Accumulators for distributed active suspension system

Abstract

To improve ride comfort quality of a vehicle through appropriate volume determination of accumulators of a suspension system component of the vehicle.SOLUTION: A suspension system component comprises: a piston slidably inserted into a hydraulic cylinder, thereby dividing it into a compression chamber and an extension chamber; a piston rod extending out of the hydraulic cylinder; a hydraulic pump that includes a first port in fluid communication with the compression chamber and a second port in fluid communication with the extension chamber; a compression accumulator arranged to exchange fluid with the compression chamber; and an extension accumulator arranged to exchange fluid with the extension chamber. When the hydraulic pump generates a first preset pressure differential of at least 68.95 bar, an observed stiffness of the suspension system component in response to an external input having a frequency of 12 Hz and a peak-to-peak amplitude of 5 mm does not exceed 80 N / mm.SELECTED DRAWING: Figure 3

Description

[Technical Field]

[0001] This application claims the benefit under 35 U.S.C. §119(e) of U.S. Provisional Patent Application No. 62 / 684,899, filed June 14, 2018, U.S. Provisional Patent Application No. 62 / 691,132, filed June 28, 2018, and U.S. Provisional Patent Application No. 62 / 740,823, filed October 3, 2018, the disclosures of each of which are incorporated by reference in their entirety. [Background technology]

[0002] Active suspension systems for vehicles have the potential to solve some of the known problems and compromises associated with passive or semi-active suspension systems. However, efforts to implement active suspension systems in production vehicles have been hampered by several complex problems whose causes and solutions are not yet understood in the art. Summary of the Invention

[0003] In one aspect, a suspension system component is disclosed that includes a hydraulic cylinder at least partially defining an interior space; a piston slidably inserted into the hydraulic cylinder, thereby dividing the interior space into a compression chamber and an expansion space; a piston rod attached to the piston and extending outward from the hydraulic cylinder; a hydraulic pump including a first port in fluid communication with the compression chamber and a second port in fluid communication with the expansion chamber; a compression accumulator configured to exchange fluid with the compression chamber; and an expansion accumulator configured to exchange fluid with the expansion chamber. In certain embodiments, when the hydraulic pump generates a first set differential pressure, the observed stiffness of the suspension system component in response to an external input having a frequency of 12 Hz and a peak-to-peak amplitude of 5 mm does not exceed 80 N / mm, and the first set differential pressure has a value of at least 1,000 psi. In certain embodiments, the first set differential pressure has a value of 1,000 psi. In certain embodiments, the observed stiffness is 5 N / mm, 10 N / mm, or 25 N / mm or greater. In certain embodiments, the observed stiffness is no greater than 80 N / mm, 70 N / mm, or 50 N / mm.

[0004] In another aspect, a vehicle is disclosed that includes a suspension system that includes the aforementioned suspension system components. In certain embodiments, the vehicle includes a suspension system that includes a plurality of suspension system components as described herein. In certain embodiments, the vehicle further includes a sprung mass and an unsprung mass, and each suspension system component of the plurality of suspension system components is positioned between the unsprung mass of the vehicle and the sprung mass of the vehicle. In certain embodiments, each suspension system component of the vehicle is fluidly decoupled from any other suspension system components of the vehicle.

[0005] In another aspect, a suspension system component for a vehicle is disclosed, the suspension system component including: a hydraulic cylinder containing a first fluid quantity; a piston slidably received in the hydraulic cylinder, thereby dividing at least a portion of the hydraulic cylinder into an extension chamber and a compression chamber; a piston rod attached to the piston; a compression accumulator containing a second fluid quantity, the compression accumulator in direct fluid communication with the compression chamber; an expansion accumulator containing a third fluid quantity, the expansion accumulator in direct fluid communication with the expansion chamber; and a pump in fluid communication with both the compression chamber and the expansion chamber, the second quantity being greater than the first quantity and the third quantity being greater than the first quantity. In certain embodiments, the second quantity is greater than the first quantity by a first factor that is 3 or greater. In certain embodiments, the first quantity is greater than the second quantity by a second factor that is 3 or greater. In certain embodiments, the first factor and the second factor are each less than 10. A vehicle suspension system including a plurality of the suspension system components and a vehicle including the suspension system are also disclosed.

[0006] In yet another aspect, a recessed accumulator assembly is disclosed that includes a first accumulator including a first accumulator and a second accumulator. The first accumulator can include a first accumulator housing at least partially defining a first interior space and a first piston dividing the first interior space into a first gas-filled chamber and a first hydraulic fluid chamber. The second accumulator can include a second accumulator housing at least partially defining a second interior space and a second piston dividing the second interior space into a second gas-filled chamber and a second hydraulic fluid chamber. At least a portion of the second accumulator can be disposed within the first interior space.

[0007] In certain embodiments, the recessed accumulator assembly includes one or more ports that allow fluid to enter and exit the first hydraulic fluid chamber and one or more ports that allow fluid to enter and exit the second hydraulic fluid chamber. In certain embodiments, the second accumulator is disposed completely within the first interior space. In certain embodiments, the second accumulator is disposed at least partially within the first hydraulic fluid chamber. In certain embodiments, the entire second accumulator is disposed within the first hydraulic fluid chamber. In certain embodiments, a portion of a wall of the first accumulator housing is attached to a portion of a wall of the second accumulator housing. In certain embodiments, the first accumulator housing and the second accumulator housing share at least a portion of a wall.

[0008] In certain embodiments, the recessed accumulator assembly may include a gap between an outer surface of the first accumulator housing and an inner surface of the second accumulator housing. In certain embodiments, the gap functions as a fluid restricting element that restricts flow into the first hydraulic fluid chamber. In certain embodiments, the gap functions as a fluid restricting element that restricts flow from the first hydraulic fluid chamber. In certain embodiments, the gap functions as a fluid inertance element.

[0009] The accompanying drawings are not intended to be drawn to scale. In the drawings, each identical or nearly identical component shown in various figures may be represented by the same numeral. For clarity, not every component may be labeled in every figure. [Brief explanation of the drawings]

[0010] [Figure 1] 1 shows various exemplary shapes of surfaces on which a vehicle can travel. [Figure 2a] The damper is shown at one stage of the compression / expansion stroke. [Figure 2b] The damper is shown at one stage of the compression / expansion stroke. [Figure 2c]The damper is shown at one stage of the compression / expansion stroke. [Figure 3] 1 illustrates an embodiment of an active suspension system component. [Figure 4] 1 illustrates an embodiment of an active suspension system component in a first suspension position. [Figure 5] 5 shows the embodiment of FIG. 4 in a second suspension position. [Figure 6] 5 shows the embodiment of FIG. 4 in a third suspension position. [Figure 7] 1 illustrates an embodiment of an active suspension system component including one or more valves. [Figure 8] 1 illustrates an embodiment of an active suspension system component including one or more valves. [Figure 9] 1 shows a graph illustrating the relationship between observed stiffness and set differential pressure. [Figure 10] 1 illustrates an embodiment of an active suspension system component including a compliant mechanism. [Figure 11] 1 shows a graph illustrating the relationship between compliant mechanism stiffness and set differential pressure. [Figure 12] 1 shows a graph illustrating the relationship between observed stiffness and set differential pressure for a suspension system component having a compliant mechanism. [Figure 13] 1 illustrates an embodiment of a single hydraulic accumulator. [Figure 14] 1 illustrates an embodiment of a recessed accumulator assembly. DETAILED DESCRIPTION OF THE INVENTION

[0011] Vehicles typically include a suspension system that includes various suspension system components. A vehicle's hydraulic suspension system component may include a piston slidably received in a hydraulic cylinder, thereby dividing at least a portion of the hydraulic cylinder into a compression chamber and an expansion chamber. This piston / cylinder configuration is typically referred to as a damper. The damper may be disposed between the vehicle's unsprung mass (which may include, for example, a wheel or wheel assembly) and the vehicle's sprung mass (which may include, for example, a vehicle chassis or body). If the damper has a relatively low stiffness (e.g., the damper is sufficiently "flexible"), the damper allows some movement of the unsprung mass while at least partially impeding movement of the sprung mass. For example, in a vehicle with a sufficiently flexible suspension system, the vehicle's wheels may travel over rough terrain while the vehicle body remains relatively level, resulting in a ride that is perceived as comfortable by vehicle occupants. On the other hand, if the damper has a relatively high stiffness (e.g., the damper is sufficiently "stiff"), the movement of the unsprung mass may be transmitted from the wheel through the damper to the vehicle body with only minimal loss. As a result, vehicles with overly stiff suspension systems may be perceived as providing an uncomfortable or "rough" ride for occupants.

[0012] In a hydraulic active suspension system component, a hydraulic pump may be fluidly disposed between a damper compression chamber and a damper expansion chamber. As recognized by the inventors and described herein, rotation of this hydraulic pump (e.g., during damper compression or expansion) may involve inertia, which may result in an undesirably high damper stiffness during certain driving conditions, resulting in a harsh ride. Furthermore, as recognized by the inventors, the introduction of a hydraulic pump causes the observed system stiffness to depend on the frequency of the input; in particular, the system may exhibit an acceptably low stiffness in response to inputs having a relatively low frequency, but an undesirably high stiffness in response to inputs having a relatively high frequency. In particular, apparatus and methods are described herein for at least partially mitigating the effects of pump inertia on the performance of hydraulic suspension system components.

[0013] Furthermore, as described in detail herein, active suspension systems are capable of operating over a much wider range of fluid pressures and a wider range of sustained differential pressures than passive or semi-active suspension systems. As described herein, due to both pump inertia and the range of pressures and differential pressures seen by active suspension system components, as well as other reasons, the inventors have identified the root cause of various problems with the operation of active suspension system components. Described herein are various methods and apparatus that allow for improved performance of active suspension systems, including suspension system components configured to maintain a desired stiffness throughout the entire operating range and in response to a wide range of input frequencies, thereby allowing actively suspended vehicles to achieve a more desirable ride quality over a variety of driving conditions.

[0014] Referring now to the figures, several non-limiting embodiments of various vehicles, suspension systems, and suspension system components are described in detail below. It will be appreciated that the various systems, components, features, and methods described in connection with these embodiments may be used individually and / or in any desired combination, and the present disclosure is not limited to only the specific embodiments described herein.

[0015] Figure 1 shows a vehicle 1 moving on a road surface 2. The wheels 3 of the vehicle 1 contact the road surface at contact patch 4. As the vehicle moves across the road surface, certain vertical displacements 5 of the road surface are of sufficient magnitude to be substantially greater than can be absorbed by the suspension system. These displacements are referred to as Class I displacements. The vehicle 1 may include a passive or semi-active suspension system (not shown) that includes dampers (also called shock absorbers) and / or springs.

[0016] A close-up portion 6 of the road immediately surrounding vehicle 1 shows the road surface containing various smaller vertical displacements 7. These displacements are smaller than Class I displacements 5 and are generally within the operating range of the suspension system of vehicle 1. These displacements are referred to herein as Class II displacements. In typical vehicle travel at typical speeds (20-50 mph), Class II road-induced displacements are effectively transmitted to the vehicle body, but can be quickly damped by shock absorbers in vehicles with passive and semi-active suspension systems.

[0017] The further enlarged portion 8 of the road surface immediately surrounding the wheel 3 indicates that the road surface includes additional vertical displacements 9. These additional displacements, referred to herein as Class III displacements, are smaller than Class II displacements. The vertical motion of the unsprung mass induced by moving through these Class III road displacements is typically transmitted to the unsprung mass but effectively fails to reach the sprung mass, including the vehicle body, and can be further damped by the dampers of a passive or semi-active suspension system. For purposes of this disclosure, Class I displacements are considered to be displacements having a characteristic range greater than the operating range of the suspension system of the vehicle 1; Class II displacements are displacements having a characteristic range less than the operating range of the suspension system and greater than one-eighth of the operating range of the vehicle's suspension system; and Class III displacements are displacements having a characteristic range less than one-eighth of the operating range of the vehicle's suspension system.

[0018] Ideally, a vehicle's suspension system would absorb virtually all perturbations from road inputs, thereby eliminating any body motion. However, suspension systems typically have a limited operating range and therefore cannot practically absorb perturbations that approach or exceed such limits. The operating range of most passive and semi-active shock absorbers is typically in the range of 10 cm to 15 cm. Therefore, road elevation variations beyond these limits (e.g., road elevations can vary by several meters or more) cannot be absorbed by the vehicle's suspension system. Therefore, only the effects of road elevation changes that fall within the operating range of the vehicle's shock absorbers can be optimally mitigated. For example, the suspension system can at least partially suppress body motion due to Class II and Class III displacements, but not Class I displacements.

[0019] If the dampers of vehicle 1 are made sufficiently stiff, vertical motion induced by traveling over an uneven road surface can be effectively (and undesirably) transmitted from the unsprung mass through the dampers to the vehicle body. On the other hand, if the dampers (along with any parallel spring elements) are sufficiently flexible, the suspension system can "absorb" motion of the unsprung mass within a certain frequency range while simultaneously minimizing motion of the sprung mass (including, for example, the vehicle body). To minimize such transmission of disturbances from the road surface to the vehicle body, the vehicle's suspension system components can include dampers designed to be sufficiently flexible for motion within a certain frequency range (e.g., 10-15 Hz).

[0020] 2a-2c illustrate an exemplary hydraulic damper 101 during various stages of compression / expansion. The hydraulic damper includes a piston 107 slidably received in a hydraulic cylinder 103, thereby dividing at least a portion of the cylinder into a compression chamber 111 and an expansion chamber 109. A piston rod 113 may be attached to the piston 107 and extend out of the hydraulic cylinder 103. The length of the damper, designated L, is interpreted as the distance between the base of the hydraulic cylinder 103 and the tip of the piston rod 113. This length is determined by the position of the piston within the cylinder; moving the piston relative to the cylinder 103 in an extension direction 121 expands the damper (increases the length L), while moving the piston relative to the cylinder 103 in a compression direction 119 compresses the damper (decreases the length L). The damper further includes a flow passage 123 fluidly connecting the compression chamber 111 to the expansion chamber 109, and a flow control element 125 fluidly disposed along the flow passage 123. In a passive suspension system, the flow control element 125 may be a fixed width orifice. In a semi-active suspension system, the flow control element 125 may be a controllable (e.g., solenoid) valve or a variable throttle. In an active suspension system, the flow control element 125 may alternatively or additionally include a hydraulic pump. While the flow passage is illustrated as being external to the cylinder, in various embodiments, the flow passage may instead include at least a portion that is completely within the cylinder (e.g., the flow passage may be at least partially contained within the piston 107).

[0021] Continuing with reference to FIGS. 2a-2c, upon compression of the hydraulic damper (e.g., upon a change in damper position from that shown in FIG. 2a to that shown in FIG. 2c), an increased amount of piston rod 113 enters cylinder 103, thereby reducing the amount of volume within cylinder 103 available to fluid. As a result, fluid is displaced out of cylinder 103. As shown, accumulator 127 is available to accommodate the displaced fluid. Additionally, upon expansion of the hydraulic damper (e.g., upon a change in damper position from that shown in FIG. 20a to that shown in FIG. 20b), a decreased amount of piston rod 113 remains in cylinder 103, thereby increasing the amount of volume within cylinder 103 available to fluid. As a result, fluid flows from accumulator 127 into cylinder 103. For example, when the piston 107 contacts the cylinder housing or one or more "stops" attached to the cylinder housing, it can cause maximum expansion of the damper (as shown in FIG. 2b) or maximum compression of the damper (as shown in FIG. 2c), so that the damper cannot be expanded or compressed further.

[0022] The amount of fluid contained within the hydraulic cylinder 103 at a given time can vary as the piston rod is inserted into and / or withdrawn from the hydraulic cylinder 103. As shown in FIG. 2b, when the damper is fully extended, a maximum fluid volume (V max As shown in Figure 2c, when the damper is fully compressed, the minimum amount of fluid that can be accommodated in the hydraulic cylinder (V min As will be appreciated by those skilled in the art, the embodiments shown herein are exemplary. It will be appreciated that various damper configurations (e.g., various piston shapes, asymmetric piston configurations, various rod arrangements, etc.) may be utilized.

[0023] Therefore, when the vehicle is in operation, the amount of fluid contained in the hydraulic cylinder is V min and V maxThe volume of fluid in the cylinder can vary between 0.01 and 0.01. To address the issue of changing fluid volume in the cylinder as the damper expands and compresses, it is known to use an accumulator 127, as shown in FIG. 2. The accumulator 127 can include a floating piston or bladder that separates a fluid chamber from a pressurized gas chamber. Alternatively, the accumulator can include a spring-loaded piston. The accumulator can be configured to receive fluid from the internal chamber as the damper compresses. Additionally, the accumulator can be further configured to supply fluid to the internal chamber as the damper expands. The accumulator can assist with fluid volume changes with temperature changes.

[0024] Therefore, in a passive damper, the accumulator volume is typically sized to accommodate the change in fluid volume between maximum compression and maximum extension of the damper. In other words, the accumulator volume is V max -V min A relatively small amount of additional capacitance may be included in the accumulator to account for temperature effects, but apart from this small amount of additional capacitance, the accumulator capacitance is typically approximately equal to V max -V min is kept relatively close to the value given by

[0025] FIG. 3 illustrates embodiments of suspension system components for an active suspension system. In the illustrated embodiment, a hydraulic pump 201 is shown fluidly disposed between the damper's compression chamber 111 and the damper's expansion damper 109. In various embodiments, the pump may be a gear pump (e.g., an internal gear pump (e.g., a gerotor, a crescent pump), an external gear pump), a vane pump, a variable displacement pump, or any other type of suitable pump known in the art. As illustrated, the pump 201 may be operably coupled to a motor 205. The motor may be, for example, an electric motor (e.g., a BLDC motor). The suspension system components illustrated in FIG. 3 may be used, for example, in an active suspension system of a vehicle.

[0026] In the illustrated active suspension system component of Figure 3, pump 201 can be utilized to actively compress or expand the damper. For example, when active compression is desired, motor 205 can be used to actively drive pump 201 in a first direction to cause a net fluid flow from compression chamber 211 to expansion chamber 209. For active expansion, motor 205 can actively drive pump 201 in a second direction to cause a net fluid flow from expansion chamber 209 to compression chamber 211. Thus, the use of an active suspension system allows improved control over vehicle position (e.g., control over the distance between the sprung and unsprung masses) compared to passive or semi-active suspension systems.

[0027] In this disclosure, the terms "active compression" and "active expansion" are used to describe situations in which the damper is compressed or expanded, respectively, due to active operation of the pump. In active compression or expansion, the motor 205 can receive electrical energy from an electrical energy source (e.g., a battery, capacitor, alternator, or generator) and convert the received electrical energy into mechanical energy in the form of torque applied to the pump 201. The pump 201 can then function to convert the mechanical energy (e.g., applied torque) into hydraulic energy (e.g., fluid differential pressure, flow, etc.). Furthermore, the term "active force" refers to a force imparted to the piston due to a pressure differential created by the active operation of the pump. In contrast, the term "external force" refers to a force imparted to the piston due to an external input, which is interpreted as an input not caused by the active operation of the pump. External inputs can include inputs resulting from the shape of the road surface (e.g., dips, bumps), forces caused by cornering, braking, or accelerating the vehicle, or forces / displacements applied to the piston rod by a dynamometer during evaluation. In this disclosure, the terms "external compression" and "external expansion" are used to describe compression or expansion, respectively, of the damper caused by an external input and not by active operation of the pump. During operation of a motor vehicle, the damper may experience external compression due to external forces induced by external inputs, such as, for example, ground inputs (e.g., driving over a bump in the road) and / or vehicle maneuvers, such as braking, accelerating, and cornering. Additionally, the damper may experience external expansion due to external forces induced by, for example, ground inputs, such as driving over a pothole in the road, and / or vehicle maneuvers, such as braking, accelerating, and cornering.

[0028] Considering FIG. 3 and assuming the pump is deactivated (i.e., the pump is not actively operating and does not generate a pressure differential), an external input resulting in a compressive force acting on the piston rod can pressurize the fluid in the compression chamber 111 more than the fluid in the expansion chamber 109, thereby creating a pressure differential between the fluid in the compression chamber 111 and the fluid in the expansion chamber 109. This pressure differential persists as long as the external input (e.g., road input) continues. When this pressure differential is sufficient to overcome the inertia associated with the pump, fluid can flow from the compression chamber 111, through the pump 201 (thereby rotating the pump), and into the expansion chamber 109. This can compress the damper, which in turn damps or otherwise prevents excessive movement of the sprung mass in response to the external input. As one skilled in the art will appreciate, this behavior is very similar to that of a passive damper.

[0029] However, if the rotational inertia associated with the pump is large, the pressure differential resulting from a given external input cannot overcome the rotational inertia associated with the pump 201. In such cases, the pump 201 may remain in a fixed angular position or may not be able to rotate fast enough to deliver the required amount of fluid within the available time, thereby effectively impeding or slowing fluid flow out of the compression chamber 111. In this case, the hydraulic damper 103 may not be able to compress or contract in response to the external input (e.g., the hydraulic damper 103 may be "hydraulically locked") or may not be able to compress or contract quickly enough in response to the external input. Furthermore, a similar reduction may occur for external inputs in the extension direction 121. Thus, the damper may exhibit undesirably high stiffness, and road inputs may be transmitted from the unsprung mass through the damper to the vehicle body without or only minimal isolation or damping. As a result, ride quality may be reduced, resulting in a "rough" ride. The observed stiffness may be particularly high in response to (a) high accelerations in damper operation, (b) small amplitude external inputs or external inputs resulting in low accelerations that are unable to generate a pressure differential sufficient to overcome the inertia associated with the pump, and / or (c) high frequency external inputs that are unable to generate a pressure differential for a sufficient period of time to overcome the inertia of the pump.

[0030] The rough ride problem may be more pronounced during active pump operation. Returning to FIG. 3 , motor 205 can drive pump 201 to generate a set pressure differential by rotating in either a clockwise or counterclockwise direction. For example, motor 205 can drive pump 201 to generate a set pressure differential such that fluid pressure in compression chamber 111 exceeds fluid pressure in expansion chamber 109, thereby applying an active force in extension direction 121. This can be done, for example, to maintain a vehicle body substantially level during a cornering event. If an external input force in compression direction 119 is applied to the piston rod while the pump is being driven to generate the set pressure differential (e.g., when the vehicle hits a bump while simultaneously operating the active suspension system to support the vehicle body during cornering), fluid flow between the hydraulic chambers may be effectively blocked by the inertial mass of the pump. That is, unless the external input is large enough to overcome not only the inertia of the pump but also the torque applied to the pump 201 by the motor 205, the hydraulic damper will exhibit undesirably high stiffness in response to the external input, and the force of the external input may be transmitted through the hydraulic damper 101 to the vehicle body.

[0031] The inventors have recognized that, in certain aspects, this "rough ride" problem can be at least partially mitigated by utilizing a suspension system component having two separate accumulators, as shown in the exemplary embodiment of FIGS. 4-6. The two accumulators may include a compression accumulator 350 and an expansion accumulator 352. As shown, in certain embodiments, the compression accumulator 350 may be in direct fluid communication with the compression chamber 111. As used herein, the compression accumulator is said to be in "direct" fluid communication with the compression chamber because fluid can flow from the compression accumulator 350 to the compression chamber 111 without passing through a pump. Additionally, in certain embodiments, the expansion accumulator 352 may be in direct fluid communication with the expansion chamber 109. In certain embodiments, one or more valves (e.g., semi-active valves, variable orifice valves, etc.) may be fluidly disposed between the compression chamber 111 and the compression accumulator 350 and / or between the expansion chamber 109 and the expansion accumulator 352.

[0032] In the illustrated embodiment, during external compression of the damper, even when the pump's inertia is large and the differential pressure created by the external input is overcome, fluid can flow from the compression chamber 111 to the compression accumulator 350, and simultaneously, fluid can flow from the expansion accumulator 352 to the expansion chamber 109. As a result, the damper can compress without requiring any fluid flow through the pump. FIGS. 4-6 show the suspension system components at various times during the compression stroke. In FIG. 5, the damper is compressed compared to FIG. 4. As can be seen by comparing FIG. 5 with FIG. 4, this compression can reduce the amount of fluid contained in the compression chamber 111 (assuming no fluid flow through the pump), increase the amount of fluid contained in the compression accumulator 250, reduce the amount of fluid contained in the expansion accumulator 352, and increase the amount of fluid contained in the expansion chamber 109 (liquid is shown as a light gray fill for clarity). These trends continue as the damper is further compressed, as can be seen by comparing FIG. 6 with FIG.

[0033] Thus, the net effect of external compression is fluid flow out of the compression chamber 111 and fluid flow into the expansion chamber 109, but notably, without any fluid having to pass through the hydraulic pump 201. Furthermore, in the case of external expansion, fluid can flow from the expansion chamber 109 into the expansion accumulator 352, and simultaneously, fluid can flow from the compression accumulator 350 into the compression chamber 111. Thus, the net effect during external expansion is fluid flow out of the expansion chamber 109 and fluid flow into the compression chamber 111, again without the fluid having to pass through the hydraulic pump 201. Thus, the inertia associated with the pump can be effectively ignored, and rough ride problems can be avoided. This configuration can make the system sufficiently flexible with respect to external compression, but can also make the system less responsive under similar operating conditions.

[0034] In certain embodiments, both the compression accumulator and the expansion accumulator are disposed outside the hydraulic cylinder. In certain embodiments, as shown in FIGS. 4-6, each accumulator includes an accumulator housing 354a-b that at least partially defines an internal accumulator chamber. In certain embodiments, the accumulator housings 354a-b may be cylindrical. In certain embodiments, the accumulator may further include barriers 356a-b that divide the internal accumulator chamber into liquid chambers 358a-b at least partially filled with liquid and gas chambers 360a-b at least partially filled with gas. In the exemplary embodiment of FIGS. 4-6, the barriers 360a-b are pistons slidably received within the accumulator housings 354a-b. In other embodiments, the barriers 360a-b may be flexible bladders. As used herein, the "volume" of an accumulator refers to the total amount of fluid disposed within the accumulator housing. Thus, for example, in the case of an accumulator having separate liquid and gas chambers 358a-b, 360a-b, the volume of the accumulator is understood to refer to the sum of (a) the amount of liquid in the liquid chambers at a given time and (b) the volume of gas in the gas chambers at a given time.

[0035] The inventors have recognized that, in certain embodiments, to achieve desired performance throughout the entire travel of the damper, the compression accumulator is preferably sized to accommodate the entire amount of fluid displaced from the compression chamber during a complete compression stroke (e.g., from maximum extension to maximum compression). Additionally, the expansion accumulator is preferably sized to accommodate the entire amount of fluid displaced from the expansion chamber during a complete extension stroke (e.g., from maximum compression to maximum extension). Accordingly, in certain embodiments, at any given time, the amount of fluid contained in the compression accumulator exceeds the amount of fluid contained in the damper's hydraulic cylinder. Furthermore, in certain embodiments, at any given time, the amount of fluid contained in the expansion accumulator exceeds the amount of fluid contained in the damper's hydraulic cylinder. This makes the accumulators, respectively, significantly larger than accumulators typically found in passive or semi-active suspension systems.

[0036] In the active suspension system components shown in Figures 4-6, the volume of fluid contained in the compression accumulator 350 is approximately equal to the volume of fluid contained in the damper's hydraulic cylinder. As shown in Figure 6, as the hydraulic damper approaches maximum compression, the volume of liquid in the damper's compression chamber 111 approaches zero, and the volume of liquid contained in the compression accumulator's 350 liquid chamber increases. Because the compression accumulator includes both a gas chamber and a liquid chamber, as the volume of the liquid chamber increases, the volume of the gas chamber correspondingly decreases. According to well-known gas laws, such compression of the gas chamber increases the pressure of the gas held in the gas chamber, thereby increasing the effective stiffness of the compression accumulator. As a result, the stiffness of the compression accumulator increases as the damper is compressed. If the volume of the compression accumulator were approximately equal to the volume of fluid in the damper's hydraulic cylinder, the compression accumulator could become excessively stiff when the damper is compressed. This can lead to a harsh ride when the damper is fully compressed. For similar reasons, the expansion accumulator can become too stiff when the damper 101 is expanded, which can lead to a harsh ride when the damper is fully expanded.

[0037] To avoid the compression accumulator becoming too stiff during a large external compression event, the inventors have recognized that the compression accumulator can be sized so that the amount of fluid contained in the compression accumulator exceeds the amount of fluid contained in the damper's hydraulic cylinder by a first factor. In certain embodiments, particularly those in which a flexible compression accumulator is desired, the first factor can be greater than 3. In other embodiments, the first factor can be greater than 1.5. The first factor may be limited by packaging constraints, as larger compression accumulator volumes are more difficult to package within the constrained space of a vehicle. In various embodiments, the first factor cannot exceed 5, 2.5, or 2.

[0038] For similar reasons, to avoid the compression accumulator becoming too stiff during large external expansion events, the inventors have recognized that the expansion accumulator can be sized so that the amount of fluid contained in the expansion accumulator exceeds the amount of fluid contained in the damper's hydraulic cylinder by a second factor. In certain embodiments, particularly those in which a flexible expansion accumulator is desired, the second factor can be greater than 3. In other embodiments, the second factor can be greater than 1.5. The second factor may be limited by packaging constraints, as larger expansion accumulator volumes are more difficult to package within the constrained space of a vehicle. In various embodiments, the second factor cannot exceed 5, 2.5, or 2.

[0039] Due to limited space available for implementation in automotive applications, in certain circumstances, it may be impractical or even impossible to install accumulators having the relatively large capacities defined above in a vehicle's suspension system. For example, it may be impractical or impossible to implement suspension system components in a vehicle that include compression accumulators whose fluid volume exceeds the fluid volume of the damper's hydraulic cylinders and / or expansion accumulators whose fluid volume exceeds the fluid volume of the damper's hydraulic cylinders. In light of such implementation constraints, the inventors have recognized that various modifications can be made to active suspension system components to enable the components to maintain the desired stiffness throughout the damper's travel, even when accumulators with smaller capacities than those previously disclosed are used. These modifications may include, for example, appropriately sized bypass valves, pressure relief mechanisms, the addition of compliance elements, and / or other compliant mechanisms, as disclosed herein.

[0040] FIG. 7 illustrates an embodiment of an active suspension system component. The embodiment of FIG. 7 is similar to the embodiment of FIG. 4 and incorporates a bypass flow path 401 including one or more bypass valves 403. The bypass valves 403 can be opened to allow flow from the compression chamber 111 through the bypass flow path 401 to the expansion chamber 109 or from the expansion chamber 109 through the bypass flow path 401 to the compression chamber 111 during a large amplitude external compression or expansion event (e.g., such as that caused by moving a Class I or Class II displacement), respectively. In this manner, at least during certain operating modes of the bypass valve, fluid can flow between the compression chamber 111 and the expansion chamber 109 without requiring rotation of the pump 201 (e.g., pump inertia can be effectively ignored). Thus, the compression accumulator 350 and / or the expansion accumulator 352, respectively, can be opened to allow flow from the compression chamber 111 through the bypass flow path 401 to the expansion chamber 109 without requiring rotation of the pump 201 (e.g., pump inertia can be effectively ignored). max and V min, because the compression accumulator 350 and the expansion accumulator 352 no longer need to contain all of the fluid from the compression chamber 111 and the expansion chamber 109 during external compression or external expansion, respectively.

[0041] In certain embodiments, the bypass valve 403 can include a pair of blow-off valves, each blow-off valve of the pair being oriented in an opposite direction relative to the other blow-off valve of the pair. In these embodiments, each blow-off valve can be constructed to have a respective cracking pressure that exceeds the maximum operating differential pressure the pump is configured to generate. In certain embodiments, the bypass valve 403 can include a pair of pressure-balanced blow-off valves. In certain embodiments, the bypass valve 403 can include at least one frequency-dependent check valve. In certain embodiments, the bypass valve 403 can include a pair of frequency-dependent check valves, each frequency-dependent check valve of the pair being oriented in an opposite direction relative to the other valve of the pair. Examples of available frequency-dependent check valves are described on page 245 of Shock Absorber Handbook by John C. Dixon, ISBN 978-0-470-51020-9. In preferred embodiments, each frequency-dependent check valve can be configured to open (e.g., allow flow from one side of the valve to the other) when the differential pressure across the valve changes to a frequency that exceeds a first threshold frequency and to close (e.g., substantially prevent flow from one side of the valve to the other) when the differential pressure changes to a frequency that does not exceed the threshold frequency. Such a frequency-selective valve can be configured to close during active pump operation (typically where the frequency variation of the differential pressure is relatively small) and open to external compression or expansion caused by a minor ride discomfort event (typically where the frequency variation of the differential pressure is relatively large). Additionally or alternatively, the frequency-dependent check valve can be configured to at least partially open in response to an external input having a frequency that exceeds the first threshold frequency and to at least partially close in response to an external input having a frequency that does not exceed the first threshold frequency. In various embodiments, the first threshold frequency can be between 3 Hz (approximating the natural frequency of a typical vehicle's body) and 12 Hz (approximating the wheel hop frequency of a typical vehicle).

[0042] FIG. 7 illustrates exemplary active suspension system components when the pump 201 is actively operating to establish a set pressure differential. The term "set pressure differential" is interpreted to refer to a pressure differential created by actively driving the pump using a motor. In FIG. 7, torque can be applied to the pump 201 in a first direction 601 to command the electric motor (not shown) to generate a set pressure differential in the pump 201, causing the pressure of the fluid in the expansion chamber 109 to exceed the pressure of the fluid in the compression chamber 111 by a differential amount. For purposes of illustration, assume that the bypass valve 403 is closed, the length of the damper 101 is fixed, and the liquid in the system is effectively incompressible. Because the damper length is assumed to be fixed and the liquid is effectively incompressible, there is effectively no fluid flow into or out of the expansion chamber 111 and effectively no flow into or out of the compression chamber 109.

[0043] First, when the electric motor applies torque to the pump 201, the pump 201 rotates, causing a portion of the fluid to flow from the liquid chamber 358a of the compression accumulator 350, through the pump 201, and into the liquid chamber 358b of the expansion accumulator 352. As liquid flows out of the liquid chamber 358a of the compression accumulator 350, the volume of the gas chamber 360a of the compression accumulator 350 correspondingly increases (the pressure of the gas contained in each gas chamber 360a decreases). Furthermore, as liquid flows into the liquid chamber 358b of the expansion accumulator 352, the volume of the gas chamber 360b of the expansion accumulator 352 correspondingly decreases (the pressure of the gas contained in each gas chamber 360b increases). The difference in pressure between the gas in gas chamber 360a of compression accumulator 350 and the gas in gas chamber 360b of expansion accumulator 352 is equal to the set differential pressure. In theory (assuming there are no leaks across pump 201, piston 107, and valve 403), after the set differential pressure is established, there is no more fluid flow in the system and the pump remains in a constant angular position. In practice, there will always be some amount of leakage across pump 201, piston 107, and / or valve 403, so the pump may need to rotate continuously to compensate for the leakage and maintain the set differential pressure.

[0044] As previously discussed, the observed stiffness of the active suspension system components may reach undesirably high levels in response to external inputs that occur while the pump is actively operating to apply a pressure differential. Such a situation may occur, for example, when a vehicle travels over a road feature (e.g., a bump, a pothole) while cornering. The natural tendency of a vehicle during cornering is for the side of the vehicle closest to the center of rotation to lift. Therefore, it is desirable to actively operate the active suspension system component pump 201 during a cornering event to establish a set pressure differential to level the vehicle and / or minimize vehicle roll. During such a cornering event, the inside wheel of the vehicle may experience an external input (e.g., by traveling over a road feature), which may result in an external input being applied to the piston in the extension direction 801. Because the active suspension system components become stiffer in response to the external inputs during active operation of the pump, a degradation in ride comfort may occur if certain external inputs occur simultaneously with the active operation of the pump 201.

[0045] The increase in observed stiffness and the corresponding decrease in ride comfort may be particularly noticeable for relatively high-frequency external inputs (e.g., a series of small, repetitive bumps on the ground surface) compared to relatively low-frequency external inputs. One interpretation of such frequency dependence of observed stiffness is that external inputs with sufficiently low frequencies are less affected by the inertia of the pump 201, thereby allowing fluid to flow through the pump and lowering the observed stiffness. Relatively high-frequency external inputs, such as those associated with wheel hop, may require high acceleration of the pump and therefore may be more susceptible to the inertia of the pump. In response to such relatively high-frequency external inputs, flow through the pump (ignoring leakage) may be effectively blocked, thereby effectively increasing the observed stiffness. As an alternative or additional explanation, when an actively controlled bypass valve is used to bypass the pump, there is usually a minimum response time associated with opening and closing the bypass valve. For high-frequency external inputs, the valve may become ineffective if the duration of the event is shorter than the minimum response time required by the bypass valve. Additionally, fluid mass within the various flow paths of suspension system components can effectively impede fluid flow at high frequencies, resulting in fluid inertance and / or impedance that can increase the observed stiffness.

[0046] To avoid such degradation in ride quality, the inventors have recognized that the observed stiffness of suspension system components must be considered over the entire range of expected setpoint differential pressures and across a range of input frequencies. FIG. 9 illustrates the relationship between the stiffness of an exemplary suspension system component and the setpoint differential pressure, shown on the x-axis, in response to an external input having a frequency of 12 Hz, shown on the y-axis. A vertical dashed line 901 is shown in FIG. 16 to indicate the maximum setpoint differential pressure expected during normal driving conditions. The inventors have recognized that the maximum setpoint differential pressure can be approximately 1,000 psi for a vehicle having a typical weight, piston size, and piston rod size in a normal driving environment on a standard terrain. Additionally, a horizontal dashed line 903 is shown in FIG. 5 to highlight the desired maximum stiffness in response to a 12 Hz external input (which is close to the wheel hop frequency of a typical vehicle). Curves 910-918 show the observed stiffness of suspension system components with various external accumulator capacities (curve 910 represents the component with the smallest expansion accumulator capacity, while curve 918 represents the component with the largest expansion accumulator capacity). Similar curves can be plotted for behavior based on compression accumulator capacity. A frequency of 12 Hz was chosen because the wheel hop natural frequency for most vehicles is approximately 12 Hz, although other frequencies are contemplated. As will be appreciated by those skilled in the art, such curves may also depend on various accumulator parameters, including the precharge pressure and gas volume of each accumulator.

[0047] As can be seen from Figure 16, the suspension system components represented by curves 910, 912, and 914 have stiffnesses that exceed the desired maximum stiffness at set pressure differentials that are less than the maximum set pressure differential. Therefore, the suspension system components represented by curves 910, 912, and 914 are likely to experience ride comfort degradation, at least under certain conditions. On the other hand, the observed stiffnesses of suspension system components represented by curves 916 and 918, which have expansion accumulator volumes greater than those of 910-914, do not exceed the desired maximum stiffness for any set pressure differentials that are less than the maximum set pressure differential. Therefore, the suspension system components represented by curves 916 and 918 are likely to experience ride comfort degradation, at least under normal driving conditions.

[0048] Accordingly, the inventors have recognized that suspension system components can be constructed such that, for any set pressure differential less than the maximum set pressure, the observed stiffness of the component in response to an external input having a first frequency preferably does not exceed a desired maximum stiffness. As a result of extensive experimentation and simulation, the inventors have recognized that it is preferable to construct suspension system components such that, when the pump is operated to generate a set pressure differential of 1000 psi, the observed stiffness of the suspension system component in response to an external input having a first frequency and a first peak-to-peak amplitude does not exceed a desired first maximum stiffness value. In certain embodiments, the first frequency can be 12 Hz because 12 Hz is approximately a wheel hop frequency typical of consumer vehicles, and the first peak-to-peak amplitude can be 5 mm because 5 mm approximately corresponds to a displacement typical of a road surface. In certain embodiments, the desired first maximum stiffness value can be 80 N / mm. The inventors have recognized that the given values ​​minimize the risk of excessive stiffness in the active suspension system components degrading ride comfort in a typical vehicle, even when the pump is actively operated to actively apply sufficient force to maintain a desired vehicle attitude during normal driving conditions. In various embodiments, the first frequency can be any value between 8 and 15 Hz, and the first peak-to-peak amplitude can be any value between 3 and 7 mm.

[0049] As noted above, in certain applications it may be desirable to reduce the capacity of the expansion and / or compression accumulators to values ​​below those given by the above equations (e.g., to address packaging constraints), while still maintaining within the desired observed stiffness ranges noted above. The inventors have recognized that such suspension system components can be constructed using additional compliant mechanisms.

[0050] For example, FIG. 10 illustrates a suspension system component having a compliant mechanism 1001 including a housing, a piston 1005, a first spring 1003 attached to a first surface of the piston 1005, and a second spring 1007 attached to a second surface of the piston 1005. Although not shown, the suspension system component of FIG. 10 may also include a bypass flow path including one or more bypass valves. In the illustrated embodiment, the first surface of the piston 1005 is in contact with a fluid having a pressure equal to the pressure of the fluid in the compression chamber 111, while the second surface of the piston 1005 is in contact with a fluid having a pressure equal to the pressure of the fluid in the expansion chamber 109. If the area of ​​the first surface in contact with the fluid is equal to the area of ​​the second surface in contact with the fluid, a force acting on the piston 1005 exists that is proportional to the set pressure differential. This force can expand the first spring 1003 and compress the second spring 1007. In certain embodiments, the area of ​​the first surface in contact with the fluid and the area of ​​the second surface in contact with the fluid are substantially equal, while in other embodiments, these areas can be substantially different.

[0051] In certain embodiments, the first spring 1003 and the second spring 1007 have a substantially constant spring constant over at least some range of displacement, as is known for standard coil springs. In these embodiments, the stiffness of the compliant mechanism 1001 is constant regardless of the effective differential pressure. The stiffness of the compliant mechanism 1001 can be maintained constant until the set differential pressure provides sufficient piston travel to "bottom out" at least one of the springs 1001, 1007, at which point the springs become effectively infinitely stiff. This behavior is illustrated in FIG. 11, which shows the stiffness of the compliant mechanism on the y-axis and the effective differential pressure on the x-axis. As can be seen from FIG. 11, the stiffness is substantially constant over a first range 1101 of set differential pressures until, for a given effective differential pressure 1107, at least one of the first spring 1003 and the second spring 1007 bottoms out, thereby causing the stiffness of the compliant mechanism to increase at a steeper slope 1105. To avoid spring bottoming, the length of each of the first spring 1003 and the second spring 1007 should be determined taking into account the expected maximum available pressure differential in each direction, the spring constant of each of the first spring 1003 and the second spring 1007, the area of ​​the first face of the piston 1005 in contact with the fluid, and the area of ​​the second face of the piston 1005 in contact with the fluid. Alternatively or additionally, the first spring and / or the second spring may include one or more non-linear sections designed to prevent abrupt changes in stiffness.

[0052] Continuing to refer to FIG. 11, compliance on the expansion chamber 109 side of the pump can be provided by both the expansion accumulator 352 and the compliant mechanism 1001. Typically, when two compliant elements are present in a circuit, the compliance of the overall system is determined by the compliance of the most compliant (i.e., least stiff) element. In FIG. 17, the stiffness of the expansion accumulator 352 (and therefore its compliance) varies with the effective differential pressure. On the other hand, the stiffness of the compliant mechanism 1001 is independent of the effective pressure, at least over a first range of setpoint differential pressures 1101.

[0053] The resulting observed combined stiffness of a suspension system component incorporating both the expansion accumulator 352 and the compliant mechanism 1001 is shown in Figure 12. Over an initial range 1201 of available differential pressure, the expansion accumulator 352 is softer than the spring mechanism 1001, and therefore the observed stiffness of the suspension system component is determined by the stiffness of the expansion accumulator 352. As the set differential pressure increases, the stiffness of the expansion accumulator 352, and therefore the observed overall stiffness of the suspension system component, increases. Eventually, the expansion accumulator 352 becomes stiffer than the spring mechanism 1001, at which point the observed stiffness of the suspension system component is determined by the stiffness of the compliant mechanism 1001, at least until one of the springs 1003, 1007 of the compliant mechanism 1001 bottoms out. When a compliant mechanism 1001 is used, the expansion accumulator 352 no longer needs to remain less than the desired maximum stiffness over the entire set pressure range (up to the maximum anticipated set pressure), thus allowing for the use of a smaller expansion accumulator. For set pressure differentials in the opposite direction, the same theory applies to compression accumulators. In certain embodiments, the compliant mechanism 1001 can be constructed such that (a) the suspension system components respond to an external input having a frequency of 12 Hz and a peak-to-peak amplitude of 5 mm and have an observed stiffness less than the desired maximum stiffness, and (b) the compliant mechanism 1001 has sufficient length to not bottom out even at the maximum set differential pressure. In certain embodiments, the compliant mechanism 1001 can be constructed such that when the pump is driven to generate a set pressure differential of 1000 psi, the suspension system components respond to an external input having a frequency of 12 Hz and a peak-to-peak amplitude of 5 mm and have an observed stiffness less than the desired maximum stiffness. In certain embodiments, the desired maximum stiffness is 80 N / mm. In certain embodiments, the observed stiffness is within resp.

[0054] An exemplary method for determining the observed stiffness of an active suspension system component (e.g., an active suspension system component according to any of the embodiments disclosed herein) in response to a vibration external input having a particular frequency and amplitude is disclosed below. In the exemplary method, a damper is held approximately mid-stroke in a shock dynamometer (often referred to in the art as a dyno). To evaluate the observed stiffness of the suspension system component during active operation of the pump, the pump of the suspension system component can be optionally driven (e.g., a motor can apply torque to the pump) to generate a set differential pressure across the pump (e.g., a set differential pressure of 1000 psi). This set differential pressure can be in either direction. Because the damper is held mid-stroke (i.e., cannot contract or expand), active operation of the pump can generate and continuously maintain the set differential pressure. Once stationary conditions are achieved with the pump actively driven to maintain the set differential pressure, the dynamometer can be operated to apply a programmed external input to the tip of the piston rod. This programmed external input may include an oscillating series of displacements (e.g., applied to one end of the piston rod and transmitted to the piston) that varies according to a first frequency (e.g., the wheel hop frequency of the vehicle, e.g., 12 Hz). For example, a dynamometer may command the housing of the hydraulic cylinder to continuously and repeatedly move the piston rod (and therefore the piston) in an oscillatory manner by a specific amount. In certain embodiments, the peak-to-peak amplitude of the external input may be 5 mm. In certain cases, the external input may consist solely of an oscillating series of displacements. In other cases, it may be beneficial to superimpose the oscillating series of displacements with a DC offset force, so that the piston is constantly moving and static friction effects can be neglected.During application of an external input, the (directly related) hydraulic damper length and piston position, and / or the reaction force applied by the hydraulic system to the piston rod in response to the external input, can be observed in real time and used together with the known characteristics of the external input to determine the observed stiffness. As explained herein, this stiffness depends on the set differential pressure and / or the input frequency, so when referring to the observed stiffness, it may be necessary to specify the corresponding set differential pressure and input frequency.

[0055] In certain embodiments, a vehicle may include multiple suspension system components, such as those disclosed herein. For example, one suspension system component may be located on each wheel of the vehicle. Thus, a four-wheel vehicle may have four different suspension system components. In certain embodiments, each suspension system component may be hydraulically isolated from each other suspension system component of the vehicle, thereby preventing fluid communication between different suspension system components within a single vehicle. In other embodiments, there may be fluid communication between different suspension system components, for example, to provide enhanced anti-roll control.

[0056] As used herein, "fluid" refers to liquids and gases. As used herein, a first component is said to be in "fluid communication" with a second component if, during at least one operating state, a flow path exists through which the fluid can flow from the first component to the second component. As used herein, a first component is said to be in "selective fluid communication" with a second component if, during at least a first operating state, a flow path exists through which the fluid can flow from the first component to the second component and, during at least a second operating state, a flow path exists through which the fluid is prevented from flowing from the first component to the second component. As will be appreciated by those skilled in the art, selective fluid communication can be achieved, for example, using at least one valve that, in a first operating state (e.g., when the valve is in an "open" position), effectively allows fluid flow through the valve, but, in a second operating state (e.g., when the valve is in a "closed" position), effectively prevents fluid flow through the valve. As used herein, a first component is said to be in "variable fluid communication" with a second component if a flow path exists through which fluid can flow from the first component to the second component and the hydraulic resistance of the flow path can be controllably varied. As will be appreciated by those skilled in the art, variable fluid communication can be achieved using, for example, a variable orifice valve or other flow control valve.

[0057] As used herein, an electric motor refers to any device capable of converting electrical energy into mechanical energy. A non-limiting example of an electric motor is a brushless DC motor (BLDC motor). An electric motor can operate as an electric generator in certain operating modes and as an electric motor in other operating modes. A hydraulic pump refers to any device capable of converting mechanical energy into hydraulic energy (e.g., a pressure difference between two different chambers). A hydraulic pump can operate as a hydraulic motor in certain operating modes and as a hydraulic pump in other operating modes.

[0058] Recessed Hydraulic Accumulator Assembly Hydraulic systems are used in a variety of applications, including, for example, within vehicles, particularly within vehicle suspension systems. A particular application may provide only a limited amount of space within which a given hydraulic system must physically fit. Therefore, packaging size and space constraints can be a significant obstacle to incorporating a hydraulic system into a particular application, such as, for example, an automotive application. This obstacle can be particularly pronounced for hydraulic systems that use multiple accumulators, because each accumulator in the hydraulic system occupies some space and contributes to the total space of the hydraulic system. Disclosed herein are embodiments of an embedded accumulator assembly that allow for increased flexibility in packaging multiple accumulators into hydraulic systems and / or reduced packaging size for hydraulic systems that use multiple accumulators.

[0059] In certain hydraulic systems, it may be desirable to use multiple accumulators, for example, to reduce hydraulic noise present in various portions of the hydraulic system and / or to provide flexibility to various portions of the hydraulic system. However, in some applications using hydraulic systems, there may be severe limitations on the amount of space available for packaging the hydraulic system. In automotive applications, for example, the space available for packaging a hydraulic system is typically constrained by the requirement that such a system fit within available space, such as under the hood of the vehicle, under the body, or in the wheel wells of the vehicle. Because the use of multiple accumulators can increase the packaging size of the hydraulic system, incorporating multiple accumulators may be prohibitively difficult in some applications (e.g., vehicle suspension systems) where physical packaging tolerances are severely constrained. Accordingly, the present inventors have recognized the benefits of incorporating multiple accumulators into a single housing to allow for relaxed packaging requirements while still allowing for multiple independently functioning accumulators. Additional benefits of such an “embedded accumulator” assembly may include limiting long hydraulic fluid paths and an overall simplification of the hydraulic system.

[0060] In one aspect, an embedded accumulator assembly is disclosed. The embedded accumulator assembly can include a first accumulator including a housing at least partially defining an interior space. The interior space can be divided into a first chamber and a second chamber by a mechanical member, such as a piston (e.g., a floating piston or a spring-loaded piston), an elastic diaphragm, or a closure bladder. The mechanical member can include a sealing element (e.g., an O-ring) and can be sealingly coupled to a wall of the accumulator housing such that any fluid in the first chamber is isolated from any fluid in the second chamber. In certain embodiments, the first chamber contains a compressible fluid (e.g., a gas) and can be referred to as a gas-filled chamber. The second chamber can be at least partially filled with an incompressible hydraulic fluid and can be referred to as a hydraulic fluid chamber. In certain embodiments, the first chamber can alternatively or additionally include a spring element that mechanically attaches the piston to the housing.

[0061] In certain embodiments, the secondary accumulator can be located within the interior space of the primary accumulator (e.g., within the first or second chamber of the primary accumulator). This secondary accumulator can function independently within the environment of the primary accumulator. By locating the secondary accumulator within the interior space of the primary accumulator, the total packaging space can be reduced while maintaining maximum system functionality. In certain applications, even if the total packaging size of a given hydraulic system is not reduced by utilizing an embedded accumulator, such an embedded accumulator can nevertheless provide additional flexibility for fitting the hydraulic system into constrained spaces.

[0062] In some embodiments, an array of mechanical offsets is fixedly attached to the top and / or bottom faces of the floating piston, which can help prevent hydraulic noise induced viscous flow effects between the piston face and other flat surfaces when the piston is fully extended in one direction.

[0063] FIG. 13 illustrates an embodiment of an accumulator 1, particularly a floating-piston-type gas-filled accumulator. The accumulator 1 may include a cylindrical accumulator housing 2 that at least partially defines an interior space. In the illustrated embodiment, a floating piston 3 is slidably disposed within the accumulator housing 2, sealingly coupled to the inner surface of the wall of the accumulator housing 2 via a sealing element 4 (e.g., an O-ring). As illustrated, the piston 3 divides the accumulator's interior space into a gas-filled chamber 5 and a hydraulic fluid chamber 6. The gas-filled chamber 5 contains a gas or other compressible fluid, while the hydraulic fluid chamber is at least partially filled with an incompressible hydraulic fluid. The gas may be a pressurized inert gas that provides a restoring force to the piston 4 in response to movement of the piston 4 relative to the accumulator housing 2. The accumulator 1 may further include one or more ports configured to allow hydraulic fluid to enter or leave the hydraulic fluid chamber and other portions of the hydraulic system. In certain embodiments, the accumulator 1 includes only a single port, while in other embodiments, the accumulator 1 may include at least a first port 16 and a second port 17 .

[0064] Figure 14 is an illustration of an embodiment of a recessed accumulator assembly. The exemplary recessed accumulator assembly includes a first accumulator 1 substantially similar to that described in connection with Figure 13. The first accumulator may include a first accumulator housing 2 at least partially defining a first interior space, a piston sealingly coupled (e.g., via an O-ring) to an inner surface of a wall of the first accumulator housing, and a piston 3 dividing the first interior space into a first gas-filled chamber 5 and a first hydraulic fluid chamber 6.

[0065] The illustrated embedded accumulator assembly further includes a second accumulator 9. At least a portion of the second accumulator 9 may be disposed within the first interior space of the first accumulator 1 (e.g., within the first hydraulic fluid chamber 6). The second accumulator may include a second accumulator housing 10 that at least partially defines the second interior space, and a second piston 12 sealingly coupled to an inner surface of a wall of the second accumulator housing 10 (e.g., via a sealing element 13 (e.g., an O-ring) included on the second piston). The second piston 12 may divide the interior space of the second accumulator 9 into a second gas-filled chamber 14 and a second hydraulic fluid chamber 15. The second accumulator 9 may further include one or more ports 16, 17 that allow hydraulic fluid to enter and exit the second hydraulic fluid chamber. As will be appreciated by those skilled in the art, the secondary accumulator 9 functions via a mechanism substantially similar to the primary accumulator 1, except that the size of the secondary accumulator allows the secondary accumulator 9 to fit partially or completely within the interior space (e.g., hydraulic fluid chamber 6) of the primary accumulator. The secondary accumulator 9 can operate independently within the environment of the primary accumulator 1. The disclosed physical arrangement of the two accumulators, in which the secondary accumulator is at least partially disposed within the interior space of the primary accumulator, can reduce overall packaging requirements while maintaining system functionality.

[0066] In certain embodiments, a gap 18 may exist between an outer portion of the second accumulator housing 10 and an inner portion of the first accumulator housing 2. In certain embodiments, the gap may be annular to accommodate the geometry of the respective housings. In certain embodiments, it may be desirable to tailor the gap 18 to obtain beneficial hydraulic fluid flow characteristics into and out of the first accumulator. For example, when sized appropriately, the gap 18 may function as a fluid restricting element and / or a fluid inertance element. By controlling the size of the gap 18, the hydraulic resistance and / or hydraulic inertance may be controlled for the entire hydraulic system.

[0067] In certain embodiments, the first accumulator housing and the second accumulator housing can share one or more walls or portions of a wall. Furthermore, in certain embodiments, the second accumulator can be at least partially disposed in another portion of the first interior space, such as, for example, the first gas-filled chamber, rather than, or in addition to, being at least partially disposed within the first hydraulic fluid chamber 6. Furthermore, while the recessed accumulator assembly of FIG. 14 depicts a two-accumulator system, the disclosure is not limited in this respect. Other embodiments can include multiple accumulators disposed within the first interior space of the first accumulator, or a third accumulator within the second interior space of the second accumulator. Furthermore, it will be apparent that the system can be readily modified so that the first accumulator and / or the second accumulator are spring-loaded accumulators rather than the illustrated gas-filled accumulators.

[0068] While the present teachings have been described in conjunction with various embodiments and examples, it is not intended that the present teachings be limited to such embodiments or examples. On the contrary, the present teachings encompass various alternatives, modifications, and equivalents, as will be apparent to those skilled in the art.

Claims

1. A hydraulic cylinder (103) that defines at least partially the first internal volume, A piston (107) is slidably inserted into the hydraulic cylinder (103), thereby dividing the internal volume into a compression chamber (111) and an expansion chamber (109), A piston rod (113) is attached to the piston (107) and extends outward from the hydraulic cylinder (103), A hydraulic pump (201) includes a first port that is in fluid communication with the compression chamber (111) and a second port that is in fluid communication with the expansion chamber (109), A compression accumulator (350) configured to be in direct fluid communication with the compression chamber (111) and to exchange fluid with the compression chamber (111), wherein the compression accumulator (350) defines a second internal volume that is larger than the first internal volume, An expansion accumulator (352) is configured to be in direct fluid communication with the expansion chamber (109) and to exchange fluid with the expansion chamber (109), wherein the expansion accumulator (352) defines a third internal volume that is larger than the first internal volume, A suspension system component including, A suspension system component in which the accumulator parameters (including at least one of the accumulator internal volume, precharge pressure, and gas volume) of the compression accumulator (350) and the expansion accumulator (352) are set so that the observed stiffness of the suspension system component in response to an external input having a frequency of 12 Hz and an inter-peak amplitude of 5 mm does not exceed 80 N / mm when the hydraulic pump (201) generates a first set differential pressure of at least 68.95 bar.

2. The suspension system component according to claim 1, wherein the first set differential pressure has a value of 68.95 bar.

3. The suspension system component according to claim 1 or 2, wherein the observed stiffness is 5 N / mm, 10 N / mm, or 25 N / mm or more.

4. The suspension system component according to any one of claims 1 to 3, wherein the observed stiffness is 80 N / mm, 70 N / mm, or 50 N / mm or less.

5. A vehicle including a suspension system which includes the suspension system components described in any one of claims 1 to 4.

6. A vehicle including a suspension system which includes a plurality of suspension system components according to any one of claims 1 to 4.

7. The vehicle according to claim 6, further comprising a sprung mass and an unsprung mass, wherein each of the plurality of suspension system components is positioned between the unsprung mass and the sprung mass of the vehicle.

8. The vehicle according to claim 6 or 7, wherein each of the plurality of suspension system components is fluidly separated.

9. A method for sizing at least one accumulator of an active suspension system component, wherein the active suspension system component includes a hydraulic cylinder, a pump, a compression accumulator, and an expansion accumulator. The aforementioned method, The process involves instructing the pump to generate a pressure difference of at least 68.95 bar, The process involves applying an external input having a frequency of 12 Hz and a peak-to-peak amplitude of 5 mm to the hydraulic cylinder. A step of selecting the volume of at least one of the compression accumulator and the expansion accumulator such that the observed stiffness does not exceed 80 N / mm when the pump is commanded to generate the pressure difference and the external input is applied, A method that includes this.

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