Radial hydraulic motor for a hydraulic hybrid vehicle

Abstract

A radial hydraulic pump/motor for use in place of a mechanical transmission for a hydraulic hybrid vehicle is described. Means for disengaging and reengaging individual working pistons are provided, allowing partial or zero effective displacement. Methods for using the radial pump/motor to supplement power from other pump/motors for the vehicle are also described.

Description

BACKGROUND OF THE INVENTION

The disclosed embodiments are directed generally to hydraulic hybrid vehicles, and more specifically to transmissions for such vehicles.

DESCRIPTION OF THE RELATED ART

The drivetrain of a motor vehicle typically must provide a greater amount of torque to the drive wheels during launch and acceleration than at any other time. The traditional method for meeting this peak torque is by providing a multi-speed mechanical transmission and selecting the lowest gear during launch. The lowest gear, having the largest gear ratio, provides the maximum torque multiplication. It also allows the engine to operate at a sufficient speed to generate the necessary power, despite the low rotational speed of the drivetrain at launch.

Some hybrid vehicles employ an electric or hydraulic motor for motive power. These may not require a multi-speed transmission because, unlike an internal combustion engine, they can deliver high torque even at zero speed. Still, even an electric or hydraulic motor has a limit on the amount of torque it can efficiently and smoothly deliver while still being sized appropriately to the rest of the duty cycle. A transmission is therefore often desirable for efficiency and sizing reasons, even for these vehicles.

A multi-speed mechanical transmission adds cost, weight, and bulk. An alternative would be to provide the additional torque needed at launch by supplementing the torque provided by the primary motor. This could be done by adding a supplementary motor that only operates during launch, being disengaged at other times.

Using a supplementary motor instead of a transmission would dispense with the need for frequent shifting of gears. The output torque can follow the constant power curve with no torque interruptions or at most one interruption occurring when the supplementary motor is disengaged, rather than the several interruptions caused by shifts among the gears of a transmission. This is particularly advantageous for passenger-carrying applications where smooth acceleration is desirable for passenger comfort and safety. Also, a compact supplementary motor may be smaller, lighter, and easier to package in a vehicle than a bulky multi-speed transmission.

Eliminating the need for gear shifts may also provide for faster acceleration. In many applications, acceleration rate is important to productivity. For example, delivery vehicles and refuse collection trucks perform many starts and stops in succession. Faster acceleration between stops can increase the number of addresses served in a given amount of time.

However, use of a supplementary motor in place of a transmission presents several problems. First, because it is needed only during launch, it has little purpose at other times and may introduce friction, drag, or inertial effects if it remains engaged. Second, to disengage it suggests the need for a clutching mechanism, adding complexity and cost. Third, the amount of supplementary torque necessary to substitute for the torque multiplication of a transmission may be several times the torque rating of the primary motor, suggesting that the supplementary motor would have to be even larger than the primary motor. As a result, most powertrains continue to employ multi-speed mechanical transmissions.

Series hydraulic hybrid powertrains are increasingly being applied to heavy duty vehicles in order to improve fuel efficiency at a low cost. These powertrains employ one or more hydraulic pump/motors to power the drivetrain, often through a multi-speed mechanical transmission. Here the most logical alternative to a transmission would be a supplementary hydraulic pump/motor, which could provide supplementary torque for acceleration as well as supplemental regenerative braking capacity for very aggressive braking events.

For hydraulic powertrains, certain advantages make replacing the transmission with a motor more attractive than with other powertrains. First, some types of hydraulic pump/motors can be set to zero displacement with minimal parasitic drag, avoiding the need for a clutch by allowing the device to rotate relatively freely with the drivetrain even when not in use. Second, hydraulic pump/motors have a high torque density so that a large amount of supplementary torque can be provided in a smaller space than that required by a mechanical transmission. Third, there are numerous varieties of hydraulic pump/motor design, varying significantly in torque, speed, and power characteristics, making it more likely that a compact but powerful design can be found that would provide enough torque to substitute for torque multiplication. Finally, use of a transmission with a hydraulic pump/motor is complicated by the need to assure that the pump/motor is always at zero displacement when the transmission passes through neutral, to prevent the pump/motor speed from running away and destroying the pump/motor. A supplementary hydraulic pump/motor could replace the transmission and avoid this problem, leading to a simpler control strategy.

OBJECT OF THE INVENTION

It is an object of the invention to eliminate the need for a multi-speed mechanical transmission in a hydraulic hybrid vehicle by providing supplementary torque at vehicle launch, rather than the torque multiplication of a multi-speed transmission.

It is another object of the invention to provide for disengagement of the supplementary torque when not needed, without need for a clutch, and with minimum drag imposed on the powertrain.

SUMMARY OF THE INVENTION

Applicant's co-pending application entitled “Modular Hydraulic Hybrid Drivetrain”, 13/415,109, filed Mar. 8, 2012, describes the use of a through-shaft swash plate pump/motor to replace a multi-speed transmission. A radial pump/motor, if properly modified in its structure and operation, affords certain benefits over a swash plate design, and therefore the current disclosure relates more specifically to the adaptation and use of a unique radial pump/motor for this purpose.

In the invention, a radial hydraulic pump/motor is employed as part of a series hydraulic hybrid powertrain to provide supplementary torque at launch, while having the ability to effectively disengage after launch. The radial pump/motor employs a new and unique piston deactivation method in order to effectively disengage when not needed. A high-speed control valve design enables engagement and disengagement of individual pistons.

A typical radial pump/motor has a plurality of pistons that each reciprocate in respective fixed cylinder bores. The cylinder bores are arranged in a radial fashion, each bore extending outwardly from the center of the device along its own radial line. Each piston has a connecting rod that is essentially a cam follower to a central eccentric cam that rotates with the pump/motor output shaft about the center of the device. The eccentric cam, and hence the power output shaft, is rotated by a connecting rod force that results from fluid pressure driving a piston toward the center of the device. The cam, being eccentric to the shaft, has a minimum radius and a maximum radius about the center of rotation of the shaft.

Fluid intake and exhaust at each cylinder occurs through fixed ports and is controlled mechanically by a timing cam or similar means, in a power stroke of a piston, high pressure fluid is admitted to the space above the piston when near its top dead center (TDC) position, and force is transmitted from the piston through the connecting rod to the eccentric cam and thereby to the pump/motor shaft. Fluid then exits to low pressure as the piston returns toward TDC. Prior art radial pump/motors are limited in their maximum shaft speed by the need for the pistons to reciprocate in their bores as the shaft and cam rotate, and the need for fluid to enter and exit without throttling.

In the radial pump/motor of the invention, the pump/motor shaft is a through-shaft that forms part of a vehicle powertrain. A high-speed control valve is provided at each cylinder to control the timing of intake and exhaust of working fluid through the cylinder. When the powertrain requires torque from the radial pump/motor, the working pistons are in their engaged position, bearing upon the eccentric cam to produce output power in the way normally associated with a radial pump/motor. When torque is not needed, the pistons withdraw from the cam while the cam and shaft continue to rotate with the powertrain. The radial pump/motor thereby “freewheels” without reciprocating the pistons, making it capable of reaching a much higher speed than if the pistons were riding along, and imposing very little drag on the powertrain.

Each piston is withdrawn from the cam in an orderly and sequential fashion, by switching it to low pressure and allowing it to remain above the TDC position such that the bearing surface of the respective connecting rod is outside the maximum radius of the cam, and thereby not within reach of the cam surface. While disengaged, a protruding member such as a pin or boss retains the connecting rod in the proper orientation for its later re-engagement with the cam. To re-engage, each connecting rod is in turn sequentially reseated on the cam by switching the respective cylinder back to a higher pressure at the appropriate time.

The basic method of deactivation is as follows. Initially, the radial pump/motor is operating normally transmitting power. When the speed reaches a first maximum speed, the cylinders are deactivated by first actuating all the high-speed control valves that connect all piston volumes to the low pressure supply. Next the high pressure supply is closed with a shutoff valve. A pressure-relieving valve on the high pressure line then dumps the trapped high pressure to tank or atmospheric pressure. The high-speed control valves are actuated to bring all radial pump/motor pressures to atmospheric. The continued rotation of the cam then pushes the pistons to a position just above TDC. Because pressure has been removed, the pistons remain in this position, held by friction between the piston rings and bore, and so rest outside the eccentric orbit of rotation of the cam and therefore will not contact the cam even as it continues to rotate in a preferred embodiment a mechanical solenoid pin, a permanent magnet, or similar locking means is additionally used to retain the piston in the disengaged position.

To re-engage the pistons and return to a motoring or pumping state, the pistons are sequentially reseated on the cam. This is done by first opening all high-speed control valves to low pressure, and for each individual piston, sequentially actuating the respective high-speed control valve to provide a seating force to the respective piston just as the cam reaches what would be the TDC position for that piston.

A method of operation is also described in which the radial pump/motor serves as a transmission for a hydraulic hybrid vehicle. The radial pump/motor is teamed with a primary axial piston pump/motor (preferably an over-center bent axis pump/motor). The radial pump/motor preferably operates either at full displacement or is disengaged by piston deactivation. The axial piston pump/motor either operates alone or supplements the radial pump/motor by effectively adding or subtracting displacement to match the net displacement demanded by the vehicle.

In prior art, it is known, for example in WIPO Patent Applications WO 2004/025122 and WO 2006/109079 (Artemis intelligent Power Ltd) and U.S. Pat. No. 7,793,496 (Rampen et al.) to employ high-speed control valves at each cylinder in order to provide for varying the time averaged effective displacement of the pump/motor by controlling the duration of the hydraulically powered stroke of each working piston. This may be done by opening the cylinder to the high pressure or low pressure manifold at times other than top dead center or bottom dead center so that only part of the fixed distance of piston reciprocation is powered or transmits power. In addition, any cylinder may be placed in an idling state in which no power is transmitted by keeping it open to the low pressure manifold, in which case the piston continues to reciprocate within the cylinder but low pressure fluid flows in and out with the reciprocation, producing no work. In contrast with this known approach, the current invention provides for an idling mode by causing the piston and connecting rod to temporarily cease reciprocation within the cylinder by temporarily bringing it out of contact with the eccentric cam. This prevents fluid from circulating in the cylinder at all during the idling mode, leading to a lower potential for energy losses during idling.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 shows a perspective view of a radial pump/motor according to the invention.

FIG. 2 is a sectional view of the radial pump/motor showing the radial arrangement of the piston/cylinder assemblies, the high-speed control valves, and the eccentric cam.

FIG. 3 is a close up sectional view detailing a single piston/cylinder assembly and high-speed control valve.

FIG. 4 is a sectional view orthogonal to that of FIG. 2 and showing a first preferred locking means.

FIG. 4A is a detail corresponding to a portion of the view of FIG. 4 but detailing a second preferred locking means.

FIG. 5 is a sectional view orthogonal to that of FIG. 4, detailing the orientation member and orientation groove that receives the member when a piston is retracted.

FIG. 6 further details key features of the orientation groove.

FIG. 7 is a perspective view of the interior of the radial pump/motor case showing arrangement of the machined orientation grooves.

FIG. 8 is a close up perspective view of a single orientation groove.

FIG. 9 is a schematic of a hydraulic hybrid vehicle showing a preferred application of the invention.

FIG. 10 is a flowchart showing the primary steps in a preferred method of use of the invention. FIG. 10A is a flowchart showing the detailed steps for step 1003 of FIG. 10.

FIGS. 11-16 are connected flowcharts further detailing the primary steps in a preferred method of use of the invention, with FIGS. 12-16 detailing steps in FIGS. 11-12 where indicated.

DETAILED DESCRIPTION

In FIG. 1, a radial hydraulic pump/motor 100 is depicted. In this example, case 120 encloses a total of five working piston/cylinder assemblies (seen more clearly in FIG. 2 as for example cylinder volume 110a and piston 116a) that are driven by fluid pressure to rotate shaft 101. Alternatively, as is known to apply to conventional radial pump/motors, there may be more or less than the five assemblies here depicted. Shaft 101 is a unitary through-shaft having ends 101a,b which extend from the front and rear of the device respectively. For each piston/cylinder assembly (e.g. 110a) there is a corresponding high-speed control valve (e.g. 111a) that controls the timing of the entry and exit of fluid into the respective working cylinder.

FIG. 2 shows a cross section of radial pump/motor 100. Here it can be seen that shaft 101 includes an eccentric cam 103 upon which bear the connecting rods 117a-e. Anyone skilled in the art will be familiar with the way in which connecting rods 117a-e transmit force to the eccentric cam 103 and thereby cause shaft 101 to rotate, producing power.

Each cylinder volume 110a-e is connected to a corresponding high-speed, 2-position, 3-way control valve 111a-e through fluid passages 113a-e which are preferably cast into case 120 of the pump/motor. Each valve 111a-e provides for connection of its respective piston volume with high and low pressure supplies (not shown) that supply the radial hydraulic pump/motor with working fluid.

Fluid flow through the radial pump/motor is entirely controlled by the high-speed control valves 111a-e, which are actuated rapidly by command of an electronic control module (not shown). In this preferred embodiment, each high-speed control valve is a piloted 2-stage, pressure actuated, 2-position/3-way (2P3W) spool. The main spool is intended to be large enough to minimize pressure drop (for example, preferably less than a 10 bar pressure drop when flowing 125 liters per minute). In addition to the low pressure drop through the valve, the speed of each valve must be very fast to prevent throttling during switching events, preferably with a response time on the order of 1 ms for either high-to-low or low-to-high pressure.

Each valve 111a-e preferably includes a respective pilot valve 121a-c which ensures fast actuation of a respective main spool 122a-e. Pilot valve 121a-e opens a very small volume to high pressure which in turn applies an actuating force to main spool 122a-e. Because the actuating volume is small the main stage actuating force rises very rapidly. When pilot valve 121a-e is switched in the opposite direction, respective spring 123a returns main spool 122a-e to the original position. A central drain (not shown) exhausts flow from the pump case at a lower pressure, for example, atmospheric pressure.

FIG. 3 shows a detail of a single piston/cylinder assembly and high-speed control valve. When high-speed control valve 111a is open to high pressure, high pressure fluid is admitted from the high pressure source (not shown) into passage 113a and into the space 110a above piston 116a. This exerts a force on piston 116a which then travels through connecting rod 117a. Because the vector of this force is offset from the center of shaft 101 about which cam 103 rotates, it causes cam 103 to rotate which thereby rotates shaft 101. When piston 115a has completed its downward travel to bottom dead center (BDC), continued rotation of cam 103 will cause it to be forced upward again. At this time, valve 111a quickly switches to close communication with the high pressure source and open a path to the low pressure source. Fluid trapped in cylinder 110a will then be free to exit cylinder 110a to the low pressure source as the upward movement of the piston 116a sweeps the volume.

Referring again to FIG. 2, the collective behavior of the five piston/cylinder assemblies can now be more clearly understood.

The piston deactivation sequence for complete deactivation of radial pump/rumor 100 operates as follows. First, all five pistons 115a-e and cylinders 116a-e are switched to low pressure by commanding respective high-speed control valves 111a-e to a low pressure position that places the respective cylinders in communication with the low pressure source. During this stage, low pressure fluid will be cycled into and out of the piston volume, but no work will be done. After each piston is connected to low pressure, the high pressure supply is closed and a small pressure-relieving valve on the high pressure line opens to atmosphere. Any compressed fluid will then be drained to a sink, such as (preferably) to a deaeration tank, or to a to pressure source.

After the high pressure supply line is relieved to atmospheric pressure, in the preferred embodiment each high-speed control valve 111a-e is then switched to the position that opens to the high pressure line (now at atmospheric pressure). This occurs sequentially for each respective piston/cylinder assembly when the cam is at top dead center (TDC) for each respective assembly. As each piston passes through the TDC position, there is no pressure differential across the piston and it will be retained at or just above the TDC position due to seal friction. At this point, an optional low pressure supply shutoff valve can be closed, isolating the radial pump from the hydraulic system. The cam 103 will then freewheel with the rotation of the drivetrain (i.e. shaft 101), imposing little drag.

Optionally and preferably, a locking means positively prevents each deactivated piston from reseating on the cam surface 104 should seal friction be insufficient to retain it. Various types of locking means are possible, such as a locking pin, magnetic force, fluid pressure force, or frictional force.

A first preferred locking mechanism utilizing a solenoid pin is detailed in FIG. 4. Solenoid 310 includes pin 311. Preferably, pin 311 is retracted when solenoid 310 is energized and is extended by spring force when not energized. When solenoid 310 is energized (as depicted), pin 311 is retracted. When de-energized, pin 311 extends into cylinder 110 at a point below the bottom of piston 116, preventing piston 116 from lowering enough to re-engage the cam surface.

A second preferred locking mechanism utilizes a fluid pressure acting on a stepped piston and bore, as depicted in FIG. 4A. Each cylinder is provided with a respective passage containing pressurized fluid, and leading into a small chamber created by small steps placed in the cylinder bore and the piston. Looking at the piston and cylinder of FIG. 4A, fluid pressure is maintained in passage 601, which enters the cylinder bore wall 602. The outer surface 603 of piston 116 is provided with a preferably annular stepped feature 605. Cylinder bore wall 602 is provided with a corresponding preferably annular stepped feature 606. The two stepped features, along with the adjacent portions of the sides of the bore and piston, form a small preferably annular chamber 607. The fluid pressure present in chamber 607 presents force against the surfaces of both stepped features 605 and 606. Because the bore is fixed but the piston can move, a resultant force tends to urge feature 605 and piston 116 in a direction toward TDC. At times when the piston is inactive and therefore disengaged with the cam surface, the retaining force acts to keep the piston away from the cam surface, even if the device is jarred by road forces. When the piston 116 is active and therefore engaged with the cam surface, the retaining force is overcome by the force of the working fluid driving the piston, and the piston transmits power to the cam as normally, although diminished somewhat by the retaining force.

Because the fluid pressure is always present in passage 601 and chamber 607, the retaining force is always present, even when the piston is engaged with the cam surface, i.e. not deactivated. However, the size of the retaining force is not particularly critical to the efficiency of the device, because the energy used to oppose the retaining force when the piston is in an active state is recovered on the return stroke. Therefore the size of the retaining force may be selected to that necessary to retain the piston reliably against inertial and road forces, by simply selecting the fluid pressure to be present in passages 601 and chamber 607, and the area of the stepped features 605 and 606.

The fluid pressure in passage 601 and chamber 607 (which can also be referred to as a retaining pressure) may be any pressure sufficient to provide a desired retention force resulting from the fluid pressure reacting against the area of the stepped features. For example, a relatively high retaining pressure such as 2000 to 3500 psi may be indicated if it is desired to have a small step area, or a much lower retaining pressure near or well below that of the low pressure reservoir if it is permissible to have a larger step area. Therefore the retaining pressure may be chosen to be whatever pressure is sufficient to maintain a desired degree of retention force for a given step area. In the design of a device according to the invention, the choice of retaining pressure may be influenced as much by the proximity of a suitable fluid pressure passage or source (such as a regulated pressure provided for other purposes, or a low pressure passage leading to a low pressure reservoir or deaeration device, or any similar pre-existing fluid passage) as by anything else. If a low retaining pressure is desired, it is anticipated that a retaining pressure of not more than half the low pressure reservoir pressure, or even much less, would suffice with an appropriately sized step area.

Referring again to FIG. 2, the reactivation sequence is as follows. Reactivating the pistons 116a-e is accomplished by reseating them on the rotating surface of cam 103 in a sequential manner similar to the deactivation sequence. First, the low pressure supply valve (if utilized) is opened, again placing the fluid in the low pressure line at a first low pressure. Then any locking mechanism that mechanically retains the pistons is deactivated (e.g., retracted) as cam 103 approaches TDC for each piston. Each piston is then reactivated by switching the respective high-speed control valve 111a-e from the “high pressure” position (the high pressure supply line actually being at atmospheric) to the low pressure position when cant 103 has reached (or has come very near) the TDC position for the respective piston. The pressure-relieving valve on the high pressure fluid line is then closed, and the high pressure supply valve is opened, repressurizing the fluid in the high pressure line. Low pressure active at the top of the respective piston 116a-e causes it to travel downward, reseating on cam 103. Motoring or pumping operation can resume when all pistons 116a-e have reseated on cam 103.

While any piston is deactivated, it is critical to maintain the proper position and orientation of the respective connecting rod to prevent a collision between the cam and the curved bearing surface (or “shoe”) of the connecting rod on re-engagement. The need for orientation depends in part on the manner of connection between the piston head and the connecting rod. If the piston is joined to the connecting rod by a pin, the piston and connecting rod will maintain their orientation to each other, but if the piston were to rotate in the bore, the surface of the connecting rod shoe would no longer align with the surface of the cam. If the piston is joined to the connecting rod by a ball joint, it introduces an additional degree of freedom whereby the connecting rod could also rotate with respect to the piston head. In either case, the edges of the shoe could impact the cam on re-engagement, possibly causing damage. It is therefore important in such arrangements to constrain the connecting rod shoe to the plane of rotation of the cam and to prevent the connecting rod shoe from swinging or rotating when it is out of contact with the cam surface.

Shown in FIG. 4 are orientation members 131a,b which may be pressed into or formed as part of the bottom of connecting rod 117. The orientation members may be cylindrical pins as depicted, or protruding bosses, or of any functionally similar form. The orientation members engage with grooves 201 on the inner surface of the pump/motor housing thereby constraining unwanted movement of the connecting rod when it is disengaged from the cam surface. Normally, when piston 116 is in an active state, connecting rod 117 bears upon cam 103, and members 131a,b are generally residing within the space provided by groove 201. Optionally, a single member may be used on one side of the bottom of the connecting rod, to engage with a single corresponding groove.

The orientation grooves and members are depicted more clearly in FIGS. 5-8. In FIG. 5, orientation groove 201 is seen now in profile, where its generally parabolic edge 203 can be seen. Orientation member 131 (being one of the pair 131a,b of FIG. 4) is now seen at the top of groove 201, with piston 116 being slightly above TDC, having been disengaged. The generally parabolic shape of edge 203 is derived from the path which the member would naturally follow as the connecting rod 117 “rocks” as it follows the cam. That is, owing to the natural “rocking” motion of each connecting rod 117 as it follows the eccentrically rotating cam, member 131 will take on a variety of positions and is free to move within the groove, sweeping out a volume as the cam rotates. Therefore the edge 203 of orientation groove 201 has the generally parabolic shape seen here, being defined by this sweep.

When the piston is in an engaged state, and connecting rod 117 is sliding against the surface of cam 103, connecting rod 117 will ordinarily not rotate with respect to its axis because it is essentially constrained by the cam. However, when piston 116 has been placed in a disengaged state as pictured in FIG. 5, connecting rod 117 will not be in contact with cam 103, and will therefore be potentially free to rotate about its axis, or to swing within or outside of the plane of rotation of the cam. In this position slightly above TDC, member 131 will travel to the very top of groove 201 where it becomes constrained against such rotation or swinging by semicircular profile 203a (as described in detail hereafter). This thereby keeps connecting rod 117 oriented in the proper orientation to reengage the rotating cam surface at a later time.

Piston 116 includes bore 171 which allows fluid to lubricate the piston/connecting rod interface and to additionally enter connecting rod bore 119 to lubricate the cam/connecting rod interface. Inside bore 171 is optional one-way pressure relief valve 172 which is configured to prevent fluid from passing out of chamber 110 when at a low pressure, but allow it to pass when at a higher pressure. The primary purpose of valve 172 is to prevent fluid present in chamber 110 from leaking out and being displaced by air when the piston is in an inactive state. When inactive, fluid in chamber 110 is at or near atmospheric pressure, and therefore cannot overcome the minimum pressure differential required to open valve 172. However, at other times, the fluid pressure in chamber 110 is easily enough to overcome the force and allow the fluid to pass into the lubricating passages. As an example, in a manner known in the art, valve 172 may include a ball 174 seated against a spring 173, where fluid pressure acting against the ball must cause the ball to compress the spring in order for the valve to “crack,” allowing fluid to pass. The stiffness of spring 173 may be selected to provide a desired cracking pressure, perhaps, for example, 5-7 psi, or any other pressure appropriate to the function.

FIG. 6 shows more clearly some key features of orientation groove 201. Member 131 is seen in a representative position along partial sweep path 410. Groove 201 includes a first swept volume 181 and a second swept volume 182. Optionally and preferably, edge 203 largely coincides with the outer edge of swept volume 181.

Swept volume 181 is derived from the sweep of the member 131 as the cam rotates when the piston is in an engaged state. As the cam rotates eccentrically, the “rocking” motion of the connecting rod causes member 131 to travel back and forth along or near edge 203 between position 401 and position 405, sweeping out volume 181.

Swept volume 182 is derived from the sweep of member 131 as it makes a disengagement motion as the piston is being disengaged. At the beginning of piston disengagement, member 131 would be at or very near position 403. At the end of disengagement, member 131 would be at position 404, resting against semicircular profile 203a. In this position, member 131 is constrained by profile 203a from rotating about the axis of the connecting rod 117, or from swinging within or outside of the plane of rotation of cam 103. Arcuate relieves 203b and 203c are preferably provided to accommodate the disengagement motion of member 131 if it begins from a position slightly offset from position 403.

FIGS. 7 and 8 show the orientation grooves as positioned on the inner surface of the housing. In FIG. 7, circular chamber 170 is provided, within which the cam eccentrically rotates. Representative groove 201 is machined into the inner surface of case 120 adjacent to chamber 170. In FIG. 8, the substantially semicircular profile 203a at the top of groove 201 can also be seen.

Referring again to FIG. 5, it can be seen that despite the fact that cam 103 is at the TDC position with respect to piston 116, gap 501 exists between the lower surface of connecting rod 117 and the surface of earn 103, so that earn 103 may pass by the shoe of connecting rod 117 with no friction. Gap 501 need not be large, for example, a gap of approximately 0.05-0.20 mm will suffice. The advantage of retaining a small gap rather than a large one is that it minimizes the potential impact force from reseating connecting rod 117 on cam 103.

Method of Operation in a Hydraulic Hybrid Vehicle

The described radial pump/motor is preferably employed in a series hydraulic hybrid powertrain to supplement the torque output of a variable displacement, over-center bent-axis pump/motor which is the primary pump/motor. The through shaft of the radial pump/motor preferably is connected to the output shaft of the primary pump/motor. The torque of both pump/motors are then combined to match the desired output torque, by varying the displacement of one or both pump/motors. For example, FIG. 8 is a general schematic of a vehicle to which the invention may be applied. This is only a representative example, as the invention is equally applicable to any hydraulic hybrid vehicle that would ordinarily utilize a mechanical transmission. Referring to FIG. 8, hydraulic hybrid vehicle 800 includes internal combustion engine 801 which drives shaft 840, providing mechanical power to drive engine pump 802. Engine pump 802 pressurizes fluid by drawing it from low pressure accumulator or reservoir 805 and pumping it toward high pressure accumulator 804, where it may be stored, or utilized to hydraulically power primary drive pump/motor 803. Drive motor 803 is variable displacement (and preferably over-center). Drive motor 803 then outputs mechanical power to the drivetrain by means of output shaft 841. Output shaft 841 is preferably a through-shaft that continues through radial pump/motor 899 of the invention. Radial pump/motor 899, if in an active state, can supplement the torque coming from drive motor 803, outputting the combined torque on shaft 842 which then drives the driven wheels 861, preferably through a gear reduction differential 860. If radial pump/motor 899 is not in an active state, it does not contribute torque and instead only transmits the torque received from shaft 841.

The radial pump/motor 899 is utilized at its fixed displacement and the primary over-center pump/motor 803 is utilized at a variable displacement to control the net output to the drivetrain by adding or subtracting a variable amount of displacement as necessary. For example, suppose that the over-center pump/motor has a maximum displacement of 233 cc/rev, and the radial pump/motor has a fixed displacement of 442 cc/rev, for a maximum total displacement of 675 cc/rev. Given a desired torque output and a system pressure, the required displacement of each device, adding up to a total displacement, can be computed. When the required total displacement is less than 233 cc/rev, the over-center pump/motor would be set to that displacement, and the pistons of the radial pump/motor would be disengaged from the earn. If more than 233 cc/rev but less than 442 cc/my is required, the radial pump/motor would operate as a motor at its full 442 cc/rev displacement, while the over-center pump/motor would operate in pump mode to absorb enough of the excess torque to effectively reduce the net displacement to match the required displacement. If more than 442 cc/rev is required (up to the 675 cc/rev capacity of the pair), the radial pump/motor would again operate as a pump/motor at its full 442 cc/rev displacement while the over-center pump/motor would operate as a motor to provide the it remainder.

This method is depicted in the flowcharts of FIGS. 10-16. Herein, positive (+) displacements represent a motoring mode, and negative (−) displacements represent a pumping mode. Referring to FIG. 10, a method is provided where the first pump/motor may operate at a variable displacement between (−y) and (+y), and the second pump/motor operates at a fixed displacement (+x), where |x|>=|y|. A torque T demanded of the drivetrain (for example, by the driver of the vehicle in interacting with an accelerator pedal) is determined (1002). The torque demand T is then apportioned between the first and second pump/motors (1003), such that the torque demand is fulfilled by one of or both of, the two. The apportionment is such that the first pump/motor, if utilized, is utilized at a positive or negative displacement within its variable range, and the second pump/motor, if utilized, is utilized at its substantially fixed positive displacement.

FIG. 10A depicts an example of said apportioning. Torque demand T is known as previously described. Torque t1 is determined (1010) as the torque that the first pump/motor can provide at its maximum positive displacement. If torque demand T is less than or equal to the torque t1 that the first (variable displacement) pump/motor can provide at its maximum positive displacement +y (1011), then the first pump/motor is utilized (1012) exclusively to meet torque demand T, being set to a displacement between zero and +y that would deliver that torque to the drivetrain. If torque demand T is greater than said torque t1 (1011), a torque t2 is determined (1013) as the torque that the second pump/motor can provide at its fixed positive displacement. A torque t3 is determined (1014) as the torque demand T minus the torque t2. The result t3 can be positive, negative, or substantially zero, or set to zero if near zero. The second pump/motor is operated (1015) at its fixed positive displacement to deliver the torque t2 to the drivetrain, and, if t3 is not substantially zero, the first pump/motor is operated at a positive or negative displacement that would deliver or absorb (respectively) the torque t3.

FIGS. 11-16 detail a preferred way in which the torque demand is apportioned, in which the speed of the drivetrain is also considered in determining the usage of the second pump/motor, which, being a radial motor, is likely to have a lower speed limitation than the first pump/motor. Referring to FIG. 11, a first rotational speed of the drivetrain is determined (1101), for example, by reference to a speed sensor. The first rotational speed is then compared (1102) to a maximum rotational speed of the second pump/motor, which may for example be predetermined, or be looked up as a value dependent on other factors such as system pressure, temperature, load, etc. If the first rotational speed does not exceed (1103) the maximum rotational speed of the second pump/motor, the steps of FIG. 12 take place (to be detailed later). Otherwise, if the first rotational speed is greater than the maximum rotational speed of the second pump/motor (1104), then a fourth displacement value is determined (1105) as the (motoring) displacement of the first pump/motor that would be necessary to supply all or a selected portion of the torque demand if it were acting as a motor. This displacement can be calculated knowing the characteristics of the pump/motor as well as sensed values such as fluid pressures. etc. The first pump/motor is utilized as a motor at the fourth displacement value (1106), and the second pump/motor is in an inactive state (1107) so as not to contribute torque.

The steps of FIG. 12 are now detailed. A first torque capacity is determined (1201), as the torque available from the first pump/motor if acting as a motor at maximum displacement. A second torque capacity is determined (1202), as the torque available from the second pump/motor if acting as a motor. Either torque capacity may be calculated by reference, for example, to system fluid pressure, temperature, etc, as may be gathered by the appropriate sensors. The torque demand is then compared (1203) to the first torque capacity and the second torque capacity. If the torque demand does not exceed the first torque capacity (1204), the steps of FIG. 13 are taken (to be detailed later). If the torque demand exceeds the first torque capacity and does not exceed the second torque capacity (1205), the steps of FIG. 14 take place (to be detailed later). If the torque demand exceeds the second torque capacity and does not exceed the sum of the first and second torque capacities (1206), the steps of FIG. 15 take place (to be detailed later). Otherwise, the torque demand exceeds the sum of the first and second torque capacities (1207), and the steps of FIG. 16 take place (to be detailed later). In this case, both pump/motors are utilized at their respective maximum available displacements in an effort to provide as much torque toward the torque demand as possible.

The steps of FIG. 13 are now detailed. A first displacement value is determined (1301) as the (motoring) displacement of the first pump/motor that would be necessary to supply the torque demand if acting as a motor. This displacement can be calculated knowing the characteristics of the pump/motor as well as sensed values such as fluid pressures, etc. The first pump/motor is utilized as a motor at the first displacement value (1302), and the second pump/motor is in an inactive state (1303) so as not to contribute torque.

The steps of FIG. 14 are now detailed. A torque surplus is determined (1401) as the difference between the second torque capacity and the torque demand. A second displacement value is determined (1402) as a (pumping) displacement of the first pump/motor at which it would require a torque input equal to the torque surplus if acting as a pump. This displacement can be calculated knowing the characteristics of the pump/motor as well as sensed values such as fluid pressures, etc. The second pump/motor is utilized as a motor (1403) at its substantially fixed displacement, and the first pump/motor is utilized as a pump at the second displacement value (1404). The torque thus absorbed from the drivetrain by the first pump/motor reduces the net torque output to the drivetrain by the second pump/motor so that the net torque to the drivetrain equals the torque demand.

The steps of FIG. 15 are now detailed. A torque deficit is determined (1501) as the difference between the torque demand and the second torque capacity. A third displacement value is determined (1502) as the (motoring) displacement of the first pump/motor that would be necessary to supply the torque deficit if acting as a motor. This displacement can be calculated knowing the characteristics of the pump/motor as well as sensed values such as fluid pressures, etc. The second pump/motor is utilized as a motor (1503) at its substantially fixed displacement, and the first pump/motor is utilized as a motor at the third displacement value (1504).

The steps of FIG. 16 are now detailed. The first pump/motor is utilized as a motor at its maximum displacement (1601), and the second pump/motor is utilized as a motor at its substantially fixed displacement. It should be noted that, if the torque demand substantially exceeded the sum of the first and second torque capacities, the total torque demand may not be fully met.

Deactivation and reactivation of the radial pump/motor is controlled by an appropriate electronic control system and control strategy. A preferred control strategy would consider rotational speed of the driveline and the net torque demanded of the vehicle. Below a first rotational speed, and above a threshold torque demand, the pistons of the radial pump/motor are placed in an engaged state and the radial pump/motor produces torque. When the vehicle has accelerated enough that the drivetrain now exceeds the first rotational speed, the pistons are retracted to deactivate the radial pump/motor. The pistons being deactivated no longer constrain the rotational speed of the radial pump/motor, allowing the shaft to increase its speed up to a greater speed than the maximum rotational speed of a prior art radial pump/motor of similar specifications.

Claims

1. A hydraulic machine, comprising: a housing;a power shaft, including an eccentric cam, configured to rotate about a common longitudinal axis;a plurality of working pistons positioned in respective cylinders, wherein the cylinders are oriented in a radial arrangement each being substantially perpendicular to the axis, and wherein each piston has a connecting rod and wherein each connecting rod has a bearing surface for slidably bearing upon a bearing surface of the eccentric cam;a high-speed control valve for each cylinder, configured to switchably connect the respective cylinder with a high pressure fluid source and a low pressure fluid source;a high pressure fluid line connecting each high-speed control valve with the high pressure fluid source;a low pressure fluid line connecting each high-speed control valve with the low pressure fluid source;a shut off valve for isolating the high pressure fluid line from the high pressure fluid source; anda pressure relieving valve on the high pressure fluid line for depressurizing fluid in the high pressure fluid line to facilitate disengagement of the connecting rod from the bearing surface.

2. The hydraulic machine of claim 1, additionally comprising: one or more members, rigidly connected to each connecting rod and extending substantially perpendicularly to the central axis of the connecting rod; andone or more grooves in the inner surface of the housing;wherein each groove includes a first volume swept by an outer end of the respective member in a revolution of the cam when the respective bearing surface is in contact with the cam; andeach groove includes a second volume swept by an outer end of the respective member when the respective bearing surface moves to a position outside the maximum radius of the cam; andwherein each member extends into a respective groove.

3. The hydraulic machine of claim 2, wherein: said groove, when said second volume is occupied by the respective, member, acts to restrain the respective member from movement in such a way that the central axis of the respective connecting rod is prevented from swinging within the plane of rotation of the cam.

4. The hydraulic machine of claim 2, wherein: said groove, when said second volume is occupied by the respective member, acts to restrain the respective member from movement in such a way that the central axis of the respective connecting rod is retained in an orientation substantially parallel with the plane of rotation of the cam.

5. The hydraulic machine of claim 2, wherein: said groove, when said second volume is occupied by the respective member, acts to restrain the respective member from movement in such a way that the respective connecting rod is prevented from substantially rotating about its central axis.

6. The hydraulic machine of claim 1, additionally comprising: a locking means for holding a piston in a position in which the bearing surface of the respective connecting rod is at or outside the maximum radius of the cam.

7. The hydraulic machine of claim 6, wherein: said locking means includes a pin that enters a respective cylinder at a position that is below the position of the bottom of the respective piston when the respective connecting rod bearing surface is at or outside the maximum radius of the cam.

8. The hydraulic machine of claim 6, wherein: said locking means is a permanent magnet or electromagnet configured to exert a magnetic force on said piston.

9. The hydraulic machine of claim 6, wherein: said locking means is a fluid pressure force exerted on a surface of the piston and a surface of the cylinder.

10. A method for deactivating and reactivating a radial hydraulic pump/motor having a plurality of working pistons positioned in respective cylinders, each piston having a connecting rod, each connecting rod having a bearing surface for slidably bearing upon an eccentric cam, the cam being connected to and rotating with a mechanical power shaft, comprising: deactivating a piston by causing the respective bearing surface of the respective connecting rod to move to a position at or beyond the maximum radius of the eccentric cam; andreactivating a piston by causing the respective bearing surface of the respective connecting rod to be in contact with the cam.

11. The method of claim 10, wherein: said deactivating includes the steps of: positioning each high-speed control valve to supply low pressure fluid to each cylinder;closing a high pressure fluid supply line from a high pressure fluid source supplying the pump/motor;opening a pressure relieving valve on the high pressure fluid supply line to a second low pressure that is lower than the first low pressure;allowing rotation of the cam to push a piston to a top dead center position;opening the cylinder space above said piston to the high pressure fluid supply line thereby causing fluid in said cylinder space to be at the second low pressure; andpositioning the piston at or beyond a top dead center position; andsaid reactivating includes the steps of: allowing the cam to approach a rotational position at which a piston would be brought to a top dead center position if it were active;opening the cylinder space above said piston to the low pressure fluid source;allowing the low pressure active in the cylinder space above said piston to move the piston toward the cam until the bearing surface of the respective connecting rod is seated upon the cam;closing the pressure relieving valve on the high pressure fluid supply line; andopening the high pressure fluid supply line to the high pressure fluid source.

12. The method of claim 11, wherein: the second low pressure is substantially equal to case pressure below the pistons.

13. The method of claim 11, wherein: a piston is locked at its top dead center position or beyond.

14. The method of claim 11, wherein: the second low pressure is below case pressure.

15. A method for providing torque to a drivetrain, comprising: providing a first variable displacement pump/motor having a range of allowable displacements;providing a second pump/motor having a substantially fixed displacement larger than the maximum displacement of the first pump/motor and having a power shaft connected to the output shaft of the first pump/motor;determining a torque demand of the drivetrain; andapportioning the torque demand between the first pump/motor and the second pump/motor by selecting a displacement and/or state of utilization of each, wherein the first pump/motor, if utilized, is utilized at a displacement within its range of allowable displacements, and wherein the second pump/motor, if utilized, is utilized at its substantially fixed displacement.

16. The method of claim 15, wherein said apportioning includes the steps of: determining a first torque as the torque that the first pump/motor can provide at its maximum positive displacement;comparing the torque demand to the first torque; and if the torque demand is less than or equal to the first torque, operating the first pump/motor at a positive displacement that would deliver the first torque, and having the second pump/motor in an inactive state delivering no torque; andif otherwise, then: determining a second torque as the torque that the second pump/motor can provide at its fixed positive displacement;determining a third torque as the torque demand minus the second torque, wherein the third torque may thereby be a positive, negative, or substantially zero quantity;operating the second pump/motor at its fixed positive displacement, andif the third torque is positive, operating the first pump/motor at a positive displacement that would deliver the third torque;otherwise if the third torque is negative, operating the first pump/motor at a negative displacement that would absorb the third torque;otherwise if the third torque is substantially zero, operating the first pump/motor at a substantially zero displacement.

17. The method of claim 15, wherein said apportioning includes the steps of: determining a first rotational speed of the drivetrain;comparing the first rotational speed to a maximum rotational speed of the second pump/motor; andif the first rotational speed is not greater than the maximum rotational speed, then: determining a first torque capacity, as the torque available from the first pump/motor if acting as a motor at maximum displacement;determining a second torque capacity, as the torque available from the second pump/motor if acting as a motor;comparing the torque demand to the first torque capacity and the second torque capacity; andif the torque demand does not exceed the first torque capacity, determining a first displacement value, as a displacement of the first pump/motor necessary to supply the torque demand if acting as a motor,utilizing the first pump/motor as a motor at the first displacement value, andhaving the second pump/motor in an inactive state;if the torque demand exceeds the first torque capacity and does not exceed the second torque capacity, determining a torque surplus as the difference between the second torque capacity and the torque demand,determining a second displacement value, as a displacement of the first pump/motor at which it would require a torque input equal to the torque surplus if acting as a pump,utilizing the second pump/motor as a motor, andutilizing the first pump/motor as a pump at the second displacement value;if the torque demand exceeds the second torque capacity and does not exceed the sum of the first and second torque capacities, determining a torque deficit, as the difference between the torque demand and the second torque capacity,determining a third displacement value, as a displacement of the first pump/motor necessary to supply the torque deficit if acting as a motor,utilizing the second pump/motor as a motor, andutilizing the first pump/motor as a motor at the third displacement value; andif the torque demand exceeds the sum of the first and second torque capacities, utilizing the first pump/motor as a motor at maximum displacement, and utilizing the second pump/motor as a motor;and if the first rotational speed is greater than the maximum rotational speed, then: determining a fourth displacement value, as a displacement of the first pump/motor necessary to supply at least a portion of the torque demand if acting as a motor,utilizing the first pump/motor as a motor at the fourth displacement value, andhaving the second pump/motor in an inactive slate.

18. The method of claim 15, wherein: the first pump/motor is an axial piston pump/motor, andthe second pump/motor is a radial pump/motor.

19. The method of claim 18, wherein: the second pump/motor has a mechanically fixed displacement.

20. The method of claim 18, wherein: the second pump/motor is the hydraulic machine of claim 1.