JP2017508097A - Axial piston device - Google Patents

Abstract

The axial piston device can operate as a pump and includes a self-aligning rotary valve. The piston device includes a stationary housing that surrounds the shaft and the rotary valve. The rotary valve and the shaft are coupled to each other. As it rotates, the rotary valve self-aligns as a result of eliminating moments and forces in the pump. The pump of the present invention is a piston device. The valve is in the valve hole, which is part of the manifold. The shaft is in the manifold, and the shaft is attached at its distal end to the planar surface of the rotary valve. The shaft has a first rotation axis, and the rotary valve has a second rotation axis. During pump operation, the first axis is often offset from the second axis. The pump operates via a swash plate with a reciprocating piston, while the housing remains stationary.

Description

  No. 61/937166 filed on Feb. 7, 2014 and U.S. Provisional Application No. 62/093146 filed on Dec. 17, 2014, the contents of which are hereby incorporated by reference. Claim its benefit in relation to (incorporated in the description).

  The present disclosure relates to fluid forces, particularly fluid displacement through an axial piston device.

  Axial piston technology is typically embodied in non-rotating cylinder blocks and rotating cylinder block devices (eg, rotating cylinder block hydraulic pumps / motors, commonly referred to as “rotating groups”). A conventional non-rotating cylinder block pump is shown in FIG. 1, and a rotating cylinder block pump is shown in FIG.

  In the non-rotating cylinder block, there is a problem with the force applied to the pump shaft and the swash plate. These forces must be supported by bearings. Unfortunately, the forces acting on the pump shaft and swashplate are not balanced. If not checked, force imbalances can result in rapid pump wear. Since large bearings are required to handle the unbalanced load of the non-rotating cylinder block, the pump must be very large to accommodate the large bearings.

  These problems do not exist in the rotating cylinder block because the cylinder block absorbs most of the unbalanced forces in the non-rotating cylinder block. Thus, the rotating cylinder device can be much smaller than the non-rotating cylinder block design. The design of the rotating cylinder block has an inherent drawback in that it has a large rotational inertial mass (moment of inertia). This manifests itself as an increase in power loss, especially when the rotational speed changes. Devices constructed with this technique require additional components to accommodate the rotating cylinder block. This requires a device that is smaller than the previous device but still larger and heavier than desired. These properties are detrimental to packaging and efficiency.

  The pump of the present invention draws fluid through a non-rotating cylinder block and, in some embodiments, passes the fluid through a rotating valve. The rotary valve is always in fluid communication with the inlet and outlet cavities and the plurality of piston holes in the cylinder block.

  Pumps are generally characterized by allowing rotation of the swash plate while balancing the force on the swash plate and the force within the pump. Distributing fluid throughout the pump balances the unbalanced forces in the non-rotating cylinder block pump without the need for excessively large bearings. Embodiments of the pump include a two-sided swash plate to allow for balance of the total swash plate force. The size of this pump is no longer limited by the presence of rotating element bearings in operation. The pump can handle a wide range of operating conditions and has great advantages in terms of size reduction when used in high pressure and / or high displacement pumps. Although the design and function of prior art pumps has been limited by the need for very large bearings, this is not a problem in the design of the present invention.

  An exemplary embodiment of the hydraulic pump of the present invention is a pump having a plurality of passages in a shaft. The passage eliminates the need for a rotary valve, since the rotary valve and shaft functions are performed by a single component. One embodiment of the hydraulic pump includes a swash plate having a first surface coupled to a first set of slipper shoes and an automatic rotatable within at least one stationary cylinder block about its own axis. A aligning shaft. Self-aligning shafts have a plurality of passages for receiving, guiding and discharging fluid. The self-aligning shaft can be rotated concentrically and eccentrically during operation of the pump. Fluid pressure on the shaft passage balances forces and / or moments on the shaft. This pressure helps align the shaft.

  The swash plate is connected to the shaft and is configured to tilt around an axis orthogonal to the axis of the shaft. The inclination of the swash plate is obtained by hydraulic means, mechanical dogbone assemblies, electrical means, gearing, and the like.

  Another embodiment relates to a pump having a self-balancing shaft. The swash plate has first and second substantially planar surfaces that face each other. First and second plurality of slipper assemblies are coupled to the first and second planar surfaces. The fluid pressure on the shaft passage balances forces in all directions. The housing houses the shaft, swash plate and slipper assembly. The housing has at least one inlet for receiving fluid into the pump and at least one outlet for discharging fluid from the pump.

  The shaft of the hydraulic pump is self-aligning. A typical hydraulic pump includes at least one stationary one-piece cylinder block and a shaft that is rotatable about its own axis within the stationary cylinder block. This embodiment of the hydraulic pump also includes a first and a second plurality of pistons. Each of the plurality of pistons is accommodated inside the cylinder block, and is arranged in an arc shape coaxially with the shaft. The plurality of pistons are on both sides of the swash plate. At least one (or first) cylinder block surrounds the first plurality of pistons and a second stationary cylinder block surrounds the second plurality of pistons.

  The fluid flow through the entire pump caused by the swashplate driving the reciprocation of the piston generates forces and / or moments on the shaft. A plurality of passages are formed in the shaft so that this force and / or moment on the shaft is balanced.

  The pump can be configured such that the hydraulic pump shaft has a valve integrated therein and that the passage along the shaft performs more than the minimum function of reliably aligning the shaft. Passages are also provided for receiving, directing and discharging fluid flow at a plurality of advantageous locations along the shaft. In other words, in addition to the minimum number of passages required to align the shaft, passages can be added to the shaft based on the required fluid flow rate of the pump. When the embodiment is configured in this way, it includes at least one stationary cylinder block and a valve enclosed within the cylinder block. The valve includes a shaft that passes through a stationary cylinder block. The shaft has a plurality of passages along the length of the shaft.

  One embodiment of the hydraulic pump has a two-sided swash plate. The swash plate has a first surface coupled to the first set of slipper shoes and a second surface coupled to the second set of slipper shoes. The first surface of the swash plate is substantially parallel to the second surface of the swash plate. The component coupled to the first surface of the swash plate is substantially symmetrical with respect to the component coupled to the second surface of the swash plate with the swash plate interposed therebetween.

  The housing surrounds the swash plate. The swash plate is coupled to a plurality of slipper assemblies via fluid bearings. The slipper assembly is on both sides of the swash plate. The shaft fixes the swash plate at a fixed distance with respect to the end of the shaft. The swash plate can be fixed in place by a pin. Regardless of the means used to secure the swashplate, the swashplate is configured to be adjustable about an axis that is oblique to or perpendicular to the axis of the shaft. Or, in the case of fixed volumetric displacement, the swashplate can be fixed at a certain angle. That is, the swash plate need not necessarily be adjustable.

  Another embodiment of the pump has a two-sided swashplate and multiple pumps are housed in a single housing. This embodiment has a swash plate having a first swash plate surface. The first pump has a first plurality of components including a first swash plate surface. The first swash plate surface is coupled to the first plurality of slipper shoes. The slipper shoe is coupled to a plurality of pistons. Each piston is accommodated in a respective cylinder hole. The second pump includes a second plurality of components that are substantially the same as the components of the first pump. The components of the second pump are substantially symmetric with respect to the components of the first pump. The central plane of the swash plate is the boundary between the pumps.

  A single housing surrounds the first and second pumps and the cylinder bore is integral with the housing. The pump includes integral means for balancing internal forces and / or moments in the first and second pumps. It can be envisaged that at least one additional pump is included in a single housing and is configured substantially similar to the first and / or substitution pump. In this case, the shaft of the additional pump is arranged in the housing parallel to the first and second pumps.

  In another embodiment, the rotary valve is heterogeneously coupled to the shaft so that only thrust and torque are transmitted between the rotary valve and the shaft. The shaft is also coupled to the swash plate. The inclination angle, that is, the angle of the sliding boundary surface between the swash plate and the slipper with respect to the plane orthogonal to the rotation axis of the shaft is set to a predetermined inclination angle. That is, it is fixedly coupled to the shaft or variably controlled by an inclination control mechanism. The rotating swash plate reciprocates the piston through the slipper assembly. In one half of the swashplate, the piston is withdrawn from the respective cylinder hole, and in the other half, the piston extends into the respective cylinder hole and is pushed in.

  The rotary valve is generally cylindrical with two separate fluid passages, namely suction and discharge passages, ie inlet and outlet passages, respectively. Both passages communicate with one or more piston holes on a curved surface. The curved surfaces of the two passages face each other and each occupy less than half of the rotary valve circumference. The rotary valve is coupled to the shaft such that the pump inlet passage leads to the piston bore of the piston retracting from the piston bore, and the pump discharge passage leads to the piston bore of the piston extending into the piston bore. Thus, as the piston is pulled from the hole, the rotary valve allows fluid from the suction cavity of the valve to transfer into the increasing volume of the respective piston hole. In the piston hole on the other (diametrical) side of the pump shaft, fluid is forced from the piston hole through the discharge passage and into the valve discharge cavity.

  Variations in the design of the rotary valve are possible. The disclosed pump embodiment includes a housing with an integral manifold surrounding the rotary valve. The rotary valve includes a shaft, a passage portion, and a seal portion. The seal portion has a semicircular seal ridge. The rotary valve has first and second axial seals at the first and second planar ends of the rotary valve, and a high pressure discharge on the side of the rotary valve opposite the semicircular seal ridge and the first end of the rotary valve. The part has an inlet part. The rotary valve is generally cylindrical and the semicircular seal ridge surrounds less than 360 ° of the rotary valve. The rotary valve has a manifold engaging portion (that is, a seal ridge) and a recessed portion.

  A working gap exists between the high pressure discharge and the manifold. The width of the working gap depends on the force applied in the direction of the semicircular seal ridge at the high pressure discharge. The semi-circular seal ridge maintains a substantially unchanged working distance from the manifold when activated. The width of the working gap also depends on the speed of the shaft and the circumference of the rotary valve.

  A further embodiment of the pump includes a valve hole surrounding the rotary valve. The rotary valve has at least one axial face seal and at least one radial face seal. An operating seal gap exists between the radial face seal and the valve hole. The working seal gap maintains a generally unchanged thickness during valve actuation. The discharge gap is between the rotary valve and the hole. The rotary valve is energized by a force applied in the direction of the semicircular seal ridge, which determines the width of the working seal gap. The rotary valve includes a shaft. The radial face seal has a semi-circular sealing element and is generally coaxial with the shaft, the shaft having an inlet passage and an outlet passage. The radial face seal is free to move relative to the shaft so that it is not always coaxial with the shaft.

  Yet another embodiment of the pump includes a housing having a valve hole and a manifold. The shaft is in the manifold and the rotary valve is in the valve hole. The shaft is attached to the planar surface of the rotary valve at the distal end. The shaft has first and second axes that are offset from each other. In operation, the second axis coincides with the valve hole axis and the first axis is offset from the manifold centerline.

  The rotary valve has a high pressure outlet and a low pressure inlet. The shaft can be offset in the direction opposite to the high pressure outlet of the rotary valve and can also be offset in the direction closer to the high pressure outlet of the rotary valve.

  The shaft contacts the manifold via the manifold contact surface of the shaft. The bearing is disposed in the manifold and abuts the shaft at the manifold contact surface.

  A plurality of cylinders are arranged in the manifold in parallel and circularly around the valve hole and coaxial with the valve hole. The rotary valve is generally coaxial with the valve hole and the shaft is not coaxial with the valve hole. The shaft is integral with the rotary valve.

  The rotary valve balances the high pressure load. The rotary valve can be configured to eliminate unbalanced forces (which causes the valve to move away from its axis of rotation). An unbalanced force on the valve breaks the lubricant film barrier, thereby creating a metal-metal contact between the rotary valve and the valve hole. This can cause the valve to adhere to the valve hole and interfere with the rotating member.

  The rotary valve transfers or directs high pressure fluid to a thrust cavity between the valve and a manifold thrust shelf to balance axial loads on the shaft. In one embodiment, the area of the valve / fluid thrust cavity interface is equal to about half of the sum of all piston / fluid interfaces. Most of the axial load on the shaft is contained within the cylinder block. Roller element thrust bearings are suitable for controlling a portion of the total axial load between the shaft and the outer housing.

  In this embodiment, the low pressure fluid passes through the inlet port and enters the pump to the valve inlet cavity. This fluid is delivered to the plurality of piston holes by a valve inlet passage. Each piston flowing through the valve inlet passage is pushed out of the piston hole. As the shaft-valve continues to rotate, the piston is pushed into the piston bore, thereby compressing the fluid. At this time, the piston hole starts to open toward the valve discharge passage. The valve discharge passage allows high pressure fluid from the piston hole to be sent to the discharge cavity, and the fluid exits the pump from the discharge cavity through the discharge port. The force required to move the piston is given by the input torque to the shaft-valve. This torque is transmitted to the swash plate. The swash plate is pushed into the slipper in the high pressure side half of the swash plate by the inclination of the swash plate relative to the shaft-valve (known as the tilt angle). The slipper assembly is then pushed into the piston in fluid communication with the liquid inside the piston bore. This process produces a set of different forces / moments. The axial load is supported by the swash plate and transmitted to the valve shaft. Equal and opposite axial loads are supported by the valve shaft by the thrust cavity, thereby resulting in a zero net axial load. The moment is supported by the swashplate and transmitted to the valve-shaft. Equal and opposite moments are supported by the valve-shaft by the differential pressure between the valve discharge passage and the valve inlet passage. The final force / moment is the moment about the axis of rotation, to which the input torque required to drive the shaft-valve reacts. In this way, the shaft-valve and all other rotating members are balanced. The balance of the rotating member is independent of the swash plate control and / or slipper hold mechanism.

  In prior art pumps, the displacement of the high pressure fluid creates a large thrust or axial load on the swashplate and manifold. Thrust or axial load is transmitted to the pump outer housing. Also, in prior art pumps, the trust load creates a large separation force between the outer housing components, thereby increasing the required clamp load. When this thrust load is distributed to multiple components, a one-piece cylinder block that includes all axial loads in the body is used to reduce the required amount of clamp load, thereby reducing the size and weight of the pump. Unlike the invention, it is disadvantageous because it increases the overall size and weight of the pump.

  The pump incorporates a non-rotating cylinder block. The reciprocation is obtained by attaching a swash plate to the rotating shaft. The swash plate rotates around a single axis at a predetermined angle with respect to a single axis to reciprocate the piston within the pump. The generally cylindrical rotary valve has a suction or intake area in communication with an external source of low pressure fluid on a planar surface. The intake circulates through the valve with an inlet passage in the curved surface of the rotary valve (configured to alternately communicate with one or more cylinder holes). On the opposite curved surface of the rotary valve is the rotary output or discharge area of the rotary valve. The output area alternately communicates with one or more cylinder holes and becomes a conduit for outputting high pressure fluid from the pump.

  This pump draws fluid through the centerline of the rotary valve, and the fluid travels a more direct and consistent path, reducing the possibility of increased fluid velocity, and therefore between the inlet passage and the piston hole. Reduce pressure loss. However, this valve leads to a cylinder block conduit that consistently leads to the piston bore.

  The rotary valve balances the high pressure load. This eliminates the force that would otherwise remove the valve from the center. The rotary valve can maintain a lubricant film barrier and prevent metal-metal contact between the rotary valve and the cylinder block.

  A conventional slipper shoe is attached directly to the spherical head of each piston and maintains effective sliding contact with the flat surface portion of the swash plate by a restraining plate.

  Previous iterations of non-rotating cylinder block axial pumps required large bearings. As a result, the pump is large and heavy. The embodiments disclosed herein may reduce size and weight by as much as a third without losing capacity or efficiency, and various pump modifications may be obtained. The pump can have a larger diameter piston with a smaller stroke. The piston can be made shorter without sacrificing capacity, thereby enabling shorter pumps. Unless the piston load is balanced with the working fluid, the load must be balanced with a mechanical bearing (ie, a thrust bearing). In prior art pumps, this type of bearing with a short and wide mechanical piston must be very large. However, since this pump employs a fluid bearing, a short and wide piston can be used without increasing the size of the pump.

  The swash plate is configured to adjust the stroke length of the plurality of pistons. The means for adjusting the tilt angle of the swash plate can be mechanical or hydraulic, or a combination thereof. In the case of hydraulic pressure, it may include a plurality of liquid-filled volumes configured to adjust the tilt angle of the swashplate in response to variations in hydraulic fluid within the volume. The rotary valve of the pump can also include a moment nulling means. The moment zeroing means has an equilibrated high pressure passage that is arranged to receive high pressure fluid during rotation.

  Each of the plurality of pistons has a working fluid end. The effective (net) surface area of the planar end of the rotary valve exposed to the high pressure fluid is equal to (or approximately equal to) the total surface area of the high pressure ends of each of the plurality of pistons exposed to the high pressure fluid. A plurality of slipper assemblies connect the swash plate to each piston. A hydrodynamic bearing is disposed between each slipper assembly and the swash plate. The fluid bearing is fluidly coupled to the high pressure end of each piston.

  The shaft and rotary valve are designed to move freely with at least 4 degrees of freedom. This hydraulic pump has a stationary cylinder block and a valve in the stationary cylinder block. The valve is connected to the shaft and is substantially coaxial with the shaft within the stationary cylinder block. The valve is rotatable about the axis of the shaft, and the valve is configured to be offset with respect to the axis of the shaft with 4 degrees of freedom.

  The connection between the valve and the shaft is configured to limit the transmission of force from the shaft to the valve to a torque around the shaft rotation axis and a thrust load along the rotation axis. The valve is balanced within the cylinder block by a controlled leakage rate as a function of force and / or moment on the valve.

  Another embodiment of the hydraulic pump is a pump designed to have a drive key. This embodiment has a shaft-valve assembly. The valve is connected to the shaft through the interaction between the drive key and the keyway. The connection between the valve and the shaft is maintained by any of several types of fasteners. The interaction is configured to give the valve at least one degree of freedom with respect to the shaft. The degree of freedom (which may be plural) is a linear direction configured to be orthogonal to the rotational axis of the shaft. The swash plate is connected to the shaft. The swash plate rotates with the shaft and is configured to tilt about an axis perpendicular to the axis of the shaft via means for tilting the swash plate.

  The rotary valve is an elongated cylindrical shape that preferably has a height greater than the diameter of the rotary valve. The high-pressure crossing passage is juxtaposed on the side surface of the rotary valve. The low pressure transition zone is on the side of the rotary valve, contacts the high pressure passage and extends a set distance from the high pressure passage.

  The shaft is integral with the planar end of the rotary valve. The fluid thrust bearing surface area of the discharge thrust cavity is equal to about half of the total area of the end faces of the plurality of pistons.

  The plurality of slipper assemblies are fluidly coupled to the swash plate. The slipper assemblies each include a slipper ball, a slipper neck, and a slipper shoe. The hold-down plate maintains a fluid coupling between the slipper assembly and the swash plate. The retainer plate includes a plurality of recesses each configured to engage a respective slipper assembly. The restraining plate is prevented from rotating around the shaft.

  Another embodiment is a separate rotary valve / shaft embodiment. The rotary valve of the separation type embodiment is a rotary barrel valve. One advantage of this embodiment is that, unlike disc-type valves, it can eliminate cantilevered force by using a valve that mimics a barrel. This is because the disc tends to cantilever force. This embodiment includes a shaft having a swash plate. The shaft and swash plate rotate within the housing. The plurality of pistons oscillate in response to the rotation of the shaft and the swash plate. The rotary valve is coaxial with the shaft and is fixedly attached to the shaft at the planar end of the rotary valve. The rotary valve is configured to rotate about a common axis with the shaft.

  An advantageous feature of this embodiment is that the inlet port is centered along the main axis of the pump. One advantage of this configuration is that the number of turns of the inlet passage in the rotary valve can be reduced. Therefore, the possibility of cavitation is reduced. This embodiment includes a shaft and a rotary valve in the pump housing. The rotary valve is coaxial with the shaft and is connected to the shaft. The passage is connected to the inlet port and has a stroke path with a turn or turn of 90 degrees or more. At least one high pressure passage is disposed in the rotary valve.

  The pump housing has an integrated manifold and a plurality of cylinders. The cylinder hole is fluidly connected to the rotary valve high pressure passage. The cylinder bore fluidly connects with the pump inlet and passage and the pump outlet alternately as the rotary valve rotates.

  This embodiment includes a manifold housing having a plurality of piston holes. The rotary valve is rotatably accommodated within the manifold housing and includes an inlet hole and an inlet passage that are fluidly connected in series with the cylinder hole. The inlet passage has an axis that passes through the rotary valve. The inlet passage has a maximum bend angle of about 90 degrees to avoid excessive fluid separation, thus increasing the speed of fluid passing through the passage.

  The swash plate dominates the flow rate at least partially through the inlet, which is also governed by the rotational speed of the shaft. The rotary valve includes a plurality of high-pressure crossing passages juxtaposed on the side of the rotary valve. The high pressure passage is fluidly connected to the high pressure intersection passage and extends a set distance from the high pressure intersection passage. The planar end of the rotary valve allows for the presence of an exhaust thrust cavity configured to balance the axial force at the pump.

  The connection between the rotary valve and the drive element (shaft) can be obtained by a rotary valve keyway and one or more C-clips. The advantage of this arrangement is that the planar thrust load is supported by the interface between the keyway, C-clip and shaft, which provides a means for transmitting torque from the drive element to the rotary valve. The connection between the keyway of the rotary valve and the key of the shaft transfers torque from the shaft to the rotary valve.

  The connection between the shaft and the rotary valve is locked by at least one preferably a pair of C-clips coaxial with the shaft. The clip engages a recess in the shaft and a corresponding recess in the rotary valve. The keyway can have at least one degree of freedom on the planar surface of the rotary valve.

  The rotary valve has components integrated into the surface of the rotary valve that are configured to balance the forces acting on the rotary valve. These components can also be configured to help balance the moment of the entire pump. At least the component is configured to balance rotational and axial forces in the pump.

  Separate shaft-valve pumps (and integrated shaft-valve pumps) can include pressurization and decompression notches in the surface of the rotary valve. The advantage is reduced pump noise. Pump noise arises particularly from rapid changes in fluid pressure.

  This embodiment of the pump includes a manifold having a main hole and a rotary valve in the manifold-housing. The rotary valve includes a high pressure outlet and a low pressure outer surface. The rotary valve has a rotation axis along the center line of the main hole. At least one pressurizing notch is fluidly connected to the inlet side of the high pressure outlet.

  The rotary valve includes an inlet at one of its planar surfaces. The inlet is centered around the axis of rotation of the rotary valve. The rotary valve also includes a cross passage on its curved side. The force acting on the rotary valve by the flow of high pressure fluid is balanced by the cross passage. At least one high pressure zone and at least one low pressure zone are provided by the crossing passage. The high pressure zone and the low pressure zone are configured around the rotary valve to eliminate the moment force on the rotary valve, and the placement of the zone is determined by experiment or by a mathematical formula in which all moments and forces on the valve are balanced. .

  The advantage of the one-piece cylinder block described above is that the forces acting on the seams in the multi-part housing of the prior art cylinder block are eliminated. The cylinder block includes a plurality of cylinder holes. The piston is slidably disposed in each cylinder hole. The cylinder block is an integral type, and the cylinder hole is integral with the cylinder block.

  The fluid supply flow path of the hydraulic slipper assembly passes through the entire length of the slipper assembly. Each of the plurality of pistons has a piston fluid supply channel that is fluidly coupled to the slipper assembly fluid supply channel and fluidly coupled to the piston hole.

  The disclosed content employs a hydrodynamic bearing that allows relative movement between the valve / shaft / swash plate of the pump and the cylinder block / manifold / housing. This configuration reduces the load on all bearings in the device, thereby reducing parasitic bearing loss during operation.

  The decrease in bearing load is a function of the balance between the fluid force and the structural response to the load. Since pressure and / or energy is transferred from fluid to mechanical force or work and converted to load, fluid force is generated in the pump when mechanical energy is transferred from the drive motor to the fluid. By design, the axial force of the piston caused by the reaction of the fluid pressure resisting the piston movement and the reactive force generated by the high pressure fluid acting on the planar surface of the valve / shaft (the planar surface has a vertical axial component) There is an equilibrium with the second force which is The net force (i.e., the difference in magnitude between the first force and the second force) varies continuously during operation of the device. Because of the time variation imbalance of these forces, the structure of the device needs to support net residual forces and moments.

  The advantage of the non-rotating cylinder block technology is a smaller mass and a smaller device diameter, which gives a smaller rotating mass and thus inertia and improves the overall efficiency when operating in unsteady conditions. Also, the cylinder block can be used as a device housing, so there is no need to add a separate component or housing as a structural member of the device.

  This pump is used to balance the internal forces and / or moments described above for rotating members (ie shafts, swashplates and valves, and mechanisms used to connect and / or control them). Inevitable leakage from fluids and working fluids (ie, fluids that are displaced at high pressures. Parasitic losses because this fluid would otherwise only leak between the various components in the pump and be discharged to the tank / container. Which would have shown 100%). This pumping device does not require additional parasitic porting of the working fluid to develop a load resistance bearing (i.e. to support a hydrostatic bearing) when the working fluid is displaced at high pressure. Minimize the use of a set (by design) controlled leak of hydraulic fluid to balance the internal forces on the mechanical components generated by This has the effect of improving the overall efficiency of the device. Thus, this pump does not require the use of active element bearings (ie roller element bearings). A typical axial piston device requires a large diameter, weight and cost bearing at a given load and rotational speed.

  The disclosed pump embodiment has the shape of a cylindrical valve with a curved seal / bearing surface. The prior art uses the valve plate as a planar bearing (ie, two flat sealing surfaces separated by a working fluid acting as a fluid film). The advantage of a cylindrical shape is that when operating in off-axis loading conditions (ie, the cylinder's eccentric position relative to the hole), the gap between the cylinder and its hole creates a variable gap size. This produces a working fluid wedge effect upon rotation. This phenomenon, known as the wedge effect (FIG. 3), results in an unbalanced pressure distribution within the load region, which has two functions. First, a reaction force equal to and applied to the applied load is applied. Second, the wedge effect draws new fluid into the load area as the cylinder (or shaft) rotates, thereby replenishing fluid lost due to the differential pressure between the load area and the rest of the fluid cavity.

  The disclosed pump combines the valve and journal bearing mechanism into one component. The axial leakage flow from the discharge side to the suction side (across the bearing lands) provides a heat transfer means, thereby eliminating the need to send preload fluid (lubricant) from the dedicated port to the center of the journal land area.

1 shows a conventional pump having a non-rotating cylinder block. 1 shows a conventional pump having a rotating cylinder block. The rotary valve which receives a wedge effect is shown. 1 shows a first embodiment of a self-equilibrating hydraulic pump. 2 shows a plurality of pistons of a hydraulic pump of the present invention arranged in an arc. 1 shows a first embodiment of a hydraulic pump having a rotating swash plate. 2 shows a second embodiment of a hydraulic pump. 3 shows a third embodiment of a hydraulic pump. It is sectional drawing of the hydraulic pump of FIG. FIG. 6 is another cross-sectional view of the hydraulic pump in FIG. 5. It is detail drawing of the rotary valve of the hydraulic pump of FIG. It is detail drawing of the rotary valve of the hydraulic pump of FIG. It is detail drawing of a swash plate. It is detail drawing of a slipper assembly. It is a figure of a two-piece type rotary valve. It is a rotary valve that shows significant wear. It is sectional drawing of the pump which shows the 1st version of a rotary valve. FIG. 17 is a first view of a seal portion of the rotary valve of FIG. 16. It is a 2nd figure of the seal part of FIG. 17A. FIG. 17B is a cross-sectional view of the seal portion of FIG. 17A showing a pressure drop across the surface of the seal portion. FIG. 17B is a first view of the seal portion of FIG. 17A attached to a rotary valve. FIG. 17B is a second view of the seal portion of FIG. 17A attached to the rotary valve. It is 1st sectional drawing of the whole pump which uses the rotary valve of FIG. FIG. 17 is a second cross-sectional view of the entire pump using the rotary valve of FIG. 16 with the rotary valve rotated 90 ° from the position of FIG. FIG. 17 is a top view of the rotary valve and the seal valve of FIG. 16 inside the valve hole. It is sectional drawing of the pump which shows the 2nd version of a rotary valve. FIG. 22 is a first view of a seal portion attached to the rotary valve of FIG. 21. FIG. 22 is a second view of a seal portion attached to the rotary valve of FIG. 21. FIG. 22 is a cross-sectional view of the seal portion of FIG. 21 showing a pressure drop across the surface of the seal portion. It is a top view of the rotary valve and seal part of FIG. 21 inside a valve hole. It is sectional drawing of the pump which shows the 3rd version of a rotary valve. It is sectional drawing of the rotary valve and shaft of the pump of FIG. 3 illustrates an axial face seal embodiment and a radial face seal embodiment. It is a fragmentary figure of a rotating surface seal. It is a perspective view of a shaft surface seal. 28 shows a dual face pump having the radial face seal embodiment and the rotating face seal embodiment of FIG. FIG. 3 is a cross-sectional view of a valve and radial face seal showing the pressure gradient in the radial face seal of the valve. It is sectional drawing which shows O-ring with respect to the channel | path in a valve.

  FIG. 4 shows the hydraulic pump 2. The pump 2 includes a housing 4. A shaft-valve (shaft and valve) 6 extends the entire length of the housing 4. The shaft-valve 6 allows the forces and / or moments in the pump 2 to be balanced. As a result of force and / or moment balancing, the need for mechanical bearings in the pump is reduced. Thus, the structure and function of the pump 2 is not limited by the function of the rolling element bearings during operation because the pump 2 does not require roller element bearings.

  This embodiment of the pump 2 has a plurality of passages 10 a and 10 b in the shaft-valve 6. The flow of fluid through the entire pump 2 and the rotation of the rotary swash plate 8 produce forces and / or moments on the shaft-valve 6. Fluid flow occurs as a result of swashplate rotation. The piston produces a fluid flow through reciprocation resulting from engagement with the rotating swash plate. Thus, the rotating swashplate (and fluid flow) creates a force on the pump. A plurality of passages 10a and 10b are configured in the shaft-valve 6 so that forces and / or moments on the shaft-valve 6 are balanced.

  A rotary valve separate from the shaft is advantageous in certain embodiments and is discussed in detail below. However, in this embodiment, the passages 10a and 10b are integral with the shaft-valve 6 and the functions of the rotary valve and shaft are performed by a single component or shaft-valve 6 so that separate rotations are possible. Does not require a valve. A conventional non-rotating cylinder block pump without a rotary valve as disclosed herein is shown in FIG.

  The swash plate 8 is connected to the shaft-valve 6 and is preferably configured to tilt about an axis 28 perpendicular to the axis of the shaft-valve 6. The swash plate 8 has a first surface 12 coupled to a first plurality of slipper assemblies (slipper assemblies). The contact between the slipper assembly 14 and the swash plate 8 is shown in FIG. The shaft-valve 6 is self-centering within the housing 4 about its own physical axis 16. That is, the axis of rotation of the shaft-valve 6 is aligned as much as possible with the axis of symmetry of the shaft-valve. The swash plate 8 is a two-sided swash plate. The swash plate 8 has a disk shape and is symmetrical with respect to a plane passing through the center of the disk. The first surface 12 is preferably a planar surface. The swash plate 8 has a second surface 20 which is likewise preferably a planar surface. The first surface 12 and the second surface 20 of the slope 8 face each other. The second plurality of slipper assemblies 50 are coupled to the second surface 20 of the swash plate 8.

  The use of wedge-shaped swash plates in double sided pumps has been dominant in the prior art. However, in this pump 2, the first surface 12 of the swash plate and the second surface 20 of the swash plate are as close to parallel as possible. The first surface 12 and the second surface 20 do not have to be absolutely parallel to each other. However, the closer the first surface 12 and the second surface 20 are to parallel, the closer to perfect balance of forces and / or moments on the shaft-valve 6.

  Each of the slipper assemblies 14 includes a piston 18. The housing 4 acts as a cylinder block for each pinton 18. In the pump 2, the housing 4 is non-rotating and each piston 18 remains stationary with respect to the circumference while the shaft-valve 6 and swash plate 8 rotate. This is in contrast to prior art pumps in which the cylinder block rotates.

  Referring to FIG. 6, the first plurality of slipper assemblies 14 are coupled to the first surface 12 of the swash plate 8, and the second plurality of slipper assemblies 50 are coupled to the second surface 20. There is a slipper shoe 24 at the end of each slipper assembly 14. The slipper shoe 24 can slide a predetermined distance 44 along the surface of the swash plate 8. The connection between the slipper shoe 24 and the two faces 12 and 20 of the swash plate 8 is obtained via a fluid bearing. Accordingly, the slipper shoe 24 slides freely along the surface of the swash plate 8.

  The pin 22 fixes the swash plate 8 at a fixed distance with respect to one end of the shaft-valve 6. Regardless of the means used to secure the swash plate 8, the swash plate 8 is configured to be adjustable about an axis 28 that is oblique or orthogonal to the axis of the shaft-valve 6. Increasing or decreasing the inclination angle of the swash plate 8 adjusts the stroke capacity of the hydraulic pump 2, that is, the displacement of the displacement per shaft rotation. The greater the inclination angle of the swash plate 8, the greater the stroke capacity per rotation of the pump 2. As the inclination angle of the swash plate 8 increases, the stroke of each of the pistons 18 of the slipper assemblies 14 and 50 increases. Accordingly, the piston hole 18 that accommodates the piston 18 can accommodate more fluid, thereby increasing the stroke capacity.

  To actuate the piston 18, the shaft-valve 6 rotates on its axis 16. The rotation of the shaft-valve 6 causes the swash plate 8 to rotate about the axis 16. As the swash plate 8 tilts, the interaction between the swash plate 8 and the slipper assemblies 14 and 50 causes the piston 18 to reciprocate along the individual axis of each piston 18. The interaction between each slipper assembly 14 and 50 and the swash plate 8 is enhanced by the presence of a hydrodynamic bearing between the swash plate 8 and the slipper assemblies 14 and 50. The hydrodynamic bearing is provided by a high pressure fluid supplied through the center of the piston and slipper.

  The swash plate 8 can be tilted around the axis 28 of the pin 22. In operation, means 32 for inclining (ie angular position) the swash plate 8 with respect to the axis of rotation 16 of the shaft-valve 6 is connected to the swash plate 8. The means 32 for inclining the swash plate 8 may be hydraulic means such as hydraulic jacks or cylinders, mechanical dog-bone assemblies responsive to mechanical inputs, electrical means, gearing or any of these Any combination of these may be used. The shaft-valve 6 can cause an eccentric rotation (and center rotation) that is not intended but sometimes unavoidable during operation of the pump 2. It is therefore advantageous to have a self-centering shaft-valve 6. As shown in FIG. 4, a plurality of passages 10a and 10b receive fluid through the entire pump into the shaft-valve 6 and direct fluid therethrough and drain fluid from the shaft-valve. Each of the plurality of pistons is arranged arcuately coaxially with the shaft-valve 6. The plurality of pistons 18 are grouped into two sets, which are at both ends of the housing 4 and engage the surfaces 12 and 20 of the swash plate 8, respectively.

  The housing 4 has at least one inlet passage 36 for receiving fluid in the pump and at least one outlet passage 38 for discharging fluid from the pump 2. The passages 10a and 10b at each end of the shaft are approximately 180 ° apart.

  As can be seen from FIG. 6, fluid enters the housing 4 in the inlet passage 36 and travels through the low pressure cavity 46. The fluid enters the plurality of piston holes 48 via the low pressure ports 40a and 40b and the low pressure passage 10b (FIG. 4). When the shaft-valve 6 rotates, the low pressure passage 10b is continuously aligned with the low pressure ports 40a and 40b. High pressure passage 10a is continuously aligned with high pressure ports 42a and 42b (FIG. 6). The high pressure ports 42a and 42b are connected to the outlet passage 38 via lines 52a and 52b.

  The force acting on the shaft-valve 6 (mainly caused by the moment generated by the high pressure fluid force on the piston) is balanced by the shape and installation position of the passages 10a and 10b. With further reference to FIG. 4, when the passage 10a engages the high pressure port 42a, the passage 10b that kinetically opposes the passage 10a, ie, balances the force on the pump, engages the low pressure port 40b. This is because the position of the slipper assembly 8 is generally opposite on the swash plate 8 (ie, when the swash plate 8 tilts and rotates, one slipper assembly slides out of its hole, thereby creating a vacuum. However, the opposing slipper assembly slides into the hole and helps create a high pressure in the fluid inside the hole), because the action of the slipper assembly is opposed to each other. The effect is that the force on the pump 2 caused by the flow of fluid in the pump 2 balances the force of the slipper assemblies 14 and 50 against the swash plate 8.

The surfaces 12 and 20 of the swash plate 8 are in contact with the slipper assembly 14 on the high pressure side half of the swash plate 8 because the swash plate 8 is inclined relative to the shaft-valve 6 (known as the swash angle). Apply pressure. Each slipper shoe 24 of each slipper assembly 14 and 50 receives force and applies pressure against each respective piston 18. The piston is in fluid communication with the fluid in the piston hole 48. As shown in FIG. 4, this process produces four different forces / moments (F _{1, u} , F _{1, l} , M ₁ , M ₂ ). The axial load (F ₁ ) is supported by the swash plate 8. The axial load has two equal and opposite components (F _1u and F _1l ). The back-to-back piston arrangement has a trade-off. A perfect balance of axial load can be obtained with zero indexing (ie, misalignment), but an indexed piston results in a non-zero net axial load. However, back-to-back piston indexing provides a proper balance of axial loads, and the benefits of indexing can greatly reduce flow ripple (ie, vibration, noise, pressure and flow waves) (ie, zero). That is nearly half the amplitude of the indexing).

The radial load (F ₂ ) is supported by the shaft-valve 6. The radial load has two components (F _{2, u} and F _{2, l} ). These components F _{2, u} and F _{2, l} are equal and opposite each other, thereby producing a zero net radial load. The moment (M ₁ ) is supported by the swash plate 8 and transmitted to the shaft-valve 6 via the pivot pin 22. An equal and opposite moment (M ₂ ) is supported by the shaft-valve 6 by the differential pressure between the high pressure passage 10a and the low pressure passage 10b (M ₂ is generated by F _{2, u} and F _{2, l} . ) Another force / moment is the moment (M _z ) about the axis of rotation, to which the input torque required to drive the shaft-valve 6 reacts. Thus, the shaft-valve 6 and all other rotating members are balanced. The balance of the rotating member is independent of the tilt angle.

  The housing 4 shown in FIGS. 4 and 6 is a two-piece housing including parts 4a and 4b. However, the housing can be integral or have many components. If the housing 4 has a large number of components, it is advantageous to install a static seal, such as an O-ring or gasket, between the components. The components can be held together by a flange or socket connection. In the two-part configuration, the first stationary housing / cylinder block 4a surrounds the first plurality of pistons, and the second stationary housing / cylinder block 4b surrounds the second plurality of pistons.

  As shown in FIGS. 4 and 6, the component connected to the first surface 12 of the swash plate 8 is substantially sandwiched between the component connected to the second surface 20 of the swash plate 8 and the swash plate 8. It is symmetrical. However, the pump disclosed in the present specification is not limited to the double-sided configuration shown in FIGS. As shown in FIG. 7, an embodiment of a single sided pump 402 is envisioned. In the one-side pump 402, the swash plate 408 is configured at one shaft end of the pump 402. In this case, there is no set of opposing slipper assemblies in this embodiment, so more fluid bearings 403 are required. A single set of slipper assemblies is used. Therefore, the moment and force generated by the rotation of the swash plate and the fluid force can be made zero. A notch 415 or similar relief can be added to the shaft-valve 402 to engage the edge 417 of the housing 404 to compensate for axial forces.

  It can be envisaged that at least one additional pump (not shown) is contained in a single housing and configured substantially similar to the first and / or second pump. In this case, the additional pump is arranged in the housing parallel to the first and second pumps. Thus, the plurality of pumps are housed in a single housing such that two or more shaft-valves are in the housing and parallel to each other. The operation of each pump within the housing can be completely independent of the operation of other pumps within the housing.

  In a further embodiment of the hydraulic pump, the shaft-valve 6 is designed to move with a degree of freedom of 4 if possible. It has 2 linear degrees of freedom and at least 1 rotational degree of freedom about an axis orthogonal to the linear degrees of freedom.

  In an additional embodiment of the pump 502 of the present invention shown in FIGS. 8-12, the shaft and valve are not integrated as in the previous embodiment. With particular reference to FIGS. 8 and 9, the components of pump 502 include a pump housing 504, which includes shaft 506, swash plate 508, hold-down plate 510, one or more Cs. The clip 512, the rotary valve 514, the cylinder block 516, the manifold 518, the swash plate control link 520 for setting the angle of the swash plate 508, the plurality of pistons 522, and the plurality of slipper assemblies 524 are accommodated. Each of the slipper assemblies 524 includes a slipper shoe 526, a slipper ball 528, and a slipper neck 530, with each slipper neck 530 connecting the slipper shoe 526 to a respective slipper ball 528.

  The shaft 506 is disposed along the central axis of the pump housing 504. One end of the shaft 506 extends outside the pump housing 504 and includes a spline 532. The spline 532 is toothed for attachment to a motor gear, crank, flywheel or other motion transmission mechanism. The shaft 506 is held in place by a plurality of bearings at each end of the pump 502.

  The connection between the valve 514 and the shaft 506 is configured to limit the transmission of force from the shaft to the valve to torque around the axis of rotation of the shaft and thrust load along the axis of rotation. Valve 514 is balanced in cylinder block 516 by controlled leakage rates as a function of force and / or moment on valve 514.

  The hydraulic pump 502 takes fluid through the non-rotating cylinder block and sends the fluid through the rotary valve 514. The rotary valve 514 is always in communication with the inlet port 536, the discharge port 612, and the plurality of piston holes 546 in the cylinder block 516.

  In one embodiment, referring to FIGS. 10 and 11A-11B, the low pressure fluid enters the pump through the inlet port 536 and into the inlet cavity 602. This fluid travels through the valve inlet passage 614 to the plurality of piston holes 546. Each piston 522 communicating with the valve inlet passage 614 is pushed out from the piston hole 546. When the piston hole 546 is closed from the valve inlet passage 614, the low pressure fluid is confined. As valve 514 continues to rotate, piston 522 is forced into piston bore 546, thereby compressing fluid within bore 546. At the same time, the piston hole 546 begins to open toward the valve discharge passage 616. The valve discharge passage 616 allows high pressure fluid from the piston hole 546 to be sent to the discharge cavity 618, and the fluid exits the pump through the discharge port 612 at the end of the outlet cavity 612. The force required to move the piston is provided by the input torque to the shaft 506 (FIG. 9). This torque is transmitted to the pivot pin 534 (FIG. 9) and then to the swash plate 508.

  At the end of the shaft 506 opposite to the spline 532, the rotary valve 514 is attached via a C-clip 512 and held in place. Each time the shaft 506 rotates, the rotary valve 514 rotates at the same rotational speed. Similarly, the swash plate 508 is connected to the shaft 506 via a pin 534 between both ends of the shaft 506. Each time the shaft 506 rotates, the swash plate 508 rotates at the same rotational speed as the shaft 506. The shaft 506 functions as a rotational motion transmission element, receives rotational motion from an external motor, and rotates the rotary valve 514 and the swash plate 508, thereby generating pump force.

  Alternatively, the shaft 506 receives rotational motion from the swash plate from fluid passing through the rotary valve, and the rotational motion is converted to rotational motion that rotates the spline 532 (ie, the pump acts as a motor in reverse). In order to reverse the flow (also called going over center), an additional crossing passage must be added on the side of the valve opposite to where it is currently. The reason for adding the additional crossing passage is that the high pressure crossing passage becomes low pressure and the low pressure crossing passage becomes high pressure.

  Referring to FIG. 12, the swash plate 508 is attached to the shaft 506 via a pin 534. Pin 534 of shaft 506 tilts swash plate 506, thereby allowing the axis of swash plate 508 to change relative to the axis of shaft 506. A hollow stem 554 extends from the bottom surface of the swash plate 508. A pin 534 that engages the shaft 506 is disposed within the stem 554. Except where the stem 554 extends from the swash plate 508, the bottom surface of the swash plate 508 is flat. This flat surface of the swash plate 508 allows for a sliding motion between the flat bottom of the swash plate 508 and a slipper shoe (which will be described in more detail below).

  Pin 534 facilitates rotation of swash plate 508 about the axis of shaft 506. Pin 534 allows torque and thrust loads to be transmitted between swash plate 508 and shaft 506. The advantage of pin 534 allows for a one-piece swash plate and provides structural stability by force acting on both sides of swash plate 508.

  In previous pumps, rotating cylinder blocks were used. Therefore, a bolt for holding the housing on the block has been required. The axial pump housing was a combination of two components, and the pump was held together by all bolts. Although it was necessary to deal with the detachment forces of multiple piston areas, the design of the present invention only requires consideration of the weight of the component and the force imposed by the bearing (since it is a one piece housing). This is advantageous in that the life of the bolt is much longer because no tension is applied to the bolt. A seal will also be required to maintain the working fluid in the interface between the block and the housing.

  The radial load on the pump is limited by limiting the inclination of the swash plate. When the tilt angle is small, the side load or moment is small (radial load is small). Thus, the importance of radial bearings is reduced and lighter bearings can be used. Furthermore, the pump of the present invention allows for smaller strokes. The piston is shorter than the conventional piston, but wider. Therefore, the pump of the present invention can be displaced by the same amount as the conventional pump with a smaller swash plate inclination angle. The cylinder block is integral with the pump housing. A reciprocating pump with a fixed cylinder block has structural advantages. By eliminating the rotation of the cylinder block, centripetal forces are eliminated and smaller size and volume structural units are possible.

  Also, since the tilt angle of the swash plate is smaller (less than about 12 ° is required for operation of this pump), the aperture / piston diameter can be increased without reducing the volumetric flow rate. By increasing the diameter and decreasing the swash plate inclination angle, most of the load is not a radial load but an axial load. As a result, the friction and wear on the piston is reduced, which results in higher mechanical efficiency. A series of bearings 542 accommodates moment loads about an axis orthogonal to the pin 534. The bearings 542 are above and below the center point of the pivot pin 534 to balance the high and low pressure sides of the swash plate 508.

  The tilt axis of the swash plate and the center of the slipper ball 528 are in line along the same plane along the same plane parallel to the sliding surface of the swash plate 508. Such a configuration limits the stroke size of the slipper shoe 526 along the sliding surface of the swash plate 508.

  The lemniscate is similar to the shape of an infinite symbol. This can also be described as the number 8. This shape represents the path of the slipper shoe 526 along which the slipper shoe 526 moves around the undersurface of the rotating swash plate 508 during its displacement. In addition to the Remnant skate path, the slipper shoe 526 moves 360 ° around the swash plate 508 (more specifically, the swash plate rotates 360 ° around the non-rotating slipper shoe 526). The slipper shoe 526 moves inward and outward in the radial direction when the swash plate 508 rotates. During this rotation, the fluid in the slipper pressure pocket 550 (FIGS. 12 and 13) is consistently sheared.

  The slipper pressure pocket 550 of each slipper shoe is located at the center of the flat surface of the shoe that abuts the flat lower surface of the swash plate 508. Each respective slipper shoe pressure pocket 550 is connected to a fluid supply so that the fluid pressure at the shoe / swash plate interface is always proportional to the fluid pressure of each piston 522 head.

  Referring to FIG. 13, the slipper ball 528 engages with the piston 522 at the slipper ball receiving portion 544. The slipper ball 528 can actually enter the piston hole 546, which limits the length of the piston 522, thereby limiting the exposed surface (length) of the piston 522 at maximum extension, thereby causing an undesirable moment. Reduce load. When the pump 522 is activated, the axial motion of the piston 522 is transmitted to the slipper ball 528. The slipper ball 528 also rotates about its center, causing lateral movement and resulting forces. Lateral forces produce an undesired moment about an axis perpendicular to the axis of rotation of the shaft. In order to reduce this moment, the lateral movement (and the resulting force) is suppressed. One way to suppress lateral movement and force is to reduce the stroke of the piston 522, as described herein.

  Swash plate 508 rotates with the shaft, but slipper assembly 514 does not rotate with it. Between the swash plate 508 and the slipper shoe 526 is a fluid film bearing 576. Although the slope of the slipper shoe 526 varies constantly, the slipper shoe 526 maintains alignment with the non-rotating piston 522 and the hole 546. Accordingly, the slipper shoe 526 moves in a Remni skate shape within the slipper ball receiving recess 572. The path of the slipper shoe 526 on the lower surface of the swash plate 508 is elliptical (not circular). Further, only the slipper shoe 526 and the holding plate 510 are nutated (ie, nutate) (that is, the vibration of the shaft or the rotation of the tilt axis around the central axis).

  The holding plate 510 holds the slipper assembly 524 in a predetermined position with respect to the swash plate 508. The restraining plate 510 is maintained in place on the swash plate 508 by a series of bearings 542 that are fixedly engaged with the swash plate 508. In order to assemble the holding plate 510 on the swash plate 508, the holding plate 510 has a through hole at the center thereof. The swash plate 508 slides through the center hole of the restraining plate 510, and the series of bearings 542 slide on the swash plate 508.

  The retainer plate 510 includes a plurality of openings, each of which surrounds the neck 530 of the respective slipper assembly 524. Each special washer 578 is fixed integrally with the slipper shoe 526. Each washer 578 keeps the slipper shoe in contact with the flat lower surface of the swash plate 508 via the fluid bearing 576 at all times. The restraining plate 510 is prevented from rotating independently of the slipper shoe 526, and at the same time, the restraining plate 510 does not limit the movement of the slipper shoe 526. The holding plate 510 holds the slipper shoe 526 in the same plane as the swash plate 508 and maintains the pressure between the slipper shoe 526 and the swash plate 508.

  The angle of the swash plate 508 is adjusted by pumping fluid into one of the two volumes (spaces) 582 and 584 (FIG. 12). Fluid is pumped from an external source into the first volume 582 to raise the shift piston 580 or pumped into the second volume 584 to lower the shift piston 580. The shift piston 580 operates the shaft 506 up and down with a hydraulic pressure to change the angle of the swash plate 508. The displacement of the shift piston is controlled from the outside. The hydraulic pressure control is performed by an external hydraulic pressure control mechanism (not shown). Displacement of the shift piston 580 displaces the tilt angle, resulting in displacement of the piston 522 and fluid displacement.

  Still referring to FIG. 13, the working fluid can access the slipper shoe pressure pocket 550 through the through hole 573 of the slipper ball 528. The working fluid acts as both a seal and a bearing for the interaction of the slipper shoe 526 and the flat surface of the swash plate 508. The neck 530 connects the slipper shoe 526 to the slipper ball 528. The slipper neck 530 has a through hole 574 that is a continuous body of the through hole 573 of the slipper ball 528. The slipper ball fluid inlet 572 has a constant flow with the piston through-hole 571 and provides access to the high pressure fluid pumped by the piston 522. For this reason, the fluid inlet 572 is preferably conical due to the constant rotation of the slipper ball 528.

  Referring to FIG. 14, one embodiment of the rotary valve 514 is a two piece assembly. That is, the thrust plate 600 and the valve body 601 are held together by screws 552 (FIG. 9), adhesive or other suitable fasteners. The two parts together act as one valve 514. Thrust plate 600 supports the thrust load generated and carried by shaft 506. The shaft 506 drives the rotary valve 514 only by a torque component. Shaft 506 and rotary valve 514 mesh only to affect the torque or rotary motion of the valve. The rotary valve 514 includes an inlet 602 (FIG. 9), a discharge passage 616, and balancing recesses 606a and 606b. When the shaft 6 and the rotary valve 514 rotate and when the piston 522 swings, the balancing recesses 606a and 606b balance the moment generated by the movement of these components by the pressure in the balancing recesses 606a and 606b. Configured to receive a working fluid. A pressure notch 620 is placed directly on the curved surface of the rotary valve at the distal edge of the discharge passage 616. This pressurization notch 620 is provided to reduce the damage caused to the valve by a sudden rise in pressure in the discharge passage 616. The pressurization notch controls the pressure rise rate within the individual pump chambers to minimize the impact force on the piston caused by the pressure rise. By minimizing the impact force on the piston, the incidental noise that results from contact between the slipper shoe and the swashplate and ultimately is released to the local environment is minimized.

  The thrust plate 600 is disposed on the upper portion of the rotary valve 514 opposite to the inlet passage 602. The surface area of the thrust plate 600 is equal to half of the total surface area of the piston (since half of the surface area is subjected to high pressure at any point in time). A major obstacle for axial piston pumps is to deal with large thrust loads. The thrust load is generated by continuously shifting high pressure applied from the piston hole 546 to the piston. Depending on the position of the piston in the hole, half of the piston has a high pressure flow and the other half has a low pressure flow. Lubrication occurs at high or low pressure. The load is transmitted from the piston to the slipper shoe 526, to the swash plate 508, to the shift pin 534, and finally to the shaft 506. The thrust plate 600 is used to counteract the thrust load. Without the thrust plate, the thrust load causes the shaft 506 to move away from the cylinder block 516.

  Since the film strength along the surface of the rotary valve is sufficient to counter the smaller inlet pressure value, there is no need to install a low pressure recess in the rotary valve. In addition, adding additional tolerance channels results in excessive fluid loss.

  As can be seen from FIGS. 11A and 11B, the recesses 606 a and 606 b of the rotary valve 514 form a volume together with the cylinder hole 546. Thus, the pressure is equal throughout the volume. The recesses 606a and 606b help counteract the moment caused by pressure rotation around the plurality of pistons 522. Piston 522 does not rotate about the axis of shaft 506. However, the reciprocating motion of the piston 522 within each hole 546 produces a moment about an axis orthogonal to the axis of the shaft 506. The recesses 606a, 606b allow the inlet passage 536 to be aligned along the axis of the rotary valve (and the axis of the pump). Since it is a non-rotating piston, centripetal force is eliminated. Accordingly, the rotary valve 514 is self-aligned using a cross porting hydrodynamic balance.

  The recesses 606a and 606b in FIGS. 11A-11B represent high pressure areas, while the inlet 602 represents low pressure. The inlet pressure is close to atmospheric conditions unless pressure is supplied. The outlet pressure can reach 6000 psi (41.38 MPa) and above. The two recesses 606a and 606b are separated by a high pressure outlet passage 612. The recess 606a is larger than the recess 606b to balance the moment about an axis orthogonal to the rotational axis. The surface between the passages 606a and 606b is long enough to allow the valve to properly seal in its hole. The length of each of the passages 606a and 606b is based on the fluid pressure. The following formula is used to determine the ratio of 606a to 606b.

  The hydraulic thrust bearing 610 is a fluid filled volume between the thrust plate 600 and the manifold 518.

  The projected areas of the high pressure outlet 612 and the valve discharge passage 616 are of equal size, thereby balancing the lateral forces on the rotary valve 514. Furthermore, these areas are positioned on the y-axis such that the sum of moments around the center of mass is zero. The thrust bearing at the top of the pump can handle one piston worth the area at the working pressure.

  The rotary valve inlet flow passes along the rotary valve axis (and hence along the pump 502 axis) through the central inlet passage 602 of the valve, and the rotary valve 514 outlet flow flows around the periphery of the pump frame. . The location of the outlet 612 has the additional advantage of helping to cool the pump 502. An inlet passage 602 through the center of the rotary valve 514 allows a more direct flow path to the piston hole 546. This reduces the volume of the suction cavity and also reduces fluid friction as it provides a more direct flow path and smaller surface area, thereby reducing parasitic losses.

  The pump 502 combines a piston housing and a discharge cavity (manifold) into one unit (component). By coupling high pressure into one housing, the separation forces that would otherwise be present in prior art pumps are eliminated. Without the balancing effect of the rotary valve 514, the valve can tip or tilt, damaging the fluid barrier, and the rotating parts of the pump 502 can become stuck. This force causes the rotary valve to move toward its center inside the valve hole.

  The following embodiment is for avoiding the wear 698 of the rotary valve as shown in FIG. Such wear is typically caused by eccentric rotation and tipping of the valve within the valve bore of the pump.

  Referring to FIG. 16, a first disclosed embodiment is a pump 702 that includes a housing 704 and a manifold 706 integrated therein. The rotary valve 708 is surrounded by a manifold 706. The rotary valve 708 includes a shaft 710, a ported section 712, and a seal portion 714 (FIGS. 17A-17B). Seal portion 714 includes a semi-circular seal ridge 716. With particular reference to FIG. 17C, a pressure drop across the seal ridge 716 is shown. The fluid passing through the port is at a much lower pressure compared to the pressure outside the port as indicated by the pressure gradient 740.

  18 and 19, the rotary valve 708 includes a first axial seal 718 at the first end 720 of the rotary valve 708 and a second axial seal 722 at the second end 724 of the rotary valve 708. . The high pressure discharge 726 is on the side of the rotary valve opposite the semicircular seal ridge 716. The intake 728 is at the first end 720 of the rotary valve 708.

  As shown in FIG. 20, the working gap 730 is between the high pressure exhaust 726 and the manifold 706. The operating gap 730 is a width that depends on the force applied in the direction of the semicircular seal ridge at the high pressure discharge 726. The side of the hole that engages the seal 714 is subjected to a much lower pressure than the opposite side of the rotary valve. This is because the opposite side of the rotary valve is subjected to very high pressure by the discharge of fluid from the cylinder in the pump. The seal 714 maintains a generally unchanged operating distance from the manifold 706 throughout the life of the pump, as the high pressure on the discharge side of the rotary valve pushes the rotary valve toward the hole on the low pressure side of the rotary valve. No matter how much the low pressure side of the valve (ie, the seal ridge 716 of the seal 714) is worn, the seal 714 engages the hole surface (via a bearing / fluid bearing / mechanical bearing, etc.) Always push the valve towards the hole. Thus, the rotary valve always maintains a constant distance from the hole surface, even if the valve surface wears. The width of the working gap also depends on the speed of the shaft and the circumference of the valve.

  The rotary valve 708 is generally cylindrical and a semi-circular seal ridge 716 surrounds the rotary valve 708 less than 360 °. Semicircular seal ridge 716 includes a manifold engagement portion 734 and a recessed portion 736. Ribs 748 are included to strengthen the passage 750.

  Another disclosed embodiment is shown in FIGS. 21 and 22A, a pump 800 having a valve hole 802 surrounds a rotary valve 804. The rotary valve 804 has at least one axial face seal 806 at one end of the rotary valve 804 and at least one radial face seal 808 at the curved surface of the rotary valve 804. An operating sealing clearance gap 810 (FIG. 24) is between the radial face seal 808 and the valve hole 802. The working seal gap 810 maintains a generally unchanged thickness when the rotary valve 804 is actuated.

  The discharge gap 812 is disposed between the rotary valve 804 and the hole 802. The rotary valve 804 is biased in the direction of the radial face seal 808. The bias is generated by the applied force. The amount of force applied determines the width of the working seal gap 810. The applied force is determined by the high pressure fluid exiting pump 800 through valve 801.

  With particular reference to FIG. 23, a pressure drop 836 across the seal ridge 716 is shown. The pressure of the fluid passing through the inlet passage 820 is much lower than the pressure outside the passage portion 820 as shown by the pressure gradient 836.

  The rotary valve 804 is a unitary structure and is composed of two parts, namely a passage part 830 and a shaft 814. The radial face seal 808 has a semi-circular seal piece 816 and is generally coaxial with the shaft 814. The passage portion 830 is coaxial with the shaft 814 in this embodiment. The passage portion 830 includes an inlet passage 820 and a discharge passage 822 (FIG. 22B). The pump 800 takes fluid through the inlet passage 820 and the fluid enters the piston hole that houses the piston as described in connection with previous embodiments. The piston is actuated by a rotating swash plate as previously described. The fluid is forced out of the pump by the same piston at a much higher pressure than when the fluid enters the pump. The fluid exits the pump 800 via the discharge passage 822 in the passage portion 830. As each piston pushes fluid toward the passage portion 830, the passage portion 830 rotates with the shaft 814 so that the discharge passage 822 receives high pressure fluid from the piston. The high pressure fluid is then directed toward the outlet 832 of the pump 800.

  In yet another embodiment shown in FIGS. 25 and 26, the pump 900 has a rotary valve 906 in the valve hole 902 and the hydraulic pump 900 includes a manifold 936. The shaft 904 is disposed within the manifold and is attached to the planar surface 910 of the rotary valve 906 at the distal end 908. Referring to FIG. 26, the axis of the shaft, ie, the first axis 912, and the axis of the rotary valve, ie, the second axis 914, are offset from each other and are not coaxial. Accordingly, at least one of the shafts is not coaxial with the valve hole 902 when the shaft 904 and the rotary valve 906 are joined together and stationary.

  The second shaft 914 coincides with the shaft 916 of the valve hole 902. Sometimes the second shaft 914 is coaxial with the shaft 916 of the valve hole 902, but the second shaft 914 may not be completely coaxial with the shaft 916 of the valve hole 902. At this time, the second shaft 914 and the valve hole shaft 916 are considered to be coincident, not coaxial, because they are close together and can be coaxial and optimal coaxial rotation.

  The first axis 912 is offset from the center line 918 of the manifold 936. The rotary valve 906 has a high pressure outlet 920. The shaft 904 is offset in the opposite direction to the high pressure outlet 920 of the rotary valve 906. However, the shaft 904 can be offset closer to the high pressure outlet 920 of the rotary valve 906.

  The plurality of cylinders 926 within the manifold 936 are arranged in a circle that is parallel and coaxial with the valve hole 902. The rotary valve 906 is generally coaxial with the valve hole 902. However, the rotary valve 906 may not be coaxial with the valve hole 902. The shaft 904 can be integrated with the rotary valve 906, but this is not required.

  The shaft 904 includes a manifold contact surface 922. The bearing 924 is disposed in the manifold 934 and abuts against the shaft 904. When the shaft 904 is offset in a direction away from the high pressure outlet 920, the high pressure fluid flowing through the high pressure outlet pushes the second shaft 94 of the rotary valve 906 toward the shaft 916 of the valve hole 902. Thus, the offset of the rotary valve 906 from the shaft 904 helps to cope with the high pressure fluid flowing through the rotary valve 906 and thus helps eliminate the eccentricity caused thereby.

  Alternatively, when the shaft 904 moves away from the high pressure outlet 920, the rotary valve is moved away from the high pressure outlet 920. This helps to ensure a seal between the low pressure side of the rotary valve 906 (the side of the valve opposite the high pressure outlet 920) and the valve hole 902.

  FIG. 27 shows a portion of a hydraulic pump 1000 that includes a valve 1002 in a hole 1004. A radial face seal 1006 is on the radial surface 1008 of the valve 1002 and is disposed between the valve 1002 and the hole 1004. The radial face seal 1006 at least partially forms the circumference of the valve 1002. The seal ridge 1010 is on the outer periphery of the radial surface seal 1006. Further, the passage 1024 is disposed at an intermediate position along the height of the radial surface seal 1006.

  The seal ridge 1010 protrudes from the radial surface seal 1006 toward the hole 1004. By minimizing the gap between the seal ridge 1010 and the hole 1004, a hydrodynamic bearing can be formed between the seal ridge 1010 and the hole 1004. The hydrodynamic bearing allows for smooth rotation of the valve within the hole 1004.

  The radial face seal 1006 is C-shaped to fit coaxially with the valve 1002. As shown in FIG. 32, the O-ring 1048 is disposed between the radial face seal 1006 and the valve 1002. O-ring 1048 is generally oval to enclose passage 750. Relief 1050 can be cut into the valve for the O-ring (or elsewhere for other O-rings), but this is not necessary. Still referring to FIG. 17c, the recess 734 (similar to the portion of the radial face seal above and / or below the seal ridge 1010 in this embodiment) to the passage 712 (from the passage 1024 of this embodiment shown in FIG. 29). There is a pressure drop towards (similar). This is because the fluid entering the pump enters through the passage 1024 at a lower pressure compared to the portion of the radial face seal above and / or below the seal ridge 1010. The pressure at the upper and lower edges of the seal ridge 1010 is higher than the portion of the seal ridge 1010 close to the hole. Further, as in some previous embodiments, higher pressure on the discharge side of valve 1002 pushes valve 1002 toward hole 1002 on the low pressure side of valve 1002, ie, the seal ridge side of valve 1002. The high pressure side always pushes the radial face seal 1006 towards the hole 1004 so that the seal ridge 1010 engages the hole surface (via a bearing / fluid bearing / mechanical bearing etc.). The seal ridge 1010 completely surrounds the passage 1024 so that the passage is shielded as much as possible from the differential pressure inside the pump.

  The radial face seal 1006 is connected to the valve 1002 via at least one radial face seal pin 1016 (also referred to as a “torque translator”). Each radial face seal 1006 is biased away from the valve 1002 by a radial face seal spring 1018. Torque translator 1016 transfers the rotational force from valve 1002 to rotating surface seal 1006. The torque translator can be a pin, a protrusion, a high friction surface such as an O-ring, or any other means that can transfer torque from one component to another. Since one of the purposes of the radial face seal 1006 is to provide a seal between the seal ridge 1010 and the hole 1004, the radial face seal 1006 can be operated with no differential pressure or no pump at all. It is biased towards hole 1004 so that the seal is maintained regardless. Each radial face seal pin 1016 sits inside the recess 1036 of the valve 1002 and each radial face seal spring 1018 is generally coaxial with the pin and sits on the shoulder 1038.

  Axial seal 1012 is on the axial surface 1014 of valve 1002. Axial seal 1012 is between valve shaft surface 1014 and shaft surface 1015 of hole 1004. The ridge 1017 abuts against the axial surface 1015 of the hole via a fluid bearing. Axial seal 1012 is also connected to valve 1002 via at least one axial seal pin 1020. A second shaft surface seal 1026 configured similar to the shaft surface seal 1012 is disposed on the second shaft surface 1003 of the valve 1002. Similar to the shaft surface 1015, the seal on the second shaft surface contacts the hole through the ridge.

  Each shaft seal 1012 is biased away from the valve 1002 by a shaft seal spring 1022. As can be seen from FIGS. 27 and 28, the O-ring 1030 is installed between the radial face seal 1006 and the radial surface 1008 and between the axial face seal 1012 and the shaft. As can be seen from FIG. 29, the seal ridge 1010 has a passage 1024, which is an elliptical through hole that provides fluid access to the valve 1002. The O-ring 1030 provides a static seal between the radial seal and the valve while providing a seal while allowing small radial movement between the two components. Axial seal 1012 provides a similar seal through the O-ring while permitting movement between the axial seal and the valve.

  FIG. 30 shows a two-sided hydraulic pump 1100. The hydraulic pump 1100 includes a swash plate 1102 having a first surface 1104 and a second surface 1106. The first surface of the swash plate 1104 is substantially parallel to the second surface 1106.

  The first portion 1108 of the pump lies on the first surface 1104 of the swash plate. The second portion 1110 of the pump lies on the second surface 1106 of the swash plate. The first portion 1108 has a first partial rotary valve 1112 and is surrounded by a first partial hole 1114. The second portion 1110 has a second partial valve 1116 and is surrounded by the second partial hole 1118. Both valves have shafts 1120a, 1120b and radial seals 1126a, 1126b. Each radial seal has seal ridges 1124a, 1124b and passage portions 1122a, 1122b that communicate with the respective passages 1125a, 1125b in the respective valves 1112, 1116.

  At least one of the portions 1108, 1110 has a respective radial face seal connected to the respective valve 1112, 1116, the radial face seal between the respective valve 1112, 1116 and the respective hole 1114, 1118. Give a seal. Torque translator 1128 connects each radial face seal 1126a, 1126b with each valve 1112, 1116 to transmit the rotational force from valves 1112, 1116 to each radial face seal 1126a, 1126b. O-rings 1034a, 1034b are installed between both radial face seals 1126a.

  In order to shield the fluid flow through the pump from the rest of the pump, the axial seal 1026 must be on the side of the valve opposite the portion facing the valve. As can be seen from FIG. 31, the radial face seal 1006 is subjected to a pressure gradient. The pressure gradient results from applying pressure to the valve 1002 from the fluid in the cylinder 1042 where the higher pressure portion 1040 is forced into the valve by the piston. The pressure at the side end 1044 of the radial face seal 1006 is greater than the pressure at the midpoint 1046 of the radial face seal 1006.

  The above disclosure will be described as a pump. However, one of ordinary skill in the art will appreciate that the disclosed apparatus can function as a hydraulic motor, engine, or the like.

  Although what has been described above as to what appears to be the best mode and / or other embodiments, various modifications can be made to it, and what is disclosed herein can be realized in various types of embodiments. And it should be understood that the teachings can be applied in a variety of applications, only some of which are described herein. The following claims are intended to claim all applications, modifications, and variations that fall within the true scope of the present teachings.

Claims (62)

A stationary cylinder block;
A shaft rotatable about its own axis in the stationary cylinder block, the shaft comprising a plurality of passages configured to receive, direct and discharge fluid, the shaft operating the pump A shaft, sometimes rotatable concentrically and eccentrically,
A swash plate coupled to the shaft, the swash plate having a first surface coupled to the first plurality of pistons;
A piston device.   The piston device according to claim 1, wherein the swash plate is configured to be inclined with respect to an axis orthogonal to the axis of the shaft.   The piston device according to claim 2, further comprising means for inclining the swash plate.   The piston apparatus according to claim 1, wherein fluid pressure on the passage of the shaft balances at least one force and / or moment.   The piston device according to claim 1, further comprising a respective plurality of slipper shoes each connected to a piston, wherein the swash plate comprises at least one working surface for interacting with the slipper shoes.   A fluid flow through the entire pump generates at least one of a force and a moment on the shaft, and the plurality of passages are configured on the shaft to balance at least one of the force or moment on the shaft. The piston device according to claim 1.   The piston device according to claim 6, further comprising a second plurality of pistons, wherein each of the plurality of pistons is accommodated in the cylinder block, and is arranged in an arc shape coaxially with the shaft at both ends of the cylinder block.   The piston device according to claim 7, further comprising a first working surface and a second working surface, wherein the first and second working surfaces are on opposite sides of the swash plate. A stationary cylinder block;
A shaft extending through the stationary cylinder block and having a plurality of passages along a portion of the shaft;
A valve enclosed within the stationary cylinder block, the valve comprising the shaft integrally coupled with the valve;
A piston device.   The piston device according to claim 9, further comprising a first plurality of pistons and a second plurality of pistons, wherein each of the plurality of pistons is disposed in an arc shape coaxially with the shaft on the opposite side of the shaft.   The piston apparatus of claim 9, wherein the plurality of passages are configured to receive, direct, and discharge a fluid flow through the pump.   The fluid flow through the entire pump generates at least one of a plurality of forces and moments on the shaft, and the plurality of passages are the at least one of the plurality of forces and the plurality of moments. The piston device of claim 11, wherein the piston device is configured to balance the shaft. A plurality of slipper assemblies, wherein each slipper assembly is coupled to one piston of the plurality of pistons;
A swash plate coupled to the shaft and having at least one working surface for interacting with a respective slipper assembly;
The piston device according to claim 12, comprising:   The piston device according to claim 13, further comprising a first working surface and a second working surface, wherein the first and second working surfaces are on opposite sides of the swash plate. A swash plate having a first surface coupled to the first plurality of slipper shoes and a second surface coupled to the second plurality of slipper shoes, wherein the first surface of the swash plate is the swash plate. A swash plate that is generally parallel to the second surface of
A first portion of the pump on the first surface of the swash plate and a second portion of the pump on the second surface of the swash plate, on the first surface of the swash plate A first portion and a second portion, wherein the first portion of the pump is substantially symmetrical with respect to the second portion of the pump across the swash plate;
A piston device.   The piston apparatus according to claim 15, further comprising a shaft, wherein the swash plate is coupled to the shaft.   The piston device according to claim 16, configured to be adjustable with respect to an axis that is oblique or perpendicular to the axis of the shaft.   The piston apparatus of claim 16, further comprising means for balancing at least one of a plurality of forces and a plurality of moments acting on the shaft.   18. The piston device according to claim 17, further comprising means for inclining the swash plate with respect to the shaft.   The piston apparatus according to claim 16, further comprising a housing that houses the shaft, the swash plate, and the slipper assembly. A stationary cylinder block;
A valve-shaft assembly surrounded by the stationary cylinder block, the valve-shaft assembly via an interaction between a drive key, a valve keyway of the valve and a shaft keyway of the shaft; A valve-shaft assembly comprising a valve connected to a shaft of the valve-shaft assembly, wherein the interaction is configured to provide at least one degree of freedom of the valve relative to the shaft;
A piston device.   23. The valve according to claim 21, wherein the valve is configured to balance within the cylinder block by a controlled fluid leakage rate within the hydraulic pump as a function of at least one of forces and moments against the valve. Piston device.   The piston device according to claim 22, further comprising a swash plate connected to the shaft.   24. The piston device according to claim 23, further comprising means for tilting the swash plate with respect to an axis orthogonal to the axis of the shaft.   25. A piston arrangement according to claim 24, wherein the swash plate comprises at least one working surface.   26. The piston device of claim 25, wherein the swash plate comprises a plurality of slipper assemblies coupled to the at least one working surface of the swash plate. A housing and a manifold integral with the housing;
A rotary valve surrounded by the manifold, the rotary valve comprising a shaft, a passage portion and a seal portion;
With
The seal portion comprises a semi-circular seal edge;
Piston device.   28. The piston device according to claim 27, wherein the rotary valve further comprises a first axial seal at a first end of the rotary valve and a second axial seal at a second end of the rotary valve.   29. The piston device according to claim 28, wherein the rotary valve further comprises a high pressure discharge on the side of the rotary valve opposite the semicircular seal ridge.   30. The piston device according to claim 29, wherein the rotary valve further comprises an intake at the first end of the rotary valve.   30. A working gap formed between the high pressure discharge and the manifold, the working gap having a width depending on the force applied in the direction of the semicircular seal gap at the high pressure discharge. The piston device according to 1.   32. The piston apparatus of claim 31, wherein the semi-circular seal ridge is configured to maintain a generally unchanged working distance from the manifold.   32. A piston arrangement according to claim 31, wherein the width of the working gap is further dependent on shaft speed and the circumference of the valve.   28. The piston device of claim 27, wherein the rotary valve is generally cylindrical and the semicircular seal ridge surrounds less than 360 degrees of the rotary valve.   35. The piston device of claim 34, wherein the semicircular seal ridge comprises a manifold engagement portion and a recess.   29. The piston device according to claim 28, wherein the rotary valve further comprises a low pressure suction part on a side of the rotary valve opposite to the semicircular seal ridge. Valve holes,
A rotary valve surrounded by the valve hole, the rotary valve comprising at least one axial face seal and at least one radial face seal;
A working seal gap between the radial face seal and the valve hole;
With
The operating seal gap maintains a generally unchanged thickness during operation of the rotary valve;
Piston device.   Further, a discharge gap is provided between the rotary valve and the valve hole, and the rotary valve is urged by a force applied in the direction of the semicircular seal ridge, and the force determines the width of the working seal gap. The piston device according to claim 37.   40. The piston apparatus of claim 38, wherein the rotary valve further comprises a shaft, and the radial face seal comprises a semi-circular sealing element and is generally coaxial with respect to the shaft.   40. The piston device according to claim 39, further comprising a passage portion on the shaft.   41. The piston apparatus of claim 40, wherein the shaft comprises an inlet passage and a discharge passage. A housing having a valve hole and a manifold;
A shaft in the manifold and a rotary valve in the valve bore, wherein the shaft is attached to a planar surface of the rotary valve at a distal end;
With
The shaft has a first axis, the rotary valve has a second axis, and the first axis is offset from the second axis;
Piston device.   43. The piston device according to claim 42, wherein the second axis coincides with the axis of the valve hole.   44. The piston device according to claim 43, wherein the first axis is offset from a centerline of the manifold.   43. The piston device of claim 42, wherein the rotary valve comprises a high pressure outlet.   46. The piston device of claim 45, wherein the shaft is offset in the direction of the high pressure outlet of the rotary valve.   47. The piston device of claim 46, wherein the shaft comprises a manifold contact surface.   46. The piston device of claim 45, wherein the shaft is offset in a direction opposite to the high pressure outlet of the rotary valve.   43. The piston device of claim 42, further comprising a plurality of cylinders in the manifold arranged in parallel, circularly around the valve hole and coaxial with the valve hole.   50. The piston apparatus of claim 49, wherein the rotary valve is generally coaxial with the valve hole.   51. The piston device according to claim 50, wherein the shaft is not coaxial with the valve hole.   43. The piston device of claim 42, wherein the shaft is integral with the rotary valve. A valve in the valve hole,
A radial face seal on a radial surface of the valve located between the valve and the hole, the radial face seal comprising a seal ridge; and
An axial seal on the axial surface of the valve, wherein the axial seal is between the axial surface of the valve and the axial surface of the valve bore;
A piston device.   54. The piston device according to claim 53, wherein the seal ridge protrudes away from the radial surface of the valve toward the valve hole.   54. The piston arrangement according to claim 53, wherein the radial face seal is connected to the valve via at least one radial face seal torque transmission device.   56. A piston arrangement according to claim 55, wherein the radial face seal is biased away from the valve via at least one radial face seal spring.   54. The piston device of claim 53, wherein the shaft seal is connected to the valve via at least one shaft seal torque transmission.   58. The piston device according to claim 57, wherein the at least one shaft seal is biased away from the valve via a shaft seal spring.   54. The piston device of claim 53, wherein the seal ridge further comprises a passage. A swash plate having a first surface and a second surface, wherein the first surface is substantially parallel to the second surface;
A first portion of the pump on the first surface and a second portion on the second surface;
With
The first portion includes a first partial rotary valve surrounded by a first partial hole, and the second portion includes a second partial valve surrounded by a second partial hole;
Both valves comprise a shaft, a passage and a seal ridge,
Piston device.   61. The piston device of claim 60, wherein at least one of the portions further comprises a respective radial face seal connected to the respective valve and providing a seal between the respective valve and the respective hole.   62. The piston apparatus of claim 61, further comprising a torque translator that connects the respective radial face seals to the respective valves to transmit rotational force from the valves to the respective radial face seals.

Images