering the damping force produced. An adjustment knob118engages the rebound adjustment rod116and is accessible externally of the shock absorber38to allow a user to adjust the rebound damping rate. A ball detent mechanism120operates in a known manner to provide distinct adjustment positions of the rebound damping rate.

The reservoir44includes a reservoir tube122closed on either end. A floating piston124is in sliding engagement with the interior surface of a reservoir tube122. A seal member126provides a substantially fluid-tight seal between the piston124and the interior surface of the reservoir tube122. The seal member126is preferably an annular seal having a substantially square cross-section. However, other suitable seals may also be used.

The floating piston124divides the interior chamber of the reservoir tube122into a reservoir chamber128and a gas chamber130. The reservoir chamber128portion of the reservoir tube is closed by an end cap132. The end cap132additionally receives the end of the hydraulic hose46and supports a hollow reservoir shaft134. The central passage136of the reservoir shaft134is in fluid communication with the passages114and112and, ultimately, the compression chamber96.

The reservoir shaft134supports an inertia valve assembly138and a blowoff valve assembly140. Each of the inertia valve assembly138and the blowoff valve assembly140allows selective communication between the compression chamber96, via the passages112,114,136, and the reservoir chamber128.

The gas chamber130end of the reservoir tube122is closed by a cap142which includes a valve assembly144for allowing gas, such as nitrogen, for example, to be added or removed from the gas chamber130. The pressurized gas within the gas chamber130causes the floating piston124to exert a pressure on the hydraulic fluid within the reservoir chamber128. This arrangement prevents air from being drawn into the hydraulic fluid and assists in refilling fluid into the compression chamber96during rebound motion of the shock absorber38.

With reference toFIG. 3b, the blowoff valve assembly140is supported by the reservoir shaft134and positioned above the inertia valve assembly138. The reservoir shaft134reduces in diameter to define a shoulder portion154. An annular washer156is supported by the shoulder154and the blowoff valve assembly140is supported by the washer156. The washer156also prevents direct contact between the inertia mass150and the blowoff valve assembly140.

The blowoff valve assembly140is primarily comprised of a cylindrical base158and the blowoff cap160. The base158is sealed to the reservoir shaft134by a shaft seal162. The illustrated seal162is an O-ring, however other suitable seals may also be used. The upper end of the base158is open and includes a counterbore which defines a shoulder164. The blowoff cap160is supported by the shoulder164and is sealed to the inner surface of the base158by a cap seal166. The cap seal166is preferably an O-ring, however other suitable seals may also be used. A threaded fastener168fixes the blowoff cap160and base158to the reservoir shaft134.

The blowoff cap160and base158define a blowoff chamber170therebetween. A plurality of radial fluid flow passages172are defined by the reservoir shaft134to allow fluid communication between the blowoff chamber170and the shaft passage136.

The blowoff cap160includes one or more axial blowoff passages174and one or more axial refill passages176. A blowoff shim stack178is positioned above the blowoff cap160and covers the blowoff passages174. The blowoff shim stack178is secured in place by the threaded fastener168. The individual shims of the shim stack178are capable of deflecting about the central axis of the fastener168to selectively open the blowoff passages174and allow fluid communication between the blowoff chamber170and the reservoir chamber128. The blowoff shim stack178is preferably configured to open in response to pressures within the blowoff chamber above a minimum threshold, such as approximately 800 psi, for example.

A refill shim stack180is positioned between the blowoff cap160and the reservoir shaft134and covers the refill ports176. The refill shim stack180is configured to prevent fluid from flowing from the blowoff chamber170through ports176to the reservoir128while offering little resistance to flow from the reservoir128into the blowoff chamber170.

The inertia valve assembly138includes a plurality of radially extending, generally cylindrical valve passages148, connecting the passage136to the reservoir chamber128. The inertia valve assembly138also includes a valve body, or inertia mass150, and a spring152. The spring152biases the inertia mass150into an upward, or closed, position wherein the inertia mass150covers the mouths of the valve passages148to substantially prevent fluid flow from the passage136to the reservoir chamber128. The inertia mass150is also movable into a downward, or open, position against the biasing force of the spring152. In the open position, the inertia mass150uncovers at least some of the valve passages148to allow fluid to flow therethrough.

The end cap132, which closes the lower end of the reservoir tube122, defines a cylindrical pocket, or socket,182which receives the inertia mass150in its lowermost or open position. The lowermost portion of the pocket182reduces in diameter to form a shoulder184. The shoulder184operates as the lowermost stop surface, which defines the open position of the inertia mass150, as illustrated inFIG. 5.

The inertia mass150includes a check plate190which allows fluid to be quickly displaced from the pocket182as the inertia mass150moves downward into the pocket182. The inertia mass150has a plurality of axial passages188extending therethrough. The check plate190rests on several projections, or standoff feet,192(FIG. 6) slightly above the upper surface of the inertia mass150and substantially covers the passages188. A series of stop projections193, similar to the standoff feet, are formed or installed in the upper, necked portion of the inertia mass150to limit upward motion of the check plate190.

With reference toFIG. 4a, a top plan view of the inertia mass150is shown. The axial passages188are preferably kidney-shaped, to allow the passages188to occupy a large portion of the transverse cross-sectional area of the inertia mass150. Desirably, the ratio of the passage188cross-sectional area to the inertia mass150cross-sectional area is greater than approximately 0.3. Preferably, the ratio of the passage188cross-sectional area to the inertia mass150cross-sectional area is greater than approximately 0.5, and more preferably greater than approximately 0.7.

The large area of the passages188provides a low-resistance flow path for hydraulic fluid exiting the pocket182. As a result, the flow rate of the fluid exiting the pocket182is high, and the inertia mass is able to move rapidly into the open position. In addition, the amount of fluid which must be displaced by the inertia mass188for it to move into the open position is reduced. Advantageously, such an arrangement allows the inertia mass150to respond rapidly to acceleration forces.

When the check plate190is resting against the standoff feet192on the upper surface of the inertia mass150it provides restricted fluid flow through the passages188. The check plate190also has an open position in which it moves upward relative to the inertia mass150until it contacts the stop projections193. When the check plate190is open, fluid is able to flow from the pocket182through the passages188and into the reservoir128, with desirably low resistance.

The inertia mass150also includes a third series of projections, or standoff feet,194. The standoff feet194are comprised of one or more projections located on the uppermost surface of the upper neck portion of the inertia mass150. The standoff feet194on the upper surface of the neck portion of the inertia mass150contact the washer156when the inertia mass150is in its uppermost or closed position. A fourth set of projections, or standoff feet,195are positioned on the lower surface of the inertia mass150(FIG. 4c) and contact the shoulder184when the inertia mass150is in its lower or open position.

In each set of stop projections, or standoff feet,192–195, preferably between three to five individual projections are disposed radially about the inertia mass150. However, other suitable numbers of feet may also be used. Desirably, the surface area of the stop projections, or standoff feet,192–195is relatively small. A small surface area of the standoff feet194,195lowers the resistance to movement of the inertia mass150by reducing the overall surface contact area between the inertia mass150and the washer156or shoulder184, respectively. The small surface area of the standoff feet192and stop projections193lower the resistance to movement of the check plate190relative to the inertia mass150. Desirably, the projections192–195have dimensions of less than approximately 0.025″×0.025″. Preferably, the projections192–195have dimensions of less than approximately 0.020″×0.020″ and, more preferably, the projections192–195have dimensions of less than approximately 0.015″×0.015″.

When utilized with an inertia mass150having a mass (weight) of approximately 0.5 ounces, the preferred projections192–195provide a desirable ratio of the mass (weight) of the inertia valve mass150to the contact surface area of the projections192–195. Due to the vacuum effect between two surfaces, a force of approximately 14.7 lbs/in2(i.e., atmospheric pressure) is created when attempting to separate the inertia mass150from either the washer156or shoulder184, respectively. By lowering the contact surface area between the inertia mass150and either the washer156or shoulder184, the vacuum force tending to resist separation of the contact surfaces is desirably reduced.

Preferably, the contact surface area is small in comparison with the mass (weight) of the inertia mass150because the magnitude of the acceleration force acting on the inertia mass150is proportional to it's mass (weight). Accordingly, a large ratio of the mass (weight) of the inertia valve mass150to the contact surface area of the projections192–195is desired. For example, for a set of three (3) standoff feet194,195with dimensions of approximately 0.025″×0.025″, the ratio is at least approximately 17 lbs/in2. A more desirable ratio is at least approximately 25 lbs/in2. Preferably, the ratio is at least 50 lbs/in2and more preferably is at least 75 lbs/in2. These ratios are desirable for an inertia mass utilized in the context of an off-road bicycle rear shock absorber and other ratios may be desirable for other applications and/or vehicles. Generally, however, higher ratios increase the sensitivity of the inertia mass150(i.e., allow the inertia mass150to be very responsive to acceleration forces). For example, with a ratio of 50 lbs/in2the sensitivity of the inertia mass150is about +/−⅓ G. Likewise, for a ratio of 147 lbs/in2the sensitivity of the inertia mass150is about +/− 1/10 G.

As illustrated inFIG. 6, the outside diameter of the lower portion of the inertia mass150is slightly smaller than the diameter of the pocket182. Therefore, an annular clearance space is defined between them when the inertia mass150is positioned within the pocket182. The clearance C restricts the rate with which fluid may pass to fill the pocket below the inertia mass150, to influence the rate at which the inertia mass150may exit the pocket182.

The interior surface of the inertia mass150includes an increased diameter central portion195which, together with the shaft134, defines an annular recess196. The annular recess196is preferably located adjacent to one or more of the ports148when the inertia mass150is in its closed position. Thus, fluid exiting from the shaft passage136through the passages148enters the annular recess196when the inertia mass150is its closed position.

The interior surface of the inertia mass150decreases in diameter both above and below the central portion195to create an upper intermediate portion197and a lower intermediate portion199. The upper intermediate portion197and lower intermediate portion199, together with the shaft134, define an upper annular clearance198(FIG. 7a) and a lower annular clearance200, respectively. An upper lip201(FIG. 7a) is positioned above, and is of smaller diameter than, the upper intermediate portion197. A step205(FIG. 7a) is defined by the transition between the upper intermediate portion197and the upper lip201. Similarly, a lower lip203is positioned below, and has a smaller diameter than, the lower intermediate portion199. A step205is defined by the transition between the lower intermediate portion199and the lower lip203. The upper lip201and the lower lip203, together with the shaft134, define an upper exit clearance202(FIG. 7a) and a lower exit clearance204.

With reference toFIG. 7a, the upper lip201preferably includes a labyrinth seal arrangement206. As is known, a labyrinth seal comprises a series of annular grooves formed into a sealing surface. Preferably, the lower lip203also includes a labyrinth seal arrangement substantially similar to the labyrinth seal206of the upper lip201.

Advantageously, the labyrinth seal arrangement206reduces fluid flow (bleed flow) between the reservoir shaft134and the upper lip201when the inertia mass150is in a closed position. Excessive bleed flow is undesired because it reduces the damping rate when the inertia valve138is closed. By utilizing a labyrinth seal206, the clearance between the inertia mass150and the shaft134may be increased, without permitting excessive bleed flow. The increased clearance is particularly beneficial to prevent foreign matter from becoming trapped between the inertia mass150and shaft134and thereby inhibiting operation of the inertia valve138. Thus, reliability of the shock absorber38is increased, while the need for routine maintenance, such as changing of the hydraulic fluid, is decreased.

With reference toFIG. 7b, an alternative inertia mass150is illustrated. The upper intermediate portion197of the inner surface of the inertia mass150ofFIG. 7bis inclined with respect to the outer surface of the shaft134, rather than being substantially parallel to the outer surface of the shaft134as in the inertia mass ofFIG. 7a. Thus, in the inertia mass150ofFIG. 7b, the step205is effectively defined by the entire upper intermediate portion197. The inertia mass150configuration ofFIG. 7btheoretically provides approximately one-half the self-centering force of the inertia mass150ofFIG. 7a. In addition, other suitable configurations of the inner surface of the inertia mass150may be utilized to provide a suitable self-centering force, as will be apparent to one of skill in the art based on the disclosure herein. For example, the inclined surface may begin in an intermediate point of the upper intermediate portion197. Alternatively, the step205may be chamfered, rather than orthogonal.

With reference toFIGS. 1–7, the operation of the shock absorber38will now be described in detail. As described previously, the shock absorber38is operably mounted between the main frame24and the swing arm portion26of the bicycle20and is capable of both compression and rebound motion. Preferably, the shock body42portion of the shock absorber38is connected to the swing arm portion26and the air sleeve40is connected to the main frame24. The reservoir44is desirably connected to the swing arm portion26of the bicycle20preferably near the rear axle, and preferably approximately vertical as shown inFIG. 1.

When the rear wheel30of the bicycle20encounters a bump the swing arm portion26articulates with respect to the main frame24, tending to compress the shock absorber38. If the acceleration imparted along the longitudinal axis of the reservoir44is below a predetermined threshold, the inertia mass150will remain in its closed position, held by the biasing force of the spring152, as illustrated inFIG. 3b.

For the piston68to move relative to the shock body42(i.e., compression motion of the shock absorber38) a volume of fluid equal to the displaced volume of the shock shaft70must be transferred into the reservoir128. With the inertia mass150closing the passages148and the blowoff valve140remaining in a closed position, fluid flow into the reservoir128is substantially impeded and the shock absorber38remains substantially rigid.

If the compressive force exerted on the rear wheel30, and thus the shock absorber38, attains a level sufficient to raise the fluid pressure within the blowoff chamber170above a predetermined threshold, such as 800 psi for example, the blowoff shims178open to allow fluid to flow from the blowoff chamber170through the blowoff ports174and into the reservoir128. As an example, if the diameter of the shock shaft70is ⅝″ (Area=0.31 square inches) and the predetermined blow-off threshold is 800 psi, then a compressive force at the shaft of at least 248 pounds is required to overcome the blowoff threshold and commence compression of the shock absorber. This required force, of course, is in addition to the forces required, as is known in the art, to overcome the basic spring force and the compression damping forces generated at the piston68of the shock absorber. In this situation, compression of the shock absorber is allowed against the spring force produced by the combination of the positive and negative air chambers86,88. The damping rate is determined by the flow through the compression ports104of the piston68against the biasing force of the compression shim stack106. When the pressure within the blowoff chamber170falls below the predetermined threshold, the blowoff shim stack178closes the blowoff ports174and the shock absorber38again becomes substantially rigid, assuming the inertia mass150remains in the closed position.

If the upward acceleration imposed along the longitudinal axis of the reservoir44(i.e., the axis of travel of the inertia mass150) exceeds the predetermined minimum threshold, the inertia mass150, which tends to remain at rest, will overcome the biasing force of the spring152as the reservoir44moves upward relative to the inertia mass150. If the upward distance of travel of the reservoir44is sufficient, the inertia mass will move into the pocket182. With the inertia mass150in the open position, fluid is able to be displaced from the compression chamber96through the passages112,114and the shaft passage136, through the passages148and into the reservoir128. Thus, the shock38is able to compress with the compression damping force again being determined by flow through the compression ports104of the piston68.

The predetermined minimum threshold for the inertia mass150to overcome the biasing force of the spring152is determined primarily by the mass of the inertia mass150, the spring rate of the spring152and the preload on the spring152. Desirably, the mass of the inertia mass is approximately 0.5 ounces. However, for other applications, such as the front suspension fork34or vehicles other than off-road bicycles, the desired mass of the inertia mass150may vary.

The spring rate of the spring152and the preload on the spring152are preferably selected such that the spring152biases the inertia mass150into a closed position when no upward acceleration is imposed along the longitudinal axis of the reservoir44. However, in response to such an acceleration force the inertia mass150will desirably overcome the biasing force of the spring152upon experiencing an acceleration which is between 0.1 and 3 times the force of gravity (G's). Preferably, the inertia mass150will overcome the biasing force of the spring152upon experiencing an acceleration which is between 0.25 and 1.5 G's and more preferably upon experiencing an acceleration which is between 0.4 and 0.7 G's. For certain riding conditions or other applications, such as the front suspension fork34, or other applications besides off-road bicycles, however, the predetermined threshold may be varied from the values recited above.

The check plate190resting on the standoff feet193of the inertia mass150allows fluid to be easily displaced upward from the pocket182and thus allows the inertia mass150to move into the pocket182with little resistance. This permits the inertia mass150to be very responsive to acceleration inputs. As the inertia mass150moves into the pocket182, fluid within the pocket182flows through the passages188and lifts the check plate190against the stop projections193.

Once the inertia mass150is in its open position within the pocket182, as illustrated inFIG. 5, the spring152exerts a biasing force on the inertia mass150tending to move it from the pocket182. Fluid pressure above the inertia mass150causes the check plate190to engage the standoff feet192located on the upper surface of the inertia mass150restricting flow through the ports188. The height of the standoff feet192which the check plate190rests on is typically 0.003″ to 0.008″ above the exit surface of the passages188to provide an adequate level of flow restriction upon upward movement of the inertia mass150. Fluid may be substantially prevented from flowing through the passages188and into the pocket182, except for a small amount of bleed flow between the checkplate190and the upper surface of the inertia mass150. However, the height of the standoff feet192may be altered to influence the flow rate of the bleed flow and thereby influence the timer feature of the inertia mass150, as will be described below.

Fluid also enters the pocket182through the annular clearance, or primary fluid flow path, C (FIG. 6) between the interior surface, or valve seat, of the pocket182and the exterior surface of the inertia mass150. Thus, the size of the clearance C also influences the rate at which fluid may enter the pocket182thereby allowing the inertia mass150to move upward out of the pocket182.

Advantageously, with such a construction, once the inertia mass150is moved into an open position within the pocket182, it remains open for a predetermined period of time in which it takes fluid to refill the pocket behind the inertia mass150through the clearance C. This is referred to as the “timer feature” of the inertia valve assembly138. Importantly, this period of time can be independent of fluid flow direction within the shock absorber38. Thus, the shock absorber38may obtain the benefits of a reduced compression damping rate throughout a series of compression and rebound cycles, referred to above as “modal response.” Desirably, the inertia mass150remains in an open position for a period between approximately 0.05 and 5 seconds, assuming no subsequent activating accelerations are encountered. Preferably, the inertia mass150remains in an open position for a period between about 0.1 and 2.5 seconds and more preferably for a period between about 0.2 and 1.5 seconds, again, assuming no subsequent accelerations are encountered which would tend to open the inertia mass150, thus lengthening or resetting the timer period. The above values are desirable for a rear shock absorber38for an off-road bicycle20. The recited values may vary in other applications, however, such as when adapted for use in the front suspension fork34or for use in other vehicles or non-vehicular applications.

In order to fully appreciate the advantages of the modal response inertia valve assembly138of the present shock absorber38, it is necessary to understand the operation of a bicycle having an acceleration-sensitive damping system utilizing an inertia valve. With reference toFIG. 8, the relationship between vertical position P, vertical velocity V and vertical acceleration A, over time T, for a simple mass traversing two sinusoidally-shaped bumps is illustrated.FIG. 8is based on a mass that travels horizontally at a constant velocity, while tracking vertically with the terrain contour. This physical model, somewhat simplified for clarity, correctly represents the essential arrangement utilized in inertia-valve shock absorbers wherein the inertial element is directly connected to, and driven by, the unsprung mass.

The primary simplification inherent in this model, and in this analysis, is that the flexibility of an actual bicycle tire is ignored. The tire is assumed to be inflexible in its interaction with the terrain, offering no compliance. An actual tire, of course, will provide some compliance, which in turn produces some degree of influence on the position, velocity, and acceleration of the unsprung mass. The actual degree of influence in a given situation will depend on many variables, including the actual vehicle speed and the specific bump geometry, as well as the compliance parameters of the particular tire. However, the simplified analysis discussed here is a good first approximation which clearly illustrates the key operative physics principles, while avoiding these complications. The basic validity of this simplified analysis can be demonstrated by a sophisticated computer motion analysis that incorporates the effects of tire compliance and several other complicating factors.

RelatingFIG. 8to the situation of a bicycle, the heavy solid line indicating position P represents both the trail surface and, assuming the wheel of the bicycle is rigid and remains in contact with the trail surface, the motion of any point on the unsprung portion of the bicycle, such as the hub axis of the front or rear wheel, for example. The lines representing velocity V and acceleration A thus correspond to the vertical velocity and acceleration of the hub axis. InFIG. 8, the trail surface (solid line indicating position P) includes a first bump B1and a second bump B2. In this example, as shown, each bump is preceded by a short section of smooth (flat) terrain.

As the wheel begins to traverse the first bump B1, the acceleration A of the hub axis H rises sharply to a maximum value and, accordingly, the velocity V of the hub axis H increases. Mathematically, of course, the acceleration as shown is calculated as the second derivative of the sinusoidal bump curve, and the velocity as the first derivative. At a point P1, approximately halfway up the first bump B1, the second derivative (acceleration A) becomes negative (changes direction) and the velocity begins to decrease from a maximum value. At a point P2, corresponding with the peak of the bump B1, the acceleration A is at a minimum value (i.e., large negative value) and the velocity V is at zero. At a point P3, corresponding with the mid-point of the downside of the first bump B1, the acceleration A has again changed direction and the velocity V is at a minimum value (i.e., large negative value). At a point P4, corresponding with the end of the first bump B1, the acceleration A has risen again to a momentary maximum value and the velocity V is zero. The second bump B2is assumed to be sinusoidally-shaped like the first bump B1, but, as shown, to have somewhat greater amplitude. Thus, the relationship between position P, velocity V and acceleration A are substantially identical to those of the first bump B1.

When a simple inertia valve is utilized in the suspension system of a bicycle and the acceleration A reaches a threshold value, the inertia mass overcomes the biasing force of the spring and remains in a substantially stationary position as the shaft which supports the inertia mass moves upward. Once the shaft has moved upward relative to the inertia mass a sufficient distance, the inertia valve passages are uncovered and a reduced compression damping rate is achieved. Although a compression inertia valve is discussed in this example, the same principles may be applied to an inertia valve which operates during rebound.

Before the inertia valve passages are open, the shock absorber operates at its initial, firm damping rate. This results in an undesirably firm damping rate, creating a “damping spike”, over the initial portion of the bump B1. The damping spike continues until the shaft has moved upward relative to the inertia mass a sufficient distance to open the valve passages. The amount of movement of the shaft relative to the inertia mass necessary to uncover the passages is determined primarily by the size of the passages and the position of the uppermost surface of the inertia mass relative to the passages when the mass is in its fully closed position. This distance is referred to as the spike distance SD. The amount of time necessary for the inertia passages to be opened and to reduce the damping rate is dependent upon the shape of the bump and the spike distance SD. and is referred to as the spike time ST. The reduction of the damping rate is at least partially dependent upon the size of the passages and, therefore, it is difficult to reduce the spike time STwithout reducing the spike distance SDwhich necessarily affects the achievable lowered damping rate.

The inertia mass begins to close (i.e., move relatively upward) when the acceleration acting upon it either ceases, changes direction, or becomes too small to overcome the biasing force of the spring. As shown graphically inFIG. 8, the acceleration A becomes zero at point P1, or at approximately the mid-point of the bump B1. Accordingly, a simple inertia valve begins to close at, or before, the middle of the bump B1. Therefore, utilizing a simple inertia valve tends to return the shock absorber to its initial, undesirably firm damping rate after only about one-half of the bump B1has been traversed. The operating sequence of the inertia valve is similar for the second bump B2and each bump thereafter.

In actual practice, the specific point on a bump where a simple inertia valve will close will vary depending on bump configuration, vehicle speed, inertia valve size and geometry, spring bias force, compliance of the tire and other factors. Thus, it should be understood that the extent of mid-bump “spiking” produced by “premature closing” of a simple inertia valve will be greater for some bumps and situations than for others.

It is desirable to extend the amount of time the inertia valve stays open so that the reduced damping rate can be utilized beyond the first half of the bump. More complex inertia valve arrangements utilize the fluid flow during compression or rebound motion to hydraulically support the inertia valve in an open position once acceleration has ceased or diminished below the level necessary for the inertia valve to remain open from acceleration forces alone. However, these types of inertia valve arrangements are dependent upon fluid flow and allow the inertia valve to close when, or slightly before, the compression or rebound motion ceases. A shock absorber using this type of inertia valve in the compression circuit could experience a reduced damping rate from after the initial spike until compression motion ceases at, or near, the peak P2of the bump B1. This would represent an improvement over the simple inertia valve shock absorber described previously. However, the flow dependent inertia valve necessarily reacts to specific terrain conditions. That is, the inertia mass responds to each individual surface condition and generally must be reactivated upon encountering each bump that the bicycle traverses. Therefore, this type of shock absorber experiences an undesirably high damping rate “spike” as each new bump is encountered.

In contrast, the inertia valve arrangement138of the present shock absorber38is a modal response type. That is, the inertia valve138differentiates rough terrain conditions from smooth terrain conditions and alters the damping rate accordingly. During smooth terrain conditions, the inertia valve138remains in a closed position and the damping rate is desirably firm, thereby inhibiting suspension motion due to the movement of the rider of the bicycle20. When the first bump B1is encountered, the inertia valve138opens to advantageously lower the damping rate so that the bump may be absorbed by the shock absorber38. The timer feature retains the inertia valve138in an open position for a predetermined period of time thereby allowing the shock absorber38to maintain the lowered damping rate for the entire bump (not just the first half), and to furthermore absorb the second bump B2and subsequent bumps possibly without incurring any additional “spikes.” As discussed above, the timer period may be adjustable by altering the rate at which fluid may refill the timer pocket182.

Once the shock absorber38has been compressed, either by fluid flow through the blowoff valve140or the inertia valve138, the spring force generated by the combination of the positive air chamber86and the negative air chamber88tend to bias the shock body42away from the air sleeve40. In order for the shock absorber38to rebound, a volume of fluid equal to the displaced volume of the shock shaft70must be drawn from the reservoir128and into the compression chamber96. Fluid flow is allowed in this direction through the refill ports176in the blowoff valve140against a desirably light resistance offered by the refill shim stack180. Gas pressure within the gas chamber130exerting a force on the floating piston124may assist in this refill flow. Thus, the rebound damping rate is determined primarily by fluid flow through the rebound passages108against the biasing force of the rebound shim stack110.

With reference toFIGS. 3band5, the fluid flow path during compression or rebound motion of the shock absorber38, with the inertia mass150in either of an open or closed position, is above and away from the inertia mass150itself. Advantageously, such an arrangement substantially isolates fluid flow from coming into contact with the inertia mass150, thereby inhibiting undesired movement of the inertia mass due to drag forces resulting from fluid flow. Thus, the inertia mass150advantageously responds to acceleration inputs and is substantially unaffected by the movement of hydraulic fluid during compression or rebound of the shock absorber38.

The present shock absorber38includes an inertia valve138comprising a self-centering valve body, or inertia mass150. In order to fully appreciate the advantages of the self-centering inertia mass150of the present inertia valve assembly138, it is necessary to describe the conditions which have prevented prior inertia valve designs from operating reliably, with acceptable sensitivity, and for a reasonable cost.

Each ofFIGS. 9 and 10schematically illustrate an off-center condition of the inertia mass150relative to the shaft134. The off-center condition of the inertia mass150may cause it to contact the shaft134causing friction, which tends to impede motion of the inertia mass150on the shaft134. Due to the relatively small mass of the inertia mass150and the desirability of having the inertia mass150respond to small accelerations, any friction between the inertia mass150and the shaft134seriously impairs the performance of the inertia valve138and may render it entirely inoperable. Each of the off-center conditions illustrated inFIGS. 8 and 9may result from typical manufacturing processes. However, modifying the manufacturing process to avoid these conditions often results in a prohibitively high manufacturing cost.

FIG. 9illustrates an inertia valve arrangement in which the inertia valve passages148are of slightly different diameter. Such a condition is often an unavoidable result of the typical manufacturing process of drilling in a radial direction through a tubular piece of material. Such a process may result in an entry diameter N created by the drilling tool being slightly larger than the exit diameter X created by the drilling tool. The resulting difference in area between the passages148causes the fluid pressure within the shaft passage136to exert an unequal force between the entry passage148having an entry diameter N and the exit passage148having an exit diameter X.

For example, a difference between the entry diameter N and the exit diameter X of only two thousandths of an inch (0.090″ exit diameter versus 0.092″ entry diameter) at a fluid pressure of 800 psi, results in a force differential of approximately 0.2 pounds, or 3.6 ounces, between the passages148. The inertia mass150itself may weigh only about one half of an ounce (0.5 oz.). Such a force differential will push the inertia mass150off-center and reduce the responsiveness of the inertia mass150, if not prevent it from moving entirely.

FIG. 10illustrates an off-center condition of the inertia mass150caused by the inertia valve passages148being positioned off-center relative to the shaft134. A center axis AC of the inertia valve passages148is offset from the desired diametrical axis AD of the shaft134by a distance O. Therefore, the force resulting from fluid pressure within the shaft passage136does not act precisely on a diametrical axis AD of the inertia mass150, resulting in the inertia mass150being pushed off-center with respect to, and likely contacting, the shaft134. The offset condition of the center axis AC of the passages148is the result of inherent manufacturing imperfections and cannot easily be entirely avoided, at least without raising the cost of manufacturing to an unfeasible level.

Furthermore, even if manufacturing costs were not of concern and the passages148could be made with identical diameters and be positioned exactly along the diametrical axis AD of the shaft134, additional forces may tend to push the inertia mass150off-center. For example, if the reservoir44experiences an acceleration which is not exactly aligned with the axis of travel of the inertia mass150(such as braking or forward acceleration), the transverse component of the acceleration would create a force tending to move the inertia mass150off-center and against the shaft134. If the transverse component of the acceleration is large enough, the resulting frictional force between the inertia mass150and the reservoir shaft134will inhibit, or prevent, movement of the inertia mass150. Accordingly, it is highly desirable to compensate for factors which tend to push the inertia mass150off-center in order to ensure responsive action of the inertia valve138. This is especially important in off-road bicycle applications, where it is desirable for the inertia valve assembly138to respond to relatively small accelerations and the mass of the inertia mass150is also relatively small.

As described above, the inertia valve assembly138preferably includes a self-centering inertia mass150. With reference toFIG. 11, the inertia mass150ofFIG. 5is shown without the fluid flow lines to more clearly depict the cross-sectional shape of its interior surface. The inertia mass150has a minimum internal diameter “D” while the shaft134has a constant external diameter “d,” which is smaller than the internal diameter D. The difference between the shaft diameter d and the inertia valve diameter D is desirably small. Otherwise, as described above, the bleed flow between the shaft134and the inertia mass150undesirably reduces the damping rate which may be achieved when the inertia mass150is in a closed position. Accordingly, for the rear shock38the difference between the shaft diameter d and the inertia mass diameter D is desirably less than 0.01 inches. Preferably, difference between the shaft diameter d and the inertia mass diameter D is less than 0.004 inches and more preferably is approximately 0.002 inches. For the front suspension fork34, the difference between the shaft diameter d and the inertia mass diameter D is desirably less than 0.02 inches. Preferably, difference between the shaft diameter d and the inertia mass diameter D is less than 0.008 inches and more preferably is approximately 0.004 inches. The recited values may vary in other applications, however, such as when adapted for vehicles other than off-road bicycles or non-vehicular applications.

The preferred differences between the shaft diameter d and the inertia mass diameter D recited above assume that a labyrinth seal arrangement206(FIG. 7) is provided at the upper and lower portions of the internal surface of the inertia mass150, as described above. However, the bleed rate may be influenced by factors other than the difference between the shaft diameter d and the inertia mass diameter D. Accordingly, driven by a pressure differential of 400 psi, the bleed rate between the inertia mass150and the shaft134, for an off-road bicycle shock with a shaft diameter of ⅝ inches, is desirably less than 1.0 cubic inches/sec. Preferably, the bleed rate between the inertia mass150and the shaft134is less than 0.5 cubic inches/sec and more preferably is less than 0.3 cubic inches/sec. However, for applications other than off-road bicycle shock absorbers, the preferred bleed rates may vary.

As described, an annular recess196is defined between the interior surface of the inertia mass150and the shaft134. The annular recess196is preferably located in approximately the center of the inertia mass150. The annular recess196is referred to as zone1(Z1) in the following description of the fluid flow between the shaft134and the self-centering inertia mass150. The upper annular clearance198, above the annular recess196, is referred to as zone2(Z2) and the upper exit clearance202is referred to as zone3(Z3). One half of the difference between the diameter of the upper annular clearance198and the diameter D at the upper exit clearance202defines a distance B, which is equivalent to the size of the step205. The size B of the step205(referred to as a “Bernoulli Step” inFIGS. 26,27and28) may be precisely manufactured by a computer controlled lathe operation, for example. Other suitable methods for creating a precisely sized step205may also be used.

Zone1Z1has a larger cross-sectional fluid flow area than zone2Z2which, in turn, has a larger cross-sectional flow area than zone3Z3. The cross-sectional area differential between the zones Z1, Z2, Z3causes the fluid within each zone Z1, Z2, Z3to vary in velocity, which causes a self-centering force to be exerted on the inertia mass150when it becomes off-center, as will be described below. Although the zones Z1, Z2, Z3are annular, the discussion below is in the context of a two-dimensional structure having left and right sides. Accordingly, the zones Z1, Z2, Z3of the example will vary in cross-sectional distance, rather than in cross-sectional area. Although the example is simplified, it correctly describes the general self-centering action of the inertia mass150.

A rough approximation of the centering force developed by the self-centering inertia mass150can be estimated using Bernoulli's equation. This is a rough approximation only since Bernoulli's equation assumes perfect frictionless flow, which is not valid for real fluids. However, this is a useful starting point for understanding the general principles involved, and for estimating the forces that occur. Bernoulli's equation expresses the law of conservation of energy for the flow of an incompressible fluid. In estimating the centering force of the inertia mass150, the potential energy (height) portion of Bernoulli's equation is not significant and may be ignored. Thus, for any two arbitrary points on a fluid streamline, Bernoulli's equation reduces to:
P1+(ρ/g)(V1)2=P2+(ρ/g)(V2)2
where:P1—fluid pressure (psi) at point1P2=fluid pressure (psi) at point2V1=fluid velocity (in/sec) at point1V2=fluid velocity (in/sec) at point2p=fluid densityg=gravity constant

Using the values of 0.3125 lb/in3for fluid density ρ of typical hydraulic fluid and 386 in/sec2for gravity constant g, the equation becomes:
P1+(4.05×10−5)(V1)2=P2+(4.05×10−5)(V2)2

For a simple example, assume that the fluid pressure P1in zone1is 400 psi, due to an external force tending to compress the shock absorber38and the fluid velocity V1is zero due to relatively little fluid exiting from zone1. Also, for simplicity, assume that the floating piston124is absent or is not exerting a significant pressure on the fluid within the reservoir chamber128. Accordingly, the fluid pressure P3in zone3Z3is 0 psi. Insert these values into Bernoulli's equation to find the velocity in zone3:
400+(4.05×10−5)(0)2=0+(4.05×10−5)(V3)2
V3=3,142 in/sec

Therefore, as a first approximation (accurate to the degree that the assumptions Bernoulli's equation are based upon are valid here) the velocity V3of fluid exiting zone3is 3,142 in/sec. Assuming the validity of assumptions inherent in Bernoulli's equation here, this value is true for all exit points of zone3Z3regardless of their dimensions. Further, based on flow continuity, the change in velocity of the fluid between zone2Z2and zone3Z3is proportional to the change in the clearance, or gap G, between zone2Z2and zone3Z3. The gap G is the cross-sectional distance between the outer surface of the shaft134and the relevant inner surface of the inertia mass150.

The relationship between the change in the size of the gap G and the change in velocity allows solving of the velocity in zone2Z2for both the right and left sides. Assuming that D is 0.379 inches, d is 0.375 inches and B is 0.001 inches, then the gaps on both the right and left sides, with the inertia mass150centered are:
GAP Zone 2=B+(D−d)/2=0.003
GAP Zone 3=(D−d)/2=0.002

Then, based on flow continuity, fluid velocity in Zone2is calculated as follows:

Therefore, the fluid velocity V2in zone2Z2for each of the right and left side is 2,094 in/sec. Using Bernoulli's equation to find the pressure P2in zone two gives:
400+(4.05×10−5)(0)2=(P2)+(4.05×10−5)(2,094)2
P2=222 psi

Assuming that, for a particular inertia valve, the area in zone2Z2that the fluid pressure acts upon for each of the right and left side is 0.0375 in2, then the force F at both the left and right sides of the inertia mass150can be calculated as:
F=222 psi (0.0375 in2)=8.3 lbs.

The force F acting on the inertia mass150in the above example is equal for the right and left side due to the velocity V2in zone2Z2being the same for each side. The velocity V2is the same because the ratio of gap3G3to gap2G2between the right side and the left side is equal due to the inertia mass150being centered relative to the shaft134.

With reference toFIG. 12, however, if the inertia mass150becomes off center relative to the shaft134by a distance x, for example 0.001 inches to the left, the ratio of gap3G3to gap2G2is different between the right and left sides. This results in the velocity V2being different between the right and left sides and, as a result, a force differential between the right side and left side is produced. These calculations are substantially similar to the previous calculations and are provided below (for an off-center condition 0.001 inches to the left:
V3=3,142 in/sec
Left Side: GAP Zone 3 (G3L)=(D−d)/2+x=0.003
GAP Zone 2 (G2L)=B+(D−d)/2+x=0.004

As shown, a force differential of as much as 4.7 lbs, depending on the degree of validity of the Bernoulli assumption, pushes the inertia mass150to the right to correct for the off-center condition. As noted above, preferably the lower portion of the inertia mass150also includes a step205creating a lower zone2and zone3(FIG. 12). Accordingly, a centering force acts on the lower portion of the inertia mass150when it is off-center from the shaft134. Therefore, in the example above, a force of as much as 4.7 lbs also acts on the lower portion of the inertia mass150, resulting in a total centering force of as much as 9.4 lbs acting to center the inertia mass150relative to the shaft134.

For a typical off-road bicycle application, with the inertial mass centered, the ratio of the velocity in zone2V2to the velocity in zone3V3(i.e., V2/V3) is desirably between 0.9 and 0.2. Preferably, the ratio of the velocity in zone2V2to the velocity in zone3V3is desirably between 0.8 and 0.35 and more preferably the ratio of the velocity in zone2V2to the velocity in zone3V3is desirably between 0.75 and 0.5.

The ratio of the gap G between the shaft134and the inertia mass150in zone3Z3and in zone2Z2(i.e., G3/G2), as demonstrated by the calculations above, influences the magnitude of the self-centering force produced by the inertia mass150. The ratio (G3/G2) is desirably less than one. If the ratio (G3/G2) is equal to one, then by definition there is no step205between zone2Z2and zone3Z3.

Based on flow continuity from Zone2to Zone3, the ratio of the velocity V2in Zone2to the velocity V3in Zone3(V2/V3) is equal to the ratio of the Gap G3at Zone to the Gap G2at Zone2(G3/G2). In other words, based on flow continuity it follows that:
(G3/G2)=(V2/V3).

Thus, for a typical off-road bicycle application with the inertia mass centered, the ratio of the gap at Zone3to the gap at Zone2is desirable between 0.90 and 0.20. Preferably the ratio of the gap at Zone3to the gap at Zone2is desirably between 0.80 and 0.35 and more preferably the ratio of the gap at Zone3to the gap at Zone2is desirably between 0.75 and 0.50.

Advantageously, the self-centering inertia mass150is able to compensate for force differentials due to the manufacturing variations in the passage148size and position as well as transverse accelerations, all of which tend to push the inertia mass150off-center. This allows reliable, sensitive operation of the inertia valve assembly140while also permitting cost-effective manufacturing methods to be employed without compromising performance.

Although a fluid pressure in zone1Z1of 400 psi was used in the above example, the actual pressure may vary depending on the force exerted on the shock assembly38. The upper pressure limit in zone1Z1is typically determined by the predetermined blow off pressure of the blow off valve140. Desirably, for an off-road bicycle rear shock with a shaft diameter of ⅝ inches, the predetermined blow off pressure is approximately 400 psi. Preferably, the predetermined blow off pressure within zone1Z1is approximately 600 psi and more preferably is approximately 800 psi. These predetermined blow off pressures are provided in the context of an off-road bicycle rear shock application and may vary for other applications or vehicle types.

FIG. 13illustrates an alternative arrangement for controlling the refill rate, or timer function, of fluid flow into the pocket182as the inertia mass150moves in an upward direction away from its closed position. The end cap132includes a channel208communicating with an orifice209connecting the reservoir chamber128and the pocket182. The orifice209permits fluid to flow between the reservoir chamber128and the pocket182in addition to the fluid flow through the clearance C and bleed flow between the check plate190and inertia mass150. The size of the orifice209may be varied to influence the overall rate of fluid flow into the pocket182.

FIG. 13also illustrates an adjustable pocket refill arrangement210. The adjustable refill arrangement210allows external adjustment of the refill rate of fluid flow into the pocket182. The adjustable refill arrangement includes an inlet channel212connecting the reservoir chamber128to a valve seat chamber213. An outlet channel214connects the valve seat chamber213to the pocket182.

A needle215is positioned within the valve seat chamber213and includes a tapered end portion216, which extends into the outlet channel214to restrict the flow of fluid therethrough. External threads of the needle215engage internal threads of the end cap132to allow the needle215to move relative to the outlet channel216. The needle215includes a seal217, preferably an O-ring, which creates a fluid tight seal between the needle215and the end cap132. The exposed end of the needle215includes a hex-shaped cavity218for receiving a hex key to allow the needle215to be rotated. The exposed end of the needle215may alternatively include other suitable arrangements that permit the needle215to be rotated by a suitable tool, or by hand. For example, an adjustment knob may be connected to the needle215to allow a user to easily rotate the needle without the use of tools.

Rotation of the needle215results in corresponding translation of the needle215with respect to the end cap132(due to the threaded connection therebetween) and adjusts the position of the tapered end216relative to the outlet channel214. If the needle215is moved inward, the tapered end216blocks a larger portion of the outlet channel214and slows the fluid flow rate into the pocket182. If the needle215is moved outward, the tapered end216reduces its blockage of the outlet channel214and speeds the fluid flow rate into the pocket182. This permits user adjustment of the refill rate of the pocket182and, accordingly, adjustment of the period of time the inertia mass150is held in an open position. Advantageously, the adjustable refill arrangement210allows a user to alter the period of time the inertia valve138is open and thus, the period of lowered compression damping once the inertia valve138is opened.

FIG. 14illustrates the suspension fork34detached from the bicycle20ofFIG. 1. The suspension fork34includes right and left legs220,222, as referenced by a person in a riding position on the bicycle20. The right leg220includes a right upper tube224telescoping received in a right lower tube226. Similarly, the left leg222includes a left upper tube228telescopingly received in a left lower tube230. A crown232connects the right upper tube224to the left upper tube228thereby connecting the right leg220to the left leg222of the suspension fork34. In addition, the crown232supports a steerer tube234, which passes through, and is rotatably supported by the frame22of the bicycle20. The steerer tube234provides a means for connection of the handlebar assembly36to the suspension fork34, as illustrated inFIG. 1.

Each of the right lower tube226and the left lower tube230includes a dropout236for connecting the front wheel28to the fork34. An arch238connects the right lower tube226and the left lower tube230to provide strength and minimize twisting of the tubes226,230. Preferably, the right lower tube226, left lower tube230, and the arch238are formed as a unitary piece, however, the tubes226,230and the arch238may be separate pieces and connected by a suitable fastening method.

The suspension fork34also includes a pair of rim brake bosses240to which a standard rim brake assembly may be mounted. In addition, the fork34may include a pair of disc brake bosses (not shown) to which a disc brake may be mounted. Of course, the suspension fork34may include only one or the other of the rim brake bosses240and disc brake bosses, depending on the type of brake systems desired.

FIG. 15is a cross-section view of the right leg220of the suspension fork34having the front portion cutaway to illustrate the internal components of a damping assembly244of the fork34. Preferably, the left leg222of the suspension fork34houses any of a known suitable suspension spring assembly. For example, an air spring or coil spring arrangement may be used. In addition, a portion of the suspension spring assembly may be housed within the right fork leg220along with the damper assembly244.

As described previously, the upper tube224is capable of telescopic motion relative to the lower tube226. The fork leg220includes an upper bushing246and a lower bushing248positioned between the upper tube224and the lower tube226. The bushings246,248inhibit wear of the upper tube224and the lower tube226by preventing direct contact between the tubes224,226. Preferably, the bushings246,248are affixed to the lower tube226and are made from a self-lubricating and wear-resistant material, as is known in the art. However, the bushings246,248may be similarly affixed to the upper tube224. Preferably, the bushings246,248include grooves (not shown) that allow a small amount of hydraulic fluid to pass between the bushings246,248and the upper fork tube224to permit lubrication of the bushing246and seal, described below.

The lower tube226has a closed lower end and an open upper end. The upper tube224is received into the lower tube226through its open upper end. A seal250is provided at the location where the upper224enters the open end of the lower tube226and is preferably supported by the lower tube226and in sealing engagement with the upper tube224to substantially prevent oil from exiting, or a foreign material from entering the fork leg220.

The damping assembly244is operable to provide a damping force in both compression and a rebound direction to slow both compression and rebound motion of the fork34. The damper assembly244is preferably an open bath, cartridge-type damper assembly having a cartridge tube252fixed with respect to the closed end of the lower tube226and extending vertically upward. A damper shaft254extends vertically downward from a closed upper end of the upper tube224and supports a piston258. Thus, the piston258is fixed for movement with the upper tube224while the cartridge tube252is fixed for movement with the lower tube226.

The piston258is positioned within the cartridge tube252and is in telescoping engagement with the inner surface of the cartridge tube252. A cartridge tube cap260closes the upper end of the cartridge tube252and is sealing engagement with the damper shaft254. Thus, the cartridge tube252defines a substantially sealed internal chamber which contains the piston258.

The piston258divides the internal chamber of the cartridge tube252into a variable volume rebound chamber262and a variable volume compression chamber264. The rebound chamber262is positioned above the piston258and the compression chamber264is positioned below the piston258. A reservoir266is defined between the outer surface of the cartridge tube252and the inner surfaces of the upper and lower tubes224,226. A base valve assembly268is operably positioned between the compression chamber264and the reservoir266and allows selective communication therebetween.

FIG. 16is an enlarged cross section of the damping assembly244. As described above, a cartridge tube cap260closes the upper end of the cartridge tube252. An outer seal270creates a seal between the cartridge tube cap260and the cartridge tube252while an inner seal272creates a seal between the cartridge tube cap260and the damper shaft254. Accordingly, extension and retraction of the damper shaft254with respect to the cartridge tube252is permitted while maintaining the rebound chamber262in a substantially sealed condition.

The cartridge cap260includes a one-way refill valve274which, during inward motion of the damper shaft254with respect to the cartridge tube252, allows fluid flow from the reservoir266into the rebound chamber262. The refill valve274comprises one or more axial passages276through the cap260which are closed at their lower end by refill shim stack278. Thus, the shim stack278allows fluid flow from the reservoir266to the rebound chamber262with a relatively small amount of resistance. When the fluid pressure in the rebound chamber262is greater than the fluid pressure in the reservoir266, such as during retraction of the damper shaft254, the refill shim stack278engages the lower surface of the cartridge tube cap260to substantially seal the refill passages276and prevent fluid from flowing therethrough.

The piston258is fixed to the end of the damper shaft254by a threaded fastener280. The piston includes an outer seal282which engages the inner surface of the cartridge tube252to provide a sealing engagement between the piston258and the inner surface of the cartridge tube252. Thus, fluid flow around the piston is substantially eliminated.

The piston258includes a one-way rebound valve assembly284which permits fluid flow from the rebound chamber262to the compression chamber264while preventing flow from the compression chamber264to the rebound chamber262. The rebound valve assembly284comprises one or more axial passages286through the piston258closed at their lower end by a rebound shim stack288. Fluid is able to flow from the rebound chamber262through the passages286and into the compression chamber264against the resistance offered by the shim stack288. When the pressure is greater in the compression chamber264than in the rebound chamber262, the shim stack288engages the lower surface of the piston258to substantially seal the passages286and prevent the flow of fluid therethrough.

In the illustrated embodiment, the cartridge tube252is split into an upper portion290and a lower portion292, which are each threadably engaged with a connector294to form the cartridge tube252. Optionally, a one-piece cartridge tube may be employed. A base member296is fixed to the closed end of the lower tube226and supports the cartridge252. The lower portion292of the cartridge tube252is threadably engaged with the base member296.

FIG. 17is an enlarged cross-sectional view of the base valve assembly268. The base valve assembly268is housed within the lower portion292of the cartridge tube252and is supported by a shaft298which extends in an upward direction from the base member296. The entire base valve assembly268is secured onto the shaft298by a bolt300which threadably engages the upper end of the shaft298.

The base valve assembly268includes a compression valve302, a blowoff valve304, and an inertia valve306. The compression valve302is positioned on the upper portion of the shaft298. The blowoff valve304is positioned below the compression valve302and spaced therefrom. The compression valve302and the blowoff valve304define a blowoff chamber308therebetween. A plurality of passages310connect the blowoff chamber308to a central passage312of the base valve shaft298.

A snap ring314, which is held in an annular recess of the shaft298, supports the compression valve302. A washer316positioned underneath the bolt300holds the compression valve302onto the shaft298. The compression valve302includes a compression piston318sealingly engaged with the inner surface of the lower portion292of the cartridge tube252by a seal320. The compression piston318is spaced from both the snap ring314and the washer316by a pair of spacers322,324respectively.

The compression piston318includes one or more compression passages326covered by a compression shim stack328. The compression shim stack328is secured to the lower surface of the compression piston318by the lower spacer322. The compression shim stack328deflects about the lower spacer322to selectively open the compression passages326. The compression shim stack328seals against the lower surface of the compression piston318to prevent unrestricted compression flow past the compression shim stack328.

As illustrated inFIGS. 20 and 21, which show fluid flows during the rebound stroke, the compression piston318also includes one or more refill passages330extending axially through the compression piston318. The refill passages330are covered at the upper surface of the compression piston318by a refill shim stack332. The refill shim stack332is held against the upper surface of the compression piston318by the upper spacer324and deflects to open the refill passages330. Thus, the refill shims332prevent fluid flow through the refill passages from the compression chamber264to the blowoff chamber308, but permit fluid flow from the blowoff chamber308through the refill passages330and into the compression chamber264against the slight resistance offered by the refill shim stack332.

As illustrated inFIG. 17, the blowoff valve304is positioned between a lower snap ring334and an upper snap ring336. A separator plate338is supported by the lower snap ring334and is sealingly engaged with the inner surface of the lower portion292of the cartridge tube252by a seal340. A lower spacer342spaces the blowoff piston344in an upward direction from the separator plate338. The blowoff piston344is also sealingly engaged with the inner surface of the lower portion292of the cartridge tube252by a seal346. An upper spacer348spaces the blowoff piston344from the upper snap ring336. A separator chamber350is defined between the blowoff piston344and the separator plate338.

As illustrated inFIGS. 20 and 21, the blowoff piston344includes one or more blowoff passages352covered on the lower surface of the blowoff piston344by a blowoff shim stack354. The blowoff shim stack354is positioned between the blowoff piston344and the lower spacer342to allow fluid flow from the blowoff chamber308into the separator chamber350at pressures above a predetermined threshold. The blowoff shim stack354seals passages352to prevent unrestricted (without blowoff) compression fluid flow from the blowoff chamber308to the separator chamber350.

The blowoff piston344also includes one or more refill passages356covered at the upper surface of the blowoff piston344by a refill shim stack358. The refill shim stack358is held against the upper surface of the blowoff piston344by the upper spacer348to seal the refill passages356and prevent fluid flow from the blowoff chamber308into the separator chamber350. However, the refill shims deflect about the upper spacer348to allow fluid flow from the separator chamber350into the blowoff chamber308through the refill passages356with relatively little resistance. One or more passages360are formed within the lower portion292of the cartridge tube252at a height between the separator plate338and the blowoff piston344to allow fluid communication between the separator chamber350and the reservoir266.

Preferably, the inertia valve306is substantially identical to the inertia valve previously described in relation to the shock absorber38. The inertia valve306includes an inertia mass362movable between a closed position, where the inertia mass362closes two or more passages364, and an open position, where the inertia mass362uncovers the two or more passages364. The uppermost or closed position of the inertia mass362is defined by the snap ring334, which supports the separator plate338.

The inertia mass362is biased into its closed position by a spring366. The lowermost or open position of the inertia mass362is defined when the lower surface of the inertia mass362engages the lower interior surface of a pocket368, defined by the base member296. The inertia mass362includes one or more axial passages370covered at the upper surface of the inertia mass362by a check plate372which is movable between a substantially closed position against the standoff feet394at the upper surface of the inertia mass362and an open position against the stop projections392on the upper, necked portion of the inertia mass362.

The check plate372moves into an open position when the inertia mass362moves downward in relation to the base valve shaft298to allow fluid to flow from the pocket368into an inertia valve chamber376above the inertia mass362through the passages370. The check plate372moves into a substantially closed position upon upward movement of the inertia mass362relative to the base valve shaft298to restrict fluid flow through the passages370. One or more passages378are defined by the lower portion292of the cartridge tube252to allow fluid communication between the inertia valve chamber376and the reservoir266.

An annular clearance C is defined between the inertia mass362and the pocket368when the inertia mass362is in its open position. In a similar manner to the inertia valve described in relation to the shock absorber38, the clearance C restricts fluid flow from the inertia valve chamber376into the pocket368. The inertia valve306preferably includes other features described in relation to the inertia valve of the shock absorber38. For example, the inertia mass362preferably includes a plurality of standoff feet394at the locations discussed above in relation to the inertia mass of the shock absorber38. Additionally, the inertia mass362includes an annular recess380aligned with the passages364when the inertia mass362is in its closed position. The inertia mass362also includes a step preferably on each end of the interior surface of the inertia mass362which is sliding engagement with the base valve shaft298, as described above. As shown, the inertia mass362also includes a labyrinth seal arrangement substantially as described above.

When the front wheel28of the bicycle20ofFIG. 1encounters a bump, a force is exerted on the fork34, which tends to compress the fork legs224,226in relation to each other. If the upward acceleration of the lower fork tube226along its longitudinal axis (i.e., the axis of travel of the inertia mass362) is below a predetermined threshold, the inertia mass362remains in its closed position. Pressure within the compression chamber264causes fluid to flow through the compression passages326and into the blowoff chamber308. If the pressure within the blowoff chamber308is below a predetermined threshold, the blowoff shims354remain closed and the suspension fork34remains substantially rigid.

If the pressure within the blowoff chamber308exceeds the predetermined threshold, the blowoff shim stack354deflects away from the blowoff piston344to allow fluid to flow through the blowoff passage352into the separator chamber350and into the reservoir through the passages360, as illustrated inFIG. 17. Thus, the fork34is able to compress with the compression damping rate being determined primarily by the shim stack354of the blowoff piston344.

As the upper fork leg224moves downward with respect to the lower fork leg226, and thus the piston258and damper shaft254move downward with respect to the cartridge252, fluid is drawn into the rebound chamber262through the refill valve274, as illustrated inFIG. 16.

When the upward acceleration of the lower fork leg226exceeds a predetermined threshold, the inertia mass362tends to stay at rest and overcomes the biasing force of the spring366to open the passages364. Thus, fluid flow is permitted from the central passage312of the base valve shaft298into the inertia chamber376through the passages364and from the inertia chamber376into the reservoir266through the passages378, as illustrated inFIGS. 18 and 19. Accordingly, at pressures lower than the predetermined blowoff pressure, when the inertia mass362is open (down) fluid is permitted to flow from the compression chamber264to the reservoir266and the suspension fork244is able to compress.

Upon rebound motion of the suspension fork34, the refill valve274closes and the fluid within the rebound chamber262is forced through the rebound passages286of the piston258against the resistive force of the rebound shim stack288, as illustrated inFIG. 20. A volume of fluid equal to the displaced volume of the damper shaft254is drawn into the compression chamber264from the reservoir chamber266via the passages356and330against the slight resistance offered by the refill shims358and332, as illustrated inFIG. 21.

FIGS. 22–25illustrate an alternative embodiment of the suspension fork34. The embodiment ofFIGS. 22–25operates in a substantially similar manner as the suspension fork34described in relation toFIGS. 14–21with the exception that the embodiment ofFIGS. 22–25allows flow through a compression valve382in the piston258during compression motion. This is known as a shaft-displacement type damper, because a volume of fluid equal to the displaced volume of the shaft254is displaced to the reservoir266during compression motion of the fork34. For reference, this compares with the previously-described embodiment where the displaced fluid volume equals the displaced volume of the full diameter of the piston258. Flow through the piston258into the rebound chamber during compression eliminates the need for refill passages in the cartridge cap, and thus a solid cap260is utilized.

The compression valve382is a one-way valve, similar in construction to the one-way valves described above. The compression valve382comprises one or more valve passages384formed axially in the piston258and a shim stack386closing the valve passages384. As is known, the shim stack386may comprise one or more shims. The shims may be combined to provide a desired spring rate of the shim stack386. The shim stack386is deflected to allow fluid flow between the compression chamber264and the rebound chamber262during compression of the suspension fork34. Preferably, shim stack386is significantly “softer” than shim stack328in the base valve assembly268, in order to ensure sufficient pressure for upward flow through piston258into rebound chamber262during compression strokes.

The operation of the suspension fork34ofFIGS. 22–25is substantially similar to the operation of the suspension fork34described in relation toFIGS. 14–21. However, during compression motion of the fork34ofFIGS. 22–25, fluid flows from the compression chamber264to the rebound chamber262. This results in less fluid being displaced into the reservoir266than in the previous embodiment. As will be appreciated by one of skill in the art,FIGS. 22 and 23illustrate compression fluid flow when the blow off valve304is open.FIGS. 24 and 25illustrate compression fluid flow when the inertia valve306is open.

As will be appreciated by one of ordinary skill, the illustrated suspension fork and rear shock absorber arrangements advantageously minimize unintended movement of the inertia mass150due to normal compression and rebound fluid flow. With particular reference toFIG. 3b, compression fluid flow (illustrated by the arrow inFIG. 3b) through the blow off valve140of the rear shock absorber38occurs through the passage136of the reservoir shaft134as it passes the inertia mass150. Accordingly, fluid moving with any substantial velocity does not directly contact the inertia mass150, thereby avoiding undesired movement of the inertia mass150due to forces from such a flow. Similarly, compression fluid flow through the passages148when the inertia mass150is in an open position (FIG. 5) and refill fluid flow upon rebound of the shock absorber38are similarly insulated from the inertia mass150. With reference toFIGS. 17,19and21, the inertia mass150is also insulated from contact with moving fluid in the suspension fork34.FIGS. 23 and 25illustrate similar flow paths for the second embodiment of the suspension fork34.

FIG. 26is a graph illustrating the influence of a change in the internal diameter D of a specific inertia mass150on the pressure differential between the right and left side when the inertia mass150is off-center by a distance x of 0.001 inches. As described above in relation toFIGS. 11 and 12, the reservoir shaft134, which defines an axis of motion for the inertia mass150, has a diameter referred to by the reference character “d.” The reference character “B” refers to the size of the step205, or the difference in the radial dimensions of the inner surface of the inertia mass150between zone2Z2and zone3Z3. For the purposes of illustration in the graph ofFIG. 26, the diameter d of the shaft134is given a value of 0.375 inches. The step size B is given a value of 0.001 inches.

In the graph ofFIG. 26, the value of the minimum internal diameter of the inertia mass150(i.e., the diameter at zone3Z3) is varied and the corresponding pressure differential between the left and right sides is illustrated by the line388, given the constants d, B and x. As described above, the self-centering force is proportional to the pressure differential produced by the design of zones1,2and3of the self-centering inertia mass150. Thus, as the pressure differential increases, so does the ability of the inertia mass150to center itself with respect to the shaft134. As illustrated, the value of the pressure differential between the left and right sides varies greatly with relatively small changes in the internal diameter D of the inertia mass150. The pressure differential is at its maximum value on the graph when the difference between the inertia valve diameter D and the shaft diameter d is small. The pressure differential diminishes as the difference between the inertia valve diameter D and the shaft diameter d increases.

For example, when the inertia valve diameter D is equal to 0.400 inches, the pressure differential is equal to approximately 8 psi. With the inertia valve diameter D equal to 0.400 inches and the shaft diameter d equal to 0.375 inches, the total gap at zone3G3for both the left and right sides is equal to 0.025 inches (0.400–0.375), when the inertia mass150is centered. Accordingly, each gap at zone3for the left and right side, G3Land G3R, is equal to 0.0125 inches (0.025/2), when the inertia mass150is centered (FIG. 11).

The pressure differential has substantially increased at a point when the inertia valve diameter D is equal to 0.385. At this point, the resulting pressure differential is approximately 38 psi. Following the calculation above, each gap at zone3for the left and right side, G3Land G3R, is equal to 0.005 inches, with a centered inertia mass150.

The pressure differential has again substantially increased, to approximately 78 psi, at a point when the inertia valve diameter D is equal to 0.381 inches. When the inertia diameter D is equal to 0.381 inches, each gap at zone3for the left and right side, G3Land G3R, is equal to 0.003 inches, assuming the inertia mass150is centered about the shaft134. At a point when the inertia valve diameter D is equal to 0.379, the pressure differential has increased significantly to approximately 125 psi. At this point, the gap at zone3for the left and right side, G3Land G3R, is 0.002 inches.

The illustrated pressure differential reaches a maximum when the inertia valve diameter D is equal to 0.377 inches. At this value of D, the pressure differential is approximately 180 psi and each gap at zone3for the left and right side, G3Land G3R, is equal to 0.001 inches, again assuming a centered inertia mass150and the values of d, B and x as given above. Although the gap at zone3G3may be reduced further, resulting in theoretically greater self-centering forces, a gap in zone3G3of at least 0.001 inches is preferred to allow the inertia mass150to move freely on the shaft134. A gap G3below this value may allow particulate matter within the damping fluid to become trapped between the inertia mass150and shaft134, thereby inhibiting or preventing movement of the inertia mass150.

FIG. 27is a graph illustrating the relationship between the size B of the “Bernoulli step”205and the resulting pressure differential percentage. A pressure differential of 0% indicates no pressure differential, and thus no self-centering force, is present (i.e., the pressure on the right and left sides of the inertia mass150are equal), while a pressure differential of 100% indicates a maximum pressure differential, and self-centering force, is present (i.e., zero pressure on one side of the inertia mass150). The graph is based on a gap at zone3G3of 0.002 inches, with the inertia mass150centered. In other words, the inertia mass diameter D minus the shaft diameter d is equal to 0.004 inches, which results in a gap on each of the right and left sides, G3Rand G3L, of 0.002 inches.

The graph includes individual lines390,392,394and396representing different off-center values of the inertia valve. The values are given in terms of the percentage of the total gap G3(0.002″ inFIG. 27) that the inertia mass150is off-center. For example, an off-center amount of 25% means that the center axis of the inertia mass150is offset 0.0005 inches to either the left or right from the center axis of the shaft134. Similarly, an off-center amount of 50% means that the center axis of the inertia mass150is offset 0.001 inches from the center axis of the shaft134. Line390represents an off-center amount of 25%, line392represents an off-center amount of 50%, line394represents an off-center amount of 75%, and line396represents an off-center amount of 99%.

The largest step size B illustrated on the graph ofFIG. 27is 0.008 inches. A step205of a larger size B may be provided, however, as indicated by the graphs, theoretical self-centering effects have diminished significantly at this point. Accordingly, the step size is desirably less than 0.008 inches, at least for off-road bicycle applications based on these theoretical calculations. The ratio between the gap at zone3G3and the gap at zone2G2(i.e., G3/G2) in this situation is ⅕, for a centered inertia mass150and a gap at zone3G3of 0.002 inches.

With continued reference toFIG. 27, lines390–396illustrate that the pressure differential has increased at a point when the step size B is equal to 0.006 inches in comparison to the pressure differential at a step size B of 0.008 inches. At this point, the ratio between the gap at zone3G3and the gap at zone2G2(i.e., G3/G2), for a centered inertia mass150, is ¼. As a result, the self-centering effect is more substantial for ratios which are greater than ¼. The pressure differential again increases at a point when the step size B is equal to 0.004 inches. At this point, the ratio between the gap at zone3G3and the gap at zone2G2(i.e., G3/G2), for a centered inertia mass150, is ⅓. As a result, the self-centering force for ratios above self-centering force ⅓ is increased over the self-centering force obtained with a larger step size B.

For at least a portion of the lines390–396, the pressure differential again increases for step sizes B less than 0.003. At this point, the ratio of the gap at zone3G3to the gap at zone2G2(i.e., G3/G2), for a centered inertia mass150, is ⅖. Accordingly, the self-centering effect is more substantial for ratios which are greater than ⅖. Furthermore, at least a portion of the lines390–396illustrate an increase in the pressure differential at a point when the step size B is equal to 0.002 inches. At this point, the ratio of the gap at zone3G3to the gap at zone2G2(i.e., G3/G2), for a centered inertia mass150, is ½. As a result, the self-centering effect is more substantial for ratios which are greater than ½.

The graph ofFIG. 27illustrates a general trend that, up to a point, the pressure differential percentage (and self-centering force) increases as the step size B is reduced, especially for large off-center amounts. However, practical considerations also prevent the size B of the step205from becoming too small. For example, extremely small step sizes may be difficult to manufacture, or in the very least, difficult to manufacture for a reasonable cost. Accordingly, the size B of the step205(i.e., G2–G3) is desirably greater than, or equal to, 0.0001 inches. Preferably, the size B of the step205is greater than or equal to 0.001 inches. Additionally, for the practical concerns described above, the effectiveness of the self-centering inertia mass150, at least theoretically, declines as the step sizes B become too large. Accordingly, the size B of the step205is preferably less than 0.002 inches. However, as mentioned above, the graph ofFIG. 27is based on theoretical calculations using Bernoulli's equation, which assumes perfect fluid flow. For actual fluid flows, a much larger step size B may be desirable. For example, in actual applications, a step size B of 0.02 inches, 0.03 inches, or even up to 0.05 inches is believed to provide a beneficial self-centering effect. The effectiveness of larger step sizes B in actual applications is primarily a result of boundary layers of slow-moving, or non-moving fluid adjacent the inertia mass150and shaft134surfaces resulting in a lower actual flow rate than theoretically calculated using Bernoulli's equation.

FIG. 28is a graph, similar to the graph ofFIG. 27, illustrating the relationship between the size B of the step205and the resulting pressure differential percentage, except that the gap G3is 0.001 inches when the inertia valve150is centered. That is, the inertia mass diameter D minus the shaft diameter d is equal to 0.002 inches, which results in a gap on each of the right and left sides, G3Rand G3L, of 0.001 inches.

The graph includes individual lines representing inertia mass150off-center values of 25%, 50%, 75% and 99%. Line400represents an off-center amount of 25%, line402represents an off-center amount of 50%, line404represents an off-center amount of 75%, and line406represents an off-center amount of 99%.

The largest step size B illustrated on the graph ofFIG. 28is 0.008 inches. The ratio between the gap at zone3G3and the gap at zone2G2(i.e., G3/G2) in this situation is 1/9, for a centered inertia mass150and a gap at zone3G3of 0.001 inches. A step size B of greater than 0.008 inches is possible however, as discussed above, at least for off-road bicycle applications, the step size B is preferably less than 0.008 inches based on theoretical calculations.

For at least a portion of the illustrated off-center amounts, the pressure differential increases at a point when the step size B is equal to 0.003 inches. At this point, the ratio between the gap at zone3G3and the gap at zone2G2(i.e., G3/G2), for a centered inertia mass150, is ¼. As a result, the centering effect is more substantial for ratios which are greater than ¼. The lines400–406illustrate that the pressure differential again increases at a point when the step size B is equal to 0.002 inches. At this point, the ratio between the gap at zone3G3and the gap at zone2G2(i.e., G3/G2), for a centered inertia mass150, is ⅓. As a result, the self-centering effect is greater for ratios above ⅓.

The pressure differential again increases for step sizes B less than 0.0015. At this point, the ratio of the gap at zone3G3to the gap at zone2G2(i.e., G3/G2), for a centered inertia mass150, is ⅖. Accordingly, the centering effect is more substantial for ratios which are greater than ⅖. Further, the pressure differential increases at a point when the step size B is equal to 0.001 inches. At this point, the ratio of the gap at zone3G3to the gap at zone2G2(i.e., G3/G2), for a centered inertia mass150, is ½. As a result, the centering effect is more substantial for ratios which are greater than ½.

The design parameters of the self-centering inertia mass150described above, including the size of the gaps G in the different zones (Z1, Z2, Z3) and the size B of the step205, for example, as well as other considerations, such as the length of time the inertia mass150stays open in response to an activating acceleration force, the spring rate of the biasing spring and the mass of the inertia mass150, for example, may each be varied to achieve a large number of possible combinations. More than one combination may produce suitable overall performance for a given application. In a common off-road bicycle application, the combination desirably provides a self-centering force of between 0 and 800 lbs. for an off-center amount of 25%. Preferably, a self-centering force of between 0 and 40 lbs. is produced and more preferably, a self-centering force of between 0 and 5 lbs. is produced for an off-center value of 25%. Desirably, the combination provides a self-centering force of at least 0.25 ounces for an off-center amount of 25%. Preferably, a self-centering force of at least 0.5 ounces is produced and more preferably, a self-centering force of at least 1 ounce is produced for an off-center value of 25%. Most preferably a self-centering force of at least 2 ounces is produced for an off-center value of 25%. The above values are desirable for a rear shock absorber38for an off-road bicycle20. The recited values may vary in other applications, such as when adapted for use in the front suspension fork34or for use in other vehicles or non-vehicular applications.

FIG. 29illustrates an inertia valve assembly410, which is similar to the inertia valve assembly138ofFIG. 3B. The inertia valve assembly410ofFIG. 29may be incorporated in a shock absorber, such as the shock absorber38of the bicycle20illustrated inFIG. 1. The inertia valve assembly410desirably includes an inertia mass412, which has an increased density in comparison to the inertia mass150ofFIG. 3B. As a result, the inertia mass412is more responsive to an acceleration force of a given magnitude. Preferably, the inertia valve assembly410operates in a substantially similar manner to the inertia valve arrangement138described above and, therefore, the inertia valve assembly410and associated shock absorber are described in limited detail.

Preferably, the inertia valve assembly410is disposed within a reservoir tube414and is operable to selectively permit fluid flow between a first fluid chamber416and a second fluid chamber418. In a preferred embodiment, the first fluid chamber416comprises a compression chamber of the shock absorber and the second fluid chamber418comprises a reservoir chamber of the shock absorber. Preferably, the inertia mass412is supported for axial movement on an axis Ac, which is defined by a shaft420. The inertia mass412is biased in an upward direction (with respect to the orientation of the tube414illustrated inFIG. 29) against an upper stop, defined by snap ring422, by a biasing member, such as coil spring424. In this position, the inertia mass412closes openings434in the shaft420to define a closed position of the inertia valve assembly410.

A base426is coupled to a lower end of the reservoir tube414and, preferably, includes a cavity428, which defines a pocket430below the inertia mass412. The pocket430is sized and shaped to receive at least a lower portion of the inertia mass412. A bottom surface of the cavity432functions as a lower stop for the inertia mass412. As described in detail above, preferably, the inertia mass412is responsive to an appropriate acceleration force input above a predetermined threshold. Upon being subjected to such an acceleration force, the inertia mass412moves downwardly relative to the shaft420, against the biasing force of the spring424, and into the pocket430. In this position, the inertia mass412uncovers openings434to permit fluid flow from the first fluid chamber416to the second fluid chamber418and define an open position of the inertia valve assembly410.

The inertia valve assembly410also includes a refill valve assembly436, which preferably is configured to at least partially control a flow of fluid between the reservoir chamber418and the pocket430. In the illustrated embodiment, the valve assembly436includes a plurality of hooks438(only one shown) extending in an upward direction from the base426. Preferably, the hooks438are disposed around the periphery of the cavity428adjacent an inner surface of the reservoir tube414. In a preferred arrangement, four such hooks438are equally spaced around a periphery of the cavity428.

The hooks438define an upper stop surface440and an upper surface of the base426defines a corresponding lower stop surface442. A check plate444is retained for movement between the upper stop surface438and the lower stop surface442. Preferably, the check plate444is substantially annular in shape with an inner diameter which is slightly larger than an outer diameter of an adjacent portion of the inertia mass412, such that a clearance distance C is defined therebetween.

In a preferred arrangement, the check plate444is configured to restrict a flow of fluid from the reservoir chamber418into the pocket430at a first level and permit fluid flow from the pocket430to the reservoir418at a second level, which preferably is greater than the first level. In operation, when the inertia mass412is moving downward relative to the shaft420, such as due to an appropriate acceleration force, the movement of fluid out of the pocket430lifts the check plate444in an upward direction against the upper stop surface440, as illustrated in phantom. Accordingly, a large amount of fluid is permitted to be displaced from the pocket430to the reservoir chamber418, as illustrated by the phantom flow line445.

Conversely, when the inertia mass412is moving from a lower most position, within the pocket430, toward the upper stop422, fluid within the reservoir418attempts to fill the pocket430thereby urging the check plate444against the lower stop surface442, as illustrated by the solid line position of the check plate444. In the lower position of the check plate444, fluid is restricted to entering the pocket430by passing through the clearance distance C between an inner surface of the check plate444and an outer surface of the inertia mass412, as illustrated by the solid flow line446. Preferably, with such an arrangement, the flow into the pocket430is restricted to a rate that is lower than the rate in which fluid may exit the pocket430. Accordingly, the inertia mass412may move quickly in a downward direction into the pocket430, while movement in an upward direction is slowed to delay the closing of the inertia valve410in order to extend the reduced-damping mode of the shock absorber, as described in detail above.

Desirably, the inertia mass412is configured to have a relatively high density, and thus a high mass for a given volume, so that the inertia mass412moves more easily through the damping fluid within the chambers418and430to increase the responsiveness of the inertia valve410to acceleration force inputs. Preferably, the inertia mass412includes a first section, comprising a first material, and a second section, comprising a second material having a greater density than the first material. Desirably, the second material has a density greater than about 10 g/cm3and, preferably, greater than about 15 g/cm3. More preferably, the second material has a density of about 19 g/cm3. In the illustrated arrangement, the inertia mass412comprises a body portion450, which defines an annular cavity452filled with a high density material454, so as to increase the overall mass of the inertia mass412without increasing the volume that it occupies. A presently preferred high density material454is tungsten, preferably in a powdered form.

In addition, the ratio of the mass of the inertia mass412to the surface area of a lowermost surface456of the inertia mass412, normal to the axis Ac, is also increased in comparison to the previously described inertia mass constructions. The surface456may be defined as a leading surface of the inertia mass412when the inertia mass412is moving in a downward direction (i.e., toward the open position). Accordingly, the leading surface area includes a surface456aof standoff feet455, which is generally parallel with the surface456and perpendicular to the axis Acof the shaft420. Due to the increased mass to volume, and mass to leading surface area ratios, the inertia mass412more easily displaces fluid from the pocket430to move more quickly toward the open position in response to suitable acceleration force inputs.

In a preferred arrangement, a threaded cap458closes an open, upper end of the cavity452to retain the tungsten454within the cavity. A peripheral edge of the cap458includes external threads460, which mate with internal threads462of the cavity452. Thus, the cavity452may be filled with tungsten454, or another high density material, and closed with the threaded cap458.

The embodiment illustrated inFIG. 29is preferred at least because the main body portion450of the inertia mass412may be made from a relatively dense, yet readily processable material, such as brass for example, while permitting a material with even higher density, such as tungsten powder, to be held within the cavity452without the need for it to be formed or otherwise processed. Alternatively, the entire inertia mass412may be made from a material having higher density than brass, such as solid tungsten for example. In a preferred embodiment, the cavity452, and thus the tungsten powder454or other high density material, occupies a significant portion of the total volume of the inertia mass412. For example, desirably the high density material occupies at least one-third volume of the inertia mass412. Preferably, the high density material occupies at least one-half and, more preferably, at least two-thirds of the volume of the inertia mass412. However, other ratios between the material comprising the main body450and the material within the cavity452may also be used.

An inertia mass configured substantially as described above provides advantages mass to surface area, or mass to volume, ratios so that the inertia mass is very responsive to acceleration force inputs. The tables below illustrate the change in mass to surface area and mass to volume ratios for a constant volume inertia mass and a constant mass inertia mass, respectively, having varying relative volumes of brass and tungsten. In generating the tables, the annular inertia mass was assumed to have a length of 0.875 inches, an inner diameter of 0.375 inches and, for the constant volume inertia mass, an outer diameter of one (1) inch. For the constant mass inertia mass, the outer diameter (and, thus, the leading surface area) varies. The density of brass was assumed to be 8.5539 g/cm3and the density of tungsten was assumed to be 19.3 g/cm3. The constant volume inertia mass was assumed to have a volume of 9.685 cm3and the constant mass inertia mass was assumed to have a mass of 83 grams. The ratios are provided in grams/cubic inch for mass to volume and grams/square inch for mass to surface area.

FIGS. 30,31A and31B illustrate an alternative inertia mass470, which preferably is configured to provide increased flow resistance, or drag, when moving in a first direction compared to the flow resistance when moving in a second, or opposite direction. In a preferred arrangement, the inertia mass470includes one or more collapsible drag members472, which are configured to assume a first orientation when the inertia mass470is moving in a first direction and a second orientation when the inertia mass470is moving in the opposite direction.

As in the inertia valve assemblies described above, the inertia mass470is supported for axial movement on a shaft474within a reservoir chamber476. In the illustrated embodiment, the inertia mass470includes a body portion478, the outer surface of which defines a pair of annular grooves480. The annular grooves support the drag members472, which are also annular in shape. In a preferred arrangement, the drag members472are constructed from a flexible material, such as rubber or plastic, and extend upwardly and outwardly from the outer surface of the body portion478of the inertia mass. In addition, the drag members472may curve in an upward direction from an inner diameter to an outer diameter of the drag member472. Accordingly, a peripheral edge portion of each drag member472tends to be collapsible in an upward direction relative to the inner edge portion of the drag member472.

In operation, when the inertia mass470is moving in a downward direction relative to the shaft474, or toward an open position, fluid flow illustrated by the arrows482inFIG. 31A exerts an upward force on the drag members tending to collapse the drag members radially inward. Accordingly, a leading surface area of the inertia mass470is reduced and the fluid482flows past the drag members472with, preferably, little interruption. Thus, preferably, the drag members472exert little resistive force against the downward movement of the inertia mass470toward the open position.

Conversely, when the inertia mass470is moving in an upward direction relative to the shaft474, toward the closed position, fluid484flowing beside the inertia mass470tends to open the drag members472into their relaxed, or radially extended, orientation, as illustrated inFIG. 31B. Thus, preferably, the drag members472cause turbulent flow of the fluid adjacent the body portion478. Such flow significantly increases the resistance to fluid484flowing past the inertia mass470and, thereby, slows the movement of the inertia mass470toward the closed position. Thus, the drag members472provide a delay, or timer function, to the inertia mass470, in a manner similar to the timer arrangements described above.

The drag members472may be used in addition, or in the alternative, to other delay producing devices, such as the valve436ofFIG. 29or the clearance passage C illustrated inFIG. 6. Furthermore, although two drag members472are provided in the illustrated inertia valve assembly470, a greater or lesser number of drag members472may also be used. In addition, although the drag members472are illustrated as annular members extending outwardly from a side wall of the inertia mass470, other constructions are also possible. For example, collapsible drag members may be disposed above or below the main body478of the inertia mass470and be configured in a similar manner to achieve the same, or similar, effect.

FIG. 32illustrates an alternative inertia valve assembly490in which the delay in closing of the inertia mass492is influenced by a pressure differential between the pressure of the fluid within the reservoir chamber494and the pressure of the fluid within the passage526. During a rebound stroke of the shock absorber, as fluid exits the reservoir chamber494, flowing downward (relative to the orientation shown inFIG. 32) through the central shaft496, a pressure drop occurs. For a given flow rate, the magnitude of the pressure drop is influenced by the diameter of the flow passage in the shaft496. A smaller flow passage diameter creates a larger pressure drop top to bottom.

Similar to the previous embodiments, the inertia mass492is supported by a shaft496for axial movement about an axis Ac. The inertia mass492is positioned within the reservoir chamber494defined by a reservoir tube498. A base500is connected to a lower end of the reservoir tube498and defines a recess502which, in turn, defines a pocket504for receiving at least a lower portion of the inertia mass492when the inertia mass492is in the open position. Thus, a bottom surface of the recess502functions as a lower stop for the inertia mass492. The inertia mass492is biased against an upper stop, defined by snap ring506, by a biasing member, such as coil spring508.

Preferably, the base500defines a first passage510that connects the reservoir chamber494and the pocket504. Desirably, the base500also defines a second passage512that connects the reservoir chamber494and the pocket504. A pressure actuated valve arrangement514selectively permits fluid communication through the second passage512when the pressure in the reservoir chamber is above a predetermined threshold. The valve assembly514includes a valve body516biased into a closed position by a biasing member, such as coil spring518. In the closed position, an enlarged diameter upper portion517of the valve body is arranged to block the second passage512to substantially prevent fluid flow therethrough.

Preferably, an upper stop for the valve body516is defined by a snap ring520and a lower stop is defined by a lower end of a valve seat521, which receives the upper portion517of the valve body516. Desirably, the valve body516includes an elongated lower end, or shaft portion522, which functions as a guide for the coil spring518. In addition, preferably a seal member528creates a seal between the valve body516and the base500to inhibit fluid from passing therebetween. Thus, the valve body516is normally biased into a closed position by the force of the biasing member518. If the pressure differential between the reservoir chamber494and the passage526exceeds a predetermined threshold, the valve body516moves toward the open position, against the biasing force of the spring518. In the illustrated arrangement, the predetermined threshold is determined primarily by the surface area of the upper end surface of the valve body516and the spring constant of the biasing member518.

As described above, when the inertia mass492moves into its open position, refilling of the pocket504is restricted to fluid flow between an outer surface of the inertia mass492and an inner surface of the cavity502. In addition, fluid may refill the pocket504by flowing through the passage510, if provided. Thus, the inertia mass492is delayed from moving toward its open position due to the restriction of the fluid from entering the pocket504. However, in the embodiment ofFIG. 32, if the pressure differential between the reservoir chamber494and the passage526exceeds a predetermined threshold, the pressure actuated valve assembly514opens to permit fluid flow into the pocket504through the second passage512. Preferably, the second passage512is configured to permit a greater rate of flow into the pocket504in comparison to fluid flow through the clearance between the inertia mass492and the cavity502and fluid flow through the passage510(if provided). Accordingly, when the pressure actuated valve assembly514opens, the inertia mass492may return to its closed position more quickly.

FIG. 33illustrates an alternative embodiment of a pressure activated inertia valve assembly530. In the embodiment ofFIG. 33, an inertia mass532is configured for axial movement on a shaft534about an axis Ac. Preferably, the inertia mass532is disposed within a reservoir chamber536defined at least partially by a reservoir tube538and a base540. A passage542extends through the base540and shaft534and is in fluid communication with the reservoir chamber536through openings544. Desirably, the passage542receives fluid from a compression chamber (not shown) of the shock absorber, as will be appreciated by one of skill in the art. Thus, the inertia mass532selectively permits fluid communication between the passage542and the reservoir chamber536.

In the embodiment ofFIG. 33, a slide member546is interposed between the base540and the inertia mass532. The slide546includes a recess548that defines a pocket550for receiving the inertia mass532. The inertia mass532is biased into an uppermost, or closed, position (against stop552) by a biasing member, such as coil spring554. The spring554is supported relative to the shaft534by a lower stop, defined by snap ring556. The snap ring556also defines an uppermost position of the slide546. The slide546is also axially moveably relative to the shaft534and is biased into its uppermost position by a biasing member, such as coil spring558.

The base540defines a cavity560, which receives a lower end of the slide546in a sealed arrangement. One of a lower surface562of the cavity560or an upper surface564of the base540function as a stop to define a lowermost position of the slide546. In addition, preferably one or more passages566permit fluid communication between the passage542and a pocket568defined by the cavity560. Preferably, the pocket568is substantially sealed, with the exception of the passages566, such that fluid within the pocket568is at substantially the same pressure as fluid within the passage542(and, thus, the compression chamber of the shock absorber).

In operation, the inertia mass532, upon receiving an appropriate acceleration force, moves in a downward direction relative to the shaft534and into the pocket550. Once in the pocket550, the inertia mass532is delayed in moving in an upward direction due to the restriction of fluid being permitted to refill the pocket550. Thus, the inertia mass532, when positioned within the pocket550, moves toward the closed position at a delayed rate. In the illustrated embodiment, fluid may pass from the reservoir chamber536into the pocket550through a clearance distance C between an outer diameter of the inertia mass532and an inner diameter of the cavity548.

When a difference in fluid pressure between the reservoir chamber536and the passage542(and, thus, the pressure within the compression chamber of the shock absorber) exceeds a predetermined threshold, the slide546moves downward relative to the shaft534and into the pocket568. In the illustrated embodiment, preferably, the predetermined threshold is determined primarily by a surface area of an end surface569the slide546, which is perpendicular to the center axis Acof the shaft534and disposed within the pocket568, along with the spring rate of the biasing member558.

Thus, with the inertia mass532in its open position, the slide546moves in a downward direction away from the inertia mass532. When the slide546moves downwardly a sufficient distance, the inertia mass532is no longer present within the pocket550and fluid may refill the pocket550at a relatively high rate. Thus, the inertia mass532is no longer restricted from moving in an upward direction due to the restriction of fluid moving into the pocket550and, as a result, the biasing member554returns the inertia mass532to its closed position at a normal rate, determined primarily by the weight of the inertia mass532and the spring rate of the spring554. Accordingly, with such an arrangement, when the inertia mass532is in the open position and the pressure within the reservoir chamber536exceeds the pressure within the passage542by a predetermined threshold, the inertia mass532is permitted to return to the closed position without significant delay.

FIGS. 34 and 35illustrate a bicycle that employs yet another alternative embodiment of an acceleration sensitive shock absorber. The bicycle580includes a main frame portion582, an articulating frame portion584, a front wheel586, and a rear wheel588. Preferably, a front suspension assembly590is operably positioned between the front wheel586and the main frame582and a rear suspension assembly, or shock absorber592, is operably positioned between the rear wheel588and the main frame582. Preferably, the articulating frame portion584carries the rear wheel588and the shock absorber592is connected to the articulating frame portion584to resist movement of the rear wheel588in an upward direction. Preferably, the shock absorber592is positioned on one lateral side of the rear wheel588and, desirably, on the left-hand side of the rear wheel588.

With reference toFIG. 35, desirably, the shock absorber592includes a reservoir chamber594at least partially defined by a reservoir tube596and a base598. Preferably, an acceleration sensitive valve assembly600is disposed within the reservoir chamber594. The valve assembly600preferably includes a valve body602biased into an uppermost, or open position, by a biasing member, such as coil spring604. The valve body602is supported for axial movement along an axis Ac, which is defined by a shaft606. An uppermost position of the valve body602preferably is determined by a snap ring608. In the illustrated embodiment, the uppermost position defines a closed position of the valve600.

The base598preferably includes a cavity610that defines a pocket612in which the valve body602enters in its lowermost position. In a preferred arrangement, when the valve body602is in its lowermost position, fluid flow is permitted through openings613of the shaft606. A bottom surface614of the cavity610defines a lower stop for the valve body602. Preferably, as described above, a valve assembly616is provided to permit relatively free flow of fluid from the pocket612to the reservoir chamber594while permitting restricted flow of fluid from the reservoir chamber594into the pocket612.

Desirably, the valve assembly600includes a system for sensing acceleration force inputs and for moving the valve body602to an open position and/or retaining the valve body602in an open position. In the illustrated embodiment, preferably an electromagnetic system618is provided. The system618preferably includes an electromagnetic force generator620within the base598and positioned below the valve body602. A control assembly622is operably connected to the electromagnetic force generator620. Preferably, the valve body602includes a lower portion624, which is constructed from a magnetic material. The electromagnetic force generator620desirably is configured to selectively apply an attractive force to the magnetic portion624of the valve body602. Thus, the valve body602may be moved toward, or retained in, an open position by the electromagnetic force generator620.

With reference toFIG. 34, preferably, a sensor626is positioned on the front suspension assembly590for movement with a hub axis AHof the front wheel586. In addition, or in the alternative, a sensor628may be secured to the articulating frame portion584for movement with a hub axis AHof the rear wheel588. Preferably, each of the sensors626and628are configured to sense substantially vertical acceleration force inputs to the front or rear wheels586,588, respectively.

The sensors626,628are configured to communicate with the control assembly622to provide a control signal indicative of the acceleration forces acting on the front or rear wheels586,588. In a preferred embodiment, the sensors626,628produce an electronic signal to communicate with the control assembly622. In such an embodiment, the sensors626,628may communication with the control assembly622through a hardwired system or, preferably, over a wireless communication system. Furthermore, other suitable types of sensors and methods of communication between the sensors626,628and the control assembly622may also be used, such as hydraulic or mechanical systems, for example. Thus, the control signal may include changes in hydraulic pressure, or movement of a mechanical linkage, for example. Other suitable systems apparent to one of skill in the art may also be used.

The control assembly622preferably includes a processor and a memory for storing a control algorithm, or protocol. The control assembly622uses the control signal provided by the sensors626,628along with the control algorithm to determine whether to activate the electromagnetic force generator620. Thus, when an appropriate acceleration force input is detected, the control assembly622may activate the electromagnetic force generator620to move the valve body602from its closed position into an open position and, if desirable, retain the valve body602in an open position for a period of time, or a delay period.

Desirably, the control assembly622includes an adjustment mechanism, to permit adjustment of the delay period in which the valve body602is held in an open position and/or the acceleration force threshold above which the valve assembly600is opened. Preferably, the control assembly622includes a first adjustment knob630, to permit adjustment of the delay period, and a second adjustment knob632, to permit adjustment of the acceleration force threshold.

The valve body602may be fully controlled by the electromagnetic force generator620or may be configured to be self-responsive to acceleration force inputs due to the inertia of the valve body602. Furthermore, the valve616may be provided to determine a delay period of the valve body602or the electromagnetic force generator620may be relied on to provide the delay in the valve body602from returning to the closed position. In addition, a combination of inertia forces and electromagnetic forces may be utilized to open the valve body602and a combination of fluid restriction, or fluid suction, forces and electromagnetic forces may be utilized to provide the valve body602with a delay period in moving from an open position to a closed position.

Advantageously, by positioning the sensor626to sense acceleration force inputs of the front wheel586, the valve body602in the rear shock absorber592may be moved into its open position before the object (e.g., such as a bump, rock or other irregularity in the trail surface) which caused the acceleration force is encountered by the rear wheel588. Thus, there is no delay in the altered rate of damping of the rear shock absorber592due to the valve body602having to move from its closed position to its open position upon encountering the bump, or other obstacle, because the bump has been “anticipated” by the sensor626positioned to detect acceleration of the front wheel586.

As described above, preferably, the valve body602remains in an open position, or is delayed from returning to its closed position, so that the rear wheel588may absorb a series of bumps and the valve assembly600does not have to reactivate upon encountering each individual bump. Advantageously, by permitting the delay to be controlled by the adjustment mechanism630, a rider can tune the shock absorber592to suit anticipated trail conditions by providing a relatively short or a relatively long delay time. In addition, the acceleration threshold may also be adjusted such the size of bump necessary to open the valve assembly may be varied.

Furthermore, the front suspension assembly590may also be configured to include an acceleration sensitive valve assembly, similar to the valve assembly600. In addition, the various features illustrated inFIGS. 1-35may be used in combination with one another to provide a desired result, as may be determined by one of skill in the art.

Although the present invention has been explained in the context of several preferred embodiments, minor modifications and rearrangements of the illustrated embodiments may be made without departing from the scope of the invention. For example, but without limitation, although the preferred embodiments described an inertia valve damper for altering the rate of compression damping, the principles taught may also be utilized in damper embodiments for altering rebound damping, or for responding to lateral acceleration forces, rather than vertical acceleration forces. In addition, although the preferred embodiments were described in the context of an off-road bicycle application, the present damper may be modified for use in a variety of vehicles, or in non-vehicular applications where dampers may be utilized. Furthermore, the self-centering and timer features of the inertia valve assembly may be applied to other types of valves, which may be actuated by acceleration forces or by means other than acceleration forces. Accordingly, the scope of the present invention is to be defined only by the appended claims.