Patent Application
20090223359
1. Field of the Invention
The present invention relates to a hydraulic machine arrangement, and in particular, a hydraulic machine arrangement including at least one hydraulic machine that may operate as a pump, a motor, or both.
2. Background Art
It is well known that hydraulic regenerative systems promise improved efficiency over electric regenerative systems incorporating a battery. Hydraulic regeneration involves using a pump connected in the vehicle drive train as a retarding device, and then storing the resulting high pressure fluid in an accumulator. On subsequent vehicle acceleration, the high pressure fluid from the accumulator is routed to a hydraulic motor and the stored energy is recovered in the form of mechanical work which drives the vehicle forward. A low pressure accumulator acts as a reservoir to make up for fluid volume variations within the high pressure accumulator, and also provides a charge pressure to the inlet side of the pump. Integral to a system such as this are hydraulic machines—i.e., hydraulic pumps, motors, or machines that can operate as both a pump and a motor as desired.
One method of modulating braking and driving forces in hydraulic regenerative systems is to incorporate a variable displacement hydraulic machine to operate in concert with an accumulator whose pressure is a function of its state of charge. Conventional variable displacement hydraulic machines may vary the piston strokes to achieve the desired power modulation. Such devices can be bulky, heavy and expensive. Moreover, they do not package easily in automotive passenger vehicles, especially in the front of a vehicle, where space is limited.
One way to overcome the limitations associated with conventional variable displacement hydraulic machines is to use a fixed displacement machine. Such a machine is generally smaller and lighter than its variable displacement counterpart, but it does not allow the power modulation required in most applications. One solution to this problem is to use a fixed displacement hydraulic machine in conjunction with a variable ratio hydraulic transformer to facilitate the desired power modulation. Systems utilizing transformers such as these are described in U.S. patent application Ser. No. 10/535,354, entitled “Hydraulic Regenerative Braking System for a Vehicle,” filed on 18 May 2005, which is hereby incorporated herein by reference.
As an alternative to a transformer, it may be desirable to have a system that included a relatively compact variable displacement hydraulic machine, thus eliminating the requirement of a separate variable ratio transformer. Variable displacement hydraulic machines are described in U.S. patent application Ser. No. 11/721,903, entitled “Hydraulic Regenerative Braking System and Method for a Vehicle,” filed on 15 Jun. 2007, and U.S. patent application Ser. No. 11/913,971, entitled “Hydraulic Regenerative Braking System for a Vehicle,” filed on 9 Nov. 2007, each of which is hereby incorporated herein by reference.
Embodiments of the present invention provide a hydraulic machine arrangement including at least one hydraulic machine operable as a motor, a pump, or both. In particular, embodiments of the present invention may operate as a motor, such that hydraulic pressure is provided as an input, and torque is provided as an output. Other embodiments may receive torque as an input—e.g., the rotational force of a vehicle axle or drive shaft—and provide increased hydraulic pressure as an output. Embodiments of the present invention may be selectively operable as a motor in one mode and as a pump in another.
Embodiments of the invention may also provide a hydraulic machine arrangement that includes at least one hydraulic machine operable as a pump configured to be driven by a shaft, thereby increasing the pressure of fluid flowing through the hydraulic machine. The hydraulic machine may further be operable as a motor configured to be driven by pressurized fluid, thereby providing torque to the shaft. Such a hydraulic machine may include a port housing having a high pressure fluid port and a low pressure fluid port, and a cylinder block having a plurality of radial pistons. Each of the pistons is configured to reciprocate within a corresponding cylinder in the cylinder block, and has a corresponding piston stroke. The pistons pump fluid when the hydraulic machine is operating as a pump, and provide torque when the hydraulic machine is operating as a motor.
Each of the pistons includes a corresponding cam follower. A cam is disposed at least partly within the cylinder block, and has a plurality of lobes configured to cooperate with the cam followers to translate relative rotational motion between the cam and the cylinder block into linear motion of the pistons when the hydraulic machine is operating as a pump, and to translate linear motion of the pistons into relative rotational motion between the cam and the cylinder block when the hydraulic machine is operating as a motor. The rotational motion is described as relative, since, as described more fully below, some embodiments of the present invention may employ a rotating cam and stationary cylinder block, while others may employ a rotating cylinder block and a stationary cam. A valve plate, or manifold, includes a plurality of apertures therethrough, at least one of which communicates with the high pressure fluid port and at least one of which communicates with the low pressure fluid port. The valve plate is configured to connect at least one of the cylinders with the high pressure fluid port and at least one other of the cylinders with the low pressure fluid port.
The valve plate is movable relative to the cylinder block to effect a first transition to disconnect the at least one cylinder from the high pressure fluid port and connect it with the low pressure fluid port, and to effect a second transition to disconnect the at least one other cylinder from the low pressure fluid port and connect it with the high pressure fluid port. In some embodiments, the valve plate is movable such that the first and second transitions can be effected at a plurality of piston positions within a corresponding piston stroke, thereby facilitating a “continuously variable displacement” operation of the hydraulic machine. In still other embodiments, variable displacement is achieved by disengaging one or more of the pistons, thereby providing a “discrete variable displacement”, and in some embodiments, a combination of a movable valve plate and piston disengagement may be utilized.
Disengaging one or more of the pistons to operate the machine at less than full displacement may provide efficiency gains over other configurations for varying the displacement. The disengagement of one or more of the pistons may be effected in any of a number of different ways. For example, for a hydraulic machine operating as a motor, one method involves disengaging the non-driving pistons by increasing the pressure in the housing—i.e., the case pressure—to be equal with the return pressure. This balances the hydraulic forces on the piston, and allows the centrifugal force to dominate, thereby keeping the deactivated pistons in the outer retracted position separated from the cam during particular segments of the rotation. If accumulators are used in a regenerative installation, then the return pressure and the case pressure will be set by the pressure in the low pressure accumulator. It should be noted that the disengagement is synchronized with particular cam lobes, not particular cylinders, so the disengaged cylinders alternate as they pass by the continuously low pressure ports synchronized to a particular set of cam lobes.
Another configuration that can be used in embodiments of the present invention, involves disengaging the non-driving pistons of a hydraulic machine operating as a motor by decreasing the return pressure to near zero to equal the case pressure. This may be accomplished, for example, by using a high capacity pump, such as a jet pump, in the main flow circuit to pump the near zero return pressure back up to the low pressure accumulator pressure level. Systems of this type have the advantage of allowing partial evacuation of the case with the rotating cylinder block inside, allowing just enough fluid to keep the piston/cam rollers splash lubricated and lifted off their plain bearing in the power piston. Efficiency of jet pumps is affected by the location, size, and shape of the jets as they redirect some of the output flow back to the inlet passage. Control can be attained by use of a proportional valve capable of throttling the redirected flow.
Other embodiments may connect the ports for both power and return to exhaust passages, for example, with individual two-way poppet valves. For a 9 lobe cam, there are 18 feed ports corresponding to the 18 cam ramps. The distribution of the 18 cam ramps can be, for example, as follows: 3 equally spaced deep down ramps, 3 equally spaced deep up ramps, 3 equally spaced medium down ramps, 3 equally spaced medium up ramps, 3 equally spaced shallow down ramps, 3 equally spaced shallow up ramps. In one embodiment, the deep down and up ramps may have a stroke of approximately 0.220 inches, the medium down and up ramps a stroke of approximately 0.098 inches, and the shallow down and up ramps a stroke of approximately 0.061 inches. For pump mode operation, the up ramps are connected to the high pressure ports and the down ramps are connected to the low pressure ports. For motor mode, the port housing, or manifold, which contains the ports is indexed relative to the cam, such that the down ramps are connected to the high pressure ports and the up ramps are connected to the low pressure ports.
To provide smooth and quiet operation of a hydraulic machine arrangement, embodiments of the present invention may provide cam lobes that are specifically configured such that the sum of the velocity curves for all the lobes is a straight line. The nose radius of the cam lobes may also be equal to or greater than the radius of the cam follower, or roller, to reduce Hertz stress. Embodiments of the invention also provide piston velocity profiles that are compatible with the flow area of the hydraulic fluid as the valve plate opening varies from fully closed to fully open, and back again. Thus, cams for hydraulic machine arrangements of the present invention may be configured such that the maximum piston velocity occurs when the flow area is near a maximum, not, for example, when the port is at the cracking point—i.e., just opening—and the flow area is near a minimum.
In embodiments of the hydraulic machines described above, high pressure fluid may enter the machine through a port housing, thereby imparting an axial load on at least a portion of the machine. In order to balance the force caused by the high pressure fluid, a large tapered roller bearing can be used. Such a solution has some disadvantages, however, in that such bearings tend to be expensive and occupy a large amount of space, as well as incurring parasitic losses associated with the rolling friction of high loads. As an alternative to using the large tapered roller bearing, embodiments of the present invention add a pressure balance area on the cylinder block on the opposite face from the direction of the fluid load. High pressure fluid is fed to a floating piston, such that the majority of the thrust load can be balanced hydraulically, and only a small portion of the thrust load transmitted to a lighter duty roller, ball, or journal bearing.
The balance piston described above is configured such that the area separating the piston face from the cylinder block is slightly larger than the area applying the piston. An orifice or restricted flow passage in the piston causes a pressure drop through the piston such that the pressure drop is proportional to the square of the flow velocity through the passage. This allows the balance piston to find a position such that the feed pressure times the applied area equals the separating area times the reduced pressure. The balance piston position is self-regulating. If leakage increases, the separating pressure drops, and the piston moves to decrease the leakage. Conversely, if leakage decreases, the separating pressure increases, and the piston moves to increase the leakage. In summary, the balance force on the cylinder block face is equal to the feed pressure times the applied area of the balance piston. The design of the flow restrictor is adjusted to minimize the loss due to high pressure fluid leakage while maintaining a film of fluid between the rotating cylinder block and the stationary balance piston.
With a multi-speed hydraulic machine, it may be desirable to have more than one balance piston. In such a configuration, each of the balance pistons can balance a proportional share of the unbalanced thrust load. For example, with the seven speed, 9 lobe, 13 piston machine described above, three balance pistons may be used. Each of the balance pistons connects with a feed passage through which it receives high pressure fluid. By having separate feed passages, one of the balance pistons is operational when one bank of cam lobes is operational, two of the balance pistons are operational when two bank of the cam lobes are operational, and all of the balance pistons are operational when the hydraulic machine is operating at full capacity.
Another way to balance some of the high axial forces induced in hydraulic machines of this type, is to configure a hydraulic machine arrangement with two hydraulic machines mounted back-to-back. Such an arrangement may be particularly well suited for mounting motors, particularly for automotive vehicles where two motors are used to drive two axle shafts. By mounting the motors back-to-back in a single housing, heavy duty bearings and balance pistons may be eliminated. manifold in a radial piston motor or pump. The thrusts of the two machines balance each other, and because there is minimum relative speed between the two axles, a plain thrust washer or rolling element thrust washer can withstand the high thrust loads which otherwise might require a high capacity tapered roller thrust bearing.
As noted above, variable displacement of a hydraulic machine in accordance with the present invention may be effected by moving a valve plate relative to a cam such that transitions between high and low pressure ports occur at piston positions other than top dead center (TDC) and bottom dead center (BDC). The particular point in the piston stroke chosen for the transition to occur will depend at least on the amount of displacement reduction desired. It may also be chosen based on considerations of the compression and expansion of the fluid within the cylinders.
In embodiments of the present invention acting as a pump, a cam forces a piston/roller assembly to move outward against a pressure until it reaches TDC, at which time the piston motion typically stops, the high pressure port closes, and the low pressure port opens. Whatever portion of the fluid is still in the cylinder or attaching fluid passages undergoes expansion as the pressure decreases. Therefore, the first increment of downward stroke accommodates this expansion before it begins to ingest new low pressure fluid from the supply. At the bottom of the stroke, the opposite occurs. As the piston begins the upward stroke, the first increment of travel is used to compress the new fluid before the high pressure fluid is exhausted to the receiver. This compression and expansion can contribute negatively to the volumetric efficiency of the pump, and for that reason, the clearance volume and passage volume are kept to a minimum.
In embodiments of the invention where the hydraulic machine is being operated as a pump, and where it is desired to have the transition of pressures occur at positions other than TDC and BDC to incrementally decrease the throughput of the pump, and thus its input torque, it may be desirable to have the port switching occur after TDC and BDC, rather than before. At TDC, as the piston motion reverses, fluid flow reverses whether the port is connected to the high pressure port or the low pressure port. If the piston is in downward motion after TDC at the time the pressure switches from high to low, the momentary time that the total port area is zero or very low, and the natural expansion of the fluid partially compensates for the restriction in inbound flow caused by the port switching. At BDC, the piston motion again reverses with a consequential reversal in flow direction. If the piston is in an upward motion after BDC at the time the pressure switches from low to high, the momentary time that the total port area is zero or very low, and the natural compression of the fluid partially compensates for the restriction in outbound flow caused by the port switching.
In embodiments of the invention where the hydraulic machine is being operated as a motor, and it is desired to have the transition of pressures at positions other than at top and bottom dead centers to incrementally decrease the throughput of the motor, and thus its output torque, it may be desirable to have the port switching occur before TDC and BDC, rather than after. At TDC, as the piston motion reverses, fluid flow reverses whether the port is connected to the low pressure port or the high pressure port. If the piston is in upward motion before TDC at the time the pressure switches from low to high, the momentary time that the total port area is zero or very low, and the natural compression of the fluid partially compensates for the restriction in outbound flow caused by the port switching. At BDC, the piston motion again reverses with a consequential reversal in flow direction. If the piston is in a downward motion before BDC at the time the pressure switches from high to low, the momentary time that the total port area is zero or very low, and the natural expansion of the fluid partially compensates for the restriction in inbound flow caused by the port switching.
It is worth noting that indexing the valve plate relative to the cam is used not only for changing the displacement of the hydraulic machine, but also to effect a change from pump operation to motor operation. The sequence of port indexing starts with the opening of the high pressure port with the piston/roller assembly at BDC for full pump displacement. As the valve plate, or manifold, is indexed in the direction of rotation (after BDC), the pump displacement is decreased incrementally. There is a limit to the amount of modulation that can be accommodated before the piston velocity and resulting flow velocity exceed the restricted flow capacity at the port opening. Indexing beyond this limit is a non-operating region.
Continuing to index in the same direction, but beyond the non-operating region, reaches the modulated motor position (before TDC) and then proceeds to the full displacement motor position at TDC. In other words, the modulated pump index position and the modulated motor index position can both lie between the full displacement index position of the pump and the full displacement index position of the motor. For example, for an eight lobe cam, the total index travel from BDC to TDC is 22.5 degrees from full displacement pump to full displacement motor. For a nine lobe cam, the total index travel is 20.0 degrees from full displacement pump to full displacement motor. The non-operating region consists of approximately the middle half of the respective index travel.
As discussed above, making the transition from high to low pressure at certain points in the piston stroke may take advantage of the compression and expansion of the fluid, thereby helping to reduce the occurrence of hydraulic lock. Another way to help inhibit hydraulic lock is to provide some overlap between the high and low pressure ports as they close and open into a particular cylinder. This can help to avoid hydraulic lock, where both ports are closed while the piston is in motion. Although a design utilizing an overlap of ports could eliminate the potential for hydraulic lock when the transition occurs with the pistons in motion, such an overlap could lead to a “short circuit” for the hydraulic fluid. That is, some of the fluid leaving the high pressure port could travel directly to the low pressure port, rather than entering the cylinder and working on, or being worked on by, the piston. This would lead to a decrease in the volumetric efficiency of the hydraulic machine.
Embodiments of the present invention overcome the issues associated with a hydraulic short circuit, while still providing some overlap between ports during a transition. For example, rather than a single channel in the cylinder block connecting a cylinder with the high and low pressure ports—alternated by the position of the valve plate—a split channel can be used that will provide a barrier between the high and low pressure ports when the overlap occurs. In such a configuration, the high pressure port is still connected to the low pressure port during overlap, but this connection does not take place at the channel opening, where there may be only a few millimeters between the two ports. Rather, the connection between the two ports is in the cylinder. This means that fluid leaving the high pressure port during overlap could theoretically flow to the low pressure port, but to do so, it must first flow through one channel to the cylinder and then out of the cylinder through the second channel to eventually reach the low pressure port.
In addition to increasing the distance necessary for fluid transfer between the ports, the configuration described above further inhibits the fluid transfer because inertia of fluid residing in the channels will need to be overcome. This inertia is not available when a single channel is used and the two ports have a direct connection that does not require flow through the channel for the transfer to occur. To the extent that a hydraulic lock does occur, embodiments of the present invention provide a relief valve in one or more of the cylinders to allow a path for fluid to be exhausted to the case of the hydraulic machine so that the entire system will not lock. Providing a relief valve in the cylinders puts the safety feature where it is most needed, rather than at a remote location in the system, far away from the point at which the lockup actually occurs. In addition, embodiments of the invention allow each cylinder to have its own relief valve, further reducing the likelihood of even a momentary hydraulic lock.
The energy recovery system 12 illustrated and described herein is just one use for a hydraulic machine arrangement in accordance with the present invention. It is understood that such hydraulic machine arrangements may be used for other applications—e.g., they may be used exclusively as motors to provide torque, or exclusively as pumps to provide pressurized fluid. In addition, hydraulic machine arrangements, such as the hydraulic machine arrangement 14, may be mounted in different locations on a vehicle, for example, on drive shaft 29, the transfer case 19, or half axle shafts 31, 33, illustrated in
The energy recovery system 12 also includes a second, or low pressure accumulator 30. The low pressure accumulator 30 provides a charge pressure—i.e., a relatively low pressure—to the hydraulic machine arrangement 14 to help ensure that there is always some liquid supplied to the hydraulic machine arrangement 14, thereby avoiding cavitation. The low pressure accumulator 30 may include two parts: a liquid/gas container 32, and a gas only container 34. Similarly, the high pressure accumulator 28 may include two parts: a liquid/gas container 36, and a gas only container 38. Configuring each of the accumulators 28, 30 with two containers facilitates packaging by reducing the size of each liquid/gas container 32, 36. Of course, high and low pressure accumulators, such as the high and low pressure accumulators 28, 30, may include a single liquid/gas container, rather than the two-part configuration shown in
The energy recovery system 12 also includes a control system, shown in
When the control module 40 is signaled to use regenerative braking during a braking event, it sends a control pressure to the hydraulic machine arrangement 14 to ensure that the hydraulic machine arrangement 14 operates as a pump. Conversely, when the control module 40 is signaled to provide torque to the wheels 20, 22 during a driving event, it sends a control pressure to the hydraulic machine arrangement 14 to ensure that the hydraulic machine arrangement 14 operates as a motor. In this mode, fluid from the high pressure accumulator 28 drives the hydraulic machine arrangement 14 such that torque is provided to the wheels 20, 22.
In another operating scenario, the energy recovery system 12 can be used to store energy when driving the vehicle. High powered internal combustion (IC) engines can be inefficient when operating below approximately 70% of full torque, and efficiency continues to decrease as the torque decreases further. For modern vehicles, highway driving typically operates the engine at 12% to 30% of full torque. Using a hydraulic energy recovery system, such as the system 12, the IC engine can be operated intermittently, within the operating speeds of the hydraulic machinery, at near full torque while storing the excess energy in the high pressure accumulator. When a control system, such as the control module 40, detects that the accumulator is near its maximum pressure, the IC engine is idled, and cylinders are deactivated or shut off, while the vehicle is driven from the stored energy. When the high pressure accumulator is depleted, the control system reactivates the IC engine, and the cycle starts over again. With refined controls, the cycling can become transparent to the vehicle driver.
The hydraulic machine 42 also includes a cam 58 having an aperture 60 configured to allow the shaft 27 to pass therethrough. Although the cam 58, like the other cams illustrated and described herein, are disposed inboard the pistons 48, 54, embodiments of the present invention may have cams that are radially outboard of a cylinder block. The shaft 27 turns the cylinder block 50, while the cam 58 is stationary. Riding on the cam 58 are cam followers 62, which cooperate with the pistons 48, 54 to operate the pistons 48, 54 to pump fluid to the hydraulic machine 42 when it is operating as a pump. Conversely, when the hydraulic machine 42 is operating as a motor, it receives high pressure fluid from the accumulator 28, and outputs torque to the shaft 27.
Returning to
In order to facilitate a connection between the cylinders 52, 56 and the high and low pressure fluid ports 64, 66, the hydraulic machine 42 includes a valve plate 70. The valve plate 70 also remains relatively stationary, like the cam 58, while the cylinder block 50 rotates with the shaft 27. The port housing 68 and the outer housing 69 are also stationary. It is worth noting that in other embodiments, a cam and valve plate, such as the cam 58 and the valve plate 70, may be configured to rotate with a shaft, such as the shaft 27, while a respective cylinder block is stationary. In either case, the valve plate 70 is movable relative to the cam 58, which allows the hydraulic machine 42 to switch from a pump to a motor, and vice versa.
When the hydraulic machine 42 is operating as a pump, cylinders 52, 56 will be connected to the high pressure fluid port 64 when a corresponding piston 48, 54 is in an outstroke. Conversely, when the pistons 48, 54 are in an instroke, their respective cylinders 52, 56 will be connected to the low pressure fluid port 66. In order to change the operation of the hydraulic machine 42 from a pump to a motor, the valve plate 70 is rotated relative to the cam 58, such that the fluid connections to the cylinders 52, 56 are reversed. Specifically, when the hydraulic machine 42 is operating as a motor, the cylinders 52, 56 will be connected to the high pressure fluid port 64 when their respective pistons 48, 54 are in an instroke, and they will be connected to the low pressure fluid port 66 when their respective pistons 48, 54 are in an outstroke.
In order to effect movement of the valve plate 70 relative to the cam 58, the hydraulic machine 42 includes an axial piston 72. The piston 72 drives the valve plate 70 via a link (not shown) attached to the valve plate 70 and riding in a slot 74 disposed in the shaft 27. The movement of the link in the slot 74 translates the linear movement of the axial piston 72 into rotational movement of the valve plate 70. Movement of the axial piston 72 in one direction is effected by fluid entering a mode port 76 located in the port housing 68. A spring (not shown) is provided to return the axial piston 72 to its previous position when the fluid pressure from the mode port 76 is exhausted. In other embodiments, other actuators, such as a tangential piston 77—shown in phantom in FIG. 5—may be used in place of the axial piston 72 to control rotation of the valve plate 70.
In order to the facilitate a connection between the high and low pressure ports 64, 66 and the cylinders 52, 56, the valve plate 70 includes a number of apertures or ports 78, 80, 82, 84, 86, 88, 90, 92—see
As shown in
Continuing to use the example of the hydraulic machine 42 operating as a motor, as its components are shown in
To facilitate an increase in speed of the hydraulic machine 42 as its components are shown in
To complete the example,
When the spool/poppet valve 104 is moved to a position such that the valve ports 82, 90 are connected full time to the low pressure port 66, the hydraulic machine 42 will operate at 38.2% of its full displacement. Similarly, when the spool/poppet valve 102 is moved to a position such that the partial stroke valve ports 78, 86 are connected full-time to the low pressure port 66, and the spool/poppet valve 104 is positioned to connect the full stroke valve ports 82, 90 to the high pressure port 64, the hydraulic machine 42 will operate at 61.8% of its full displacement.
It is worth noting that two of the full-time low pressure valve plate ports 80, 88 are of substantially equal size. Conversely, the low pressure valve plate port 84 is shorter than the ports 80, 88, and the low pressure valve plate port 92 is longer than the low pressure ports 80, 88. As described above, the change from high pressure to low pressure can be made to occur so that all of the cylinders do not experience this change simultaneously. Offsets in the port spacing correspond to offsets in their respective cam lobes, and result in spacing “events” occurring individually. Although the port lengths differ, the space between them is generally uniform, thus ensuring that at least one of them will always be in communication with at least one of the cylinders 52, 56, thereby avoiding a “hydraulic lock” effect.
Although
Although the tapered roller bearing 106 may provide an acceptable mechanism for supporting the cylinder block 50, an alternative is also shown in
As shown in
The position of the balance piston 110 is self-regulating. If leakage in the hydraulic machine 42 increases, the separating pressure drops, and the piston 110 moves to decrease the operating gap. Conversely, if the leakage in the hydraulic machine 42 decreases, the separating pressure increases, and the piston 110 moves to increase the operating gap. The design of the orifice 114 is adjusted to minimize the loss due to high pressure fluid leakage while maintaining a film of fluid between the rotating cylinder block 50 and the stationary balance piston 110. Also shown in
The hydraulic machine 120 also includes a 9 lobe cam 128, which actuates, or is actuated by, the pistons 124 inside a cylinder block 130. Similar to the hydraulic machine 42 described above, the cylinder block 130 rotates with a shaft 132, while the cam 128 remains stationary. The hydraulic machine 120 also includes a valve plate 133, which contains three low pressure ports A, C and E, and three high pressure ports B, D and F. Although an axial piston arrangement such as described above for the hydraulic machine 42 shown in
As described above, the hydraulic machine 42 illustrated in
As shown in
1. 3 shallow lobes 144=16.1%
2. 3 intermediate lobes 142=25.8% (
3. 3 shallow lobes 144 and 3 intermediate lobes 142=41.9%
4. 3 deep lobes 140=58.1%
5. 3 shallow lobes 144 and 3 deep lobes 140=74.2%
6. 3 intermediate lobes 142 and 3 deep lobes 140=83.9%
7. All 9 lobes 140, 142, 144=100% (
In an alternative cam design, where the 3 deep lobes produce less displacement than the sum of the 3 shallow and 3 intermediate lobes, numbers 3 and 4 in the above list would be reversed to give a smooth progression. In each of the seven cases listed above, it is assumed that each of the pistons 124 not in contact with a respective lobe 140, 142, 144 will be disengaged by one of a number of mechanisms contemplated by the present invention. For example, as described above, the pressure inside the cylinder block 130 can be increased to be approximately equal to the low pressure fluid, for example, as provided by the low pressure accumulator 30 shown in
With regard to the hydraulic machine 120 shown in
With a control valve controlling each port, any combination of ports can be exhausted when not required for a desired displacement. For example, if the minimum pump displacement is desired, the A, B, C, and D are connected to exhaust to deactivate those cam lobes. If minimum motor displacement is desired, then B, C, D, and E are connected to exhaust. Partial displacement, and in particular, almost maximum displacement, would exhaust A and F. Because of this indexing, no two ports are paired in the same way for pump and motor operation. Therefore, six two-way poppet valves can be used to control the displacement for all conditions of pump and motor operation. The following chart shows the passage connections with the six ports for both pump and motor modes.
As described above, another way to disengage pistons, such as the pistons 124, to effect variable displacement is to disengage the non-driving pistons by decreasing the return pressure to near zero by using a high capacity pump, such as a jet pump.
In
As described above, a jet pump arrangement, such as the jet pump arrangement 153, can also be used to disengage certain cylinders when a hydraulic machine is operating as a motor. Again using
In this way respective pistons 124 are disengaged, being subject to the centrifugal force of the rotating valve plate 133. The hydraulic machine 120 operates at less than full displacement, thereby allowing the hydraulic machine 120 to operate at an increased speed. With a fixed displacement hydraulic machine, the speed is limited by the maximum amount of fluid flow through the machine. This has limited applications, for example, slow speed, off road vehicles. In contrast, embodiments of the present invention provide hydraulic machines having variable displacements effected by disengaging some of the pistons, thereby making them suitable for high speed, on highway vehicles.
One issue that may need to be addressed with regard to the function of a radial piston hydraulic machine, is the output of an undesirably low torque when the machine is initially started. This can be a result of friction between a cam follower, and a piston head. One possible solution to this is to use the dual piston head configuration shown in
As shown in
Each of the cylinder blocks 186, 188 is attached to a respective shaft 202, 204, which may be, for example, axle shafts, such as the half axle shafts 31, 33 illustrated in
As described in detail above, embodiments of the present invention provide a hydraulic machine wherein certain pistons registering with certain cam lobes can be disengaged, thereby providing a variable displacement machine. This type of variable displacement is discrete—i.e., pistons must be engaged or disengaged, a binary function, and there are limits based on the particular configuration being used on which cam lobes and their associated pistons can be disengaged and which must be engaged. Therefore, such a machine can be operated at a discrete number of reduced displacement levels, but cannot be operated at displacements between these discretely defined levels. To address this, embodiments of the present invention also provide a hydraulic machine having a “continuously variable displacement”. There are of course limits to the variation in displacement that can be achieved, so the term “continuously variable displacement” is used to distinguish the type of displacement from the discrete variable displacement described above, and not to necessarily imply that there are unlimited different displacements available.
One way in which a hydraulic machine can be configured for continuously variable displacement is to configure it to switch the high and low pressure ports for a cylinder when the piston is at some position in its stroke other than TDC or BDC. To the extent that there is no overlap in the switching of the ports—i.e., one port completely closes before the other opens—there is the potential for hydraulic lock to occur.
In contrast,
As shown in
One way to address the short circuit issue is to provide a barrier between the high and low pressure ports at the opening of the channel in the cylinder block. Such a configuration is shown in
In addition to providing overlap between the high and low pressure ports during crossover, another way to help ensure that hydraulic lock is not encountered is by using one or more relief valves in individual piston cylinders.
Relief valves, such as the relief valve 248, can be disposed in some or all of the cylinders in a cylinder block of a hydraulic machine in accordance with embodiments of the invention. In one embodiment, for example, the nominal pressure of the working fluid in the cylinder may be approximately 5000 pounds per square inch (psi), and the size of the channel 250, and the spring constant of the spring 254 configured to allow fluid to be expelled when its pressure reaches 5500 psi. Placing the relief valves in the cylinder heads provides advantages over systems having relief valves at other locations in the system. For example, the type of hydraulic lock described above—i.e., caused by the oncoming and offgoing ports being closed simultaneously while a piston is in motion—occurs in the cylinder. A relief valve elsewhere in the system would not be in fluid communication with the cylinder because the closed ports would be between it and the cylinder. Therefore, this type of relief valve, outside of the cylinder, would be ineffective for this type of hydraulic lock.
While embodiments of the invention have been illustrated and described, it is not intended that these embodiments illustrate and describe all possible forms of the invention. Rather, the words used in the specification are words of description rather than limitation, and it is understood that various changes may be made without departing from the spirit and scope of the invention.
This application is a continuation in part of international Patent Application No. PCT/US2008/053747 filed 12 Feb. 2008, which claims the benefit of U.S. provisional application Ser. No. 60/900,775 filed 12 Feb. 2007, and U.S. provisional application Ser. No. 60/921,279 filed 2 Apr. 2007, each of which is hereby incorporated herein by reference. This application claims the benefit of U.S. provisional application Ser. No. 61/128,055 filed 19 May 2008, which is hereby incorporated herein by reference.
| Number | Date | Country | |
|---|---|---|---|
| 61128055 | May 2008 | US | |
| 60900775 | Feb 2007 | US | |
| 60921279 | Apr 2007 | US |
| Number | Date | Country | |
|---|---|---|---|
| Parent | PCT/US2008/053747 | Feb 2008 | US |
| Child | 12468450 | US |
