US3465680A - Hydraulic pump or motor unit - Google Patents

Description

Sept. 9, 1969 T. SAILA 3.465,68@

HYDRAULIC PUMP OR MOTOR UNIT Filed May 31, 1967 C455 PRESSURE" CoA/77601. lf/0a PRESSURE J pressi/2 IM5 v 9,52% j/SZ/ewa/x www United States Patent 3,465,680 HYDRAULIC PUMP R MOTOR UNIT Tauno Sala, Los Angeles, Calif., assgnor to Sundstrand Corporation, a corporation of Delaware Filed May 31, 1967, Ser. No. 642,426 Int. Cl. F04!) 1/02, 49/ 00 U.S. Cl. 103-38 10 Claims ABSTRACT OF THE DISCLOSURE A hydraulic pump or motor unit of the variable displacement type having a pressure compensating valve responsive to system pressure for varying the displacement of the unit, the compensator valve also providing leakage ow as a function of system pressure for increasing the damping ratio of the compensator valve to minimize hunting of the control.

BACKGROUND OF THE INVENTION This invention relates generally to hydraulic uid energy translating devices and more particularly to a control for varying the displacement of a multiple piston hydraulic unit.

In the past hydraulic pumps of the multiple piston type having been provided with pressu-re compensator valves which in response to the pressure of outlet flow from the pump vary the displacement of the hydraulic unit to maintain a substantially constant supply pressure. These variations in the displacement of the unit to achieve this pressure compensation may be necessary because of a change in demand of the hydraulic system supplied by the pump or because of variations in the speed of rotation of the pump itself, or both.

These pressure compensators, when associated with a pump, respond to an increase in the pressure of outlet ow from the pump to reduce the displacement of the pump until the desired pressure is reestablished, and respond to a decrease in the pressure of fluid from the pump to increase the displacement until the desired outlet flow pressure is reestablished. It is apparent that to achieve this end there must be provided in association with the compensator valve a closed loop control circuit. This may be effected by biasing one side of a valve member in the compensator valve with supply fluid itself and biasing the other end of the valve With-a spring so that at the desired system pressure the valve is in a balanced or equilibrium position substantially blocking flow relative to a displacement control motor and in response to increases or decreases in supply pressure the valve will become unbalanced and selectively connect the control motor with either supply pressure or a suitably arranged drain thereby effecting a change in displacement of the associated hydraulic unit.

When the supply pressure in an exemplary construction increases above the desired value the valve shifts against the opposing force of the spring, which then increases, porting supply fluid to the displacement control motor in a direction to decrease the displacement of the hydraulic unit acting as a pump. The reduction in displacement causes a reduction in supply Huid pressure permitting the valve to return toward 1ts blocking position under the influence of the increased spring force. However, as will be apparent to those skilled in this art an overshoot will 3,465,680 Patented Sept. 9, 1969 ICC often occur and the valve will proceed through a period of oscillation the amount of which decreases with time. It is desirable that these oscillations or hunting of the compensator valve under these conditions be reduced to a minimum and it is to this objective that the present invention is addressed.

Various modifications may be made in the compensator valve and the control piston to vary the damping ratio and hence the extent of oscillation of the compensator valve. These might include variations in port area, spool size, spring rate, displacement' control piston size, and the length of the moment arm of the control force applied -to the variable displacement cam associated with the unit. However, for a given hydraulic unit for certain operating conditions these parameters all have optimum values and deviations from these optimum values may significantly decrease the performance of the unit. For example, certain variations from the optimum values of these parameters may cause a natural frequency problem between the hydraulic unit itself and the compensator valve.

SUMMARY OF THE INVENTION In accordance with the present invention a leakage term is introduced into the equation for damping ratio (of the compensator valve) that is a direct function of supply pressure. This minimizes oscillations of the compensator valve and, very importantly, does not require a corresponding change in any of the parameters listed above. That is, the leakage term increases the damping ratio of the control without requiring a change in any of the other control parameters from their optimum values. The overall effect on the hydraulic unit design is a reduction in size and weight for a given requirement in damping ratio.

To accomplish this result there is provided in a compensator valve construction of the type described generally above a valving land on the compensator valve associated with an additional port communicating with the displacement control motor and arranged so that as supply pressure increases shifting the valve to a position porting supply fluid to the control motor, the additional land will progressively and proportionately increase a leakage path from the displacement control motor. Since the resulting leakage is a direct function of supply pressure it has a stabilizing affect on the control oscillations of the compensator valve when the transient cyclical control pressure reaches its high peaks. This prevents the displacement of the hydraulic unit from becoming too low when compensating for an increase in supply pressure.

It should be understood that while the advantages of the present compensator valve damping arrangement are particularly useful with respect to a hydraulic pump, that certain advantages are obtained when employing the principles of the invention in a control arrangement for a motor.

BRIEF DESCRIPTION OF THE DRAWINGS FIG. l is a schematic illustration of an axial piston hydraulic unit and a compensator valve arrangement according to the present invention, with the compensator valve enlarged relative to the hydraulic unit for clarity, and

FIG. 2 illustrates curves showing the hunting characteristics of a compensator valve with and without the present improvement in damping ratio.

3 DESCRIPTION OF THE PREFERRED EMBODIMENT Referring now to the drawings and particularly FIG. 1 thereof, the present hydraulic circuit is seen to consist generally of a variable displacement hydraulic unit and a compensator valve assembly 11 arranged to control the displacement of the hydraulic unit.

The hydraulic unit 10 is adapted to operate either as a pump or as a motor. When operating as a pump it supplies fluid under a regulated pressure to a system (not shown), and when operating as a motor receives hydraulic fluid (from a suitable source) and drives a load through a shaft indicated diagrammatically at 14. An exemplary application for a combined pump/motor of this type is as a starting motor for an aircraft engine which operates after starting as a hydraulic pump to supply fluid under a regulated pressure to hydraulically operated equipment associated with the aircraft.

The hydraulic unit 10 is generally of the type described in the copending application of Walter J. Iseman, entitled Fluid Translating Device, Ser. No. 578,356, iled Sept. 9, 1966, assigned to the assignee of the present invention, to which a reference should be made lfor a more complete description of the hydraulic unit itself. A brief description thereof will serve the present purposes. A stationary valve plate 15 has ports 16 and 17 therein (illustrated diagrammatically) communicating respectively with main system passages 18 and 19. Passage 18 carries high pressure fluid and passage 19 carries low pressure uid regardless of whether the hydraulic unit is operating as a pump or as a motor. A cylinder block 20 has axially disposed cylinders 21 therein communicating with a port face 22 on the end of the cylinder block which slidably engages valve plate 15. As the cylinder block 20 rotates with respect to valve plate 15 the cylinders 21 serially communicate with the ports 16 and 17 in well known fashion.

Pistons 24, slidably mounted in the cylinders 21, are reciprocated byv a pivotally mounted cam member 23 through articulated connected rods 26.

The cam member 23 is movable from a neutral position where cam face 27 is perpendicular to the axis of drive shaft 14 to a first maximum displacement position in the direction of arrow 30 and to a second maximum displacement position in the direction of arrow 31, As will appear hereinafter when the device operates with the cam 23 on the clockwise side of neutral, (on the arrow 30 side) it operates as a pump delivering regulated pressure to the system, and when it operates with the cam on the counterclockwise side of neutral, it operates as a motor delivering torque to shaft 14. It will be understood that during the pumping mode that a suitable Prime mover drives shaft 14 in rotation.

A lluid operable control motor 36 is provided for varying the displacement of hydraulic unit 10 and includes a cylinder 37 with a piston 38 slidably mounted therein.

Piston 38 positions the cam member 23 through a connecting rod 39 pivotally mounted at 40 to the -cam member. The control motor 36 when pressurized thus pivots the cam 23 in a counterclockwise direction reducing the displacement when the unit is pumping and increasing the displacement when it is motoring. A coil compression spring assembly 40 is provided for continuously biasing the cam member 23 in a counterclockwise direction in opposition to the control motor 36. Spring assembly 40' includes a coil compression spring 42 seated at one end against a stationary seat and at its other end in a movable seat 43 receiving an articulated connecting rod 44 pivotally connected to the cam member 23 at 46. Spring 42 thus provides a continual bias on the cam member 23 tending to increase the displacement of the hydraulic unit when pumping and decrease the displacement of the hydraulic unit when motoring.

The compensator valve assembly 11 controls the displacement of the hydraulic unit 10 by controlling the flow Cil of fluid to the cylinder 37 through control passage 48. Valve assembly 11 includes a valve sleeve 55 xedly seated in a suitable bore 56 in a valve housing member shown only diagrammatically at 58. Sleeve 55 has a central through bore 59 slidably receiving a spool valve member 61. Radial ports 62 communicate with the interior of valve sleeve 55 and with an annular recess 63 in the outside of the sleeve continually communicating with a control passage 66 in communication with control passage 48. Adjacent to ports 62 are radial ports 70 communicating with an annular recess 72 in continuous communication with a drain passage 74 in housing 58 continuously communicating with a low pressure tank (or case pressure).

The valve member 61 has a left valve land 76 which in the equilibrium position shown substantially blocks communication between the ports 62 and the interior of the sleeve 55 on either side of the land so that the control motor piston 38 is electively locked in position and the displacement of the hydraulic unit is maintained.

A coil compression spring 81 is provided for biasing the valve member 61 toward a position communicating the drain ports with control passage 66 across land 76. One end of the coil compression spring 81 is seated in the right enlarged end of bore S6 and the other end is received on a spring seat 83 engaging the right end of valve member 61.

For urging the valve member 61 in the other direction the left end of the valve member 61 defines a piston communicating with supply fluid in high pressure line 18 through supply pressure control passage 8S. Thus supply fluid pressure acts on the left end of valve member 61 urging it to the right against the bias of spring 81, Spring 81 is sized so that at the desired system pressure (acting on the left end of valve member 61) the valve land 76 will assume the position shown in FIG. 1 automatically blocking flow relative to the control motor 36 and maintaining the displacement of the unit.

As thus far described the operation of the pressure compensator valve 11 when the hydraulic unit 10 is acting as a pump is as follows. Assuming initially that the cam member is positioned in some positive displacement position, such as the one shown in FIG. 1, and the unit is delivering iluid under the desired pressure to the system through passage 18, the compensator 11 will assume the position shown. If under these conditions the demand of the system is reduced the pressure in main conduit 18 will rapidly increase. The sudden increase in pressure acting on the left side of valve member 61 will shift it to the right communicating the port 62 with supply fluid in line 85. During this initial (transient) movement of the valve member to the right the control pressure in passages 66 and 48 may approach the pressure in line 85.

Control uid under pressure is thus ported to the control motor cylinder 37 and the cam 23 is moved in a direction reducing the displacement of the hydraulic unit 10 (in the pumping mode) reducing the ilow from the pump and thereby decreasing the pressure in supply line 18. This reduction in pressure is sensed by the valve 61 (by a reduction in force on the left end thereof) and it begins return movement toward the left under the action and increased force of biasing spring 81. As will be apparent to those skilled in this art there will be some overshoot of valve land 76 on its return movement past the equilibrium position shown in FIG. 1 followed by decaying oscillations which are inherent in a closed loop control system of this type. These oscillations, in terms of stem displacement and control pressure are illustrated at 91 and 91a, respectively, in FIG. 2.

Upon an increase in demand of the system upon the hydraulic unit 10 (still acting as a pump) the pressure in passage 18 and passage 85 will decrease causing a leftward shifting of valve member 61 and communication between control ports 62 and drain ports 70 across the right edge of land 76. Thus control fluid may escape from passage 66 to drain passage 74 decreasing the pressure in control cylinder 37 and permitting the biasing spring 42 to pivot cam member 23 and increase the displacement of the unit. This results in an increase in pressure in outlet passage 18 and supply control passage 85 associated therewith so that the fluid pressure acting on the left side of valve member 61 increases, returning it to the equilibrium position shown when the desired pressure is reestablished. The movement of compensator valve stem 61 and the variation in control pressure during this increase in flow demand of the system is illustrated at 95 and 95a, respectively, in FIG. 2. As will be appreciated by those skilled in this art the problem of oscillation or a lack o'f stabilization off the compensator valve is not significant when going from a low displacement to a high displacement in the pumping mode, i.e. upon an increase in system demand.

It should be understood, that =while the operation of the device as thus far described has been referenced with respect to variations in the demand of an associated system and assumes a substantially constant speed of driving shaft 14, that the componesator valve 11 will operate to regulate system pressure when the demand is constant and speed of tbe drive shaft 14 varies.

To minimize hunting of the valve member 61 and to thereby stabilize the compensator valve 11 during a decrease in displacement of the cam 23 'when the hydraulic unit operates as a pump, a leakage path is provided for control pressure fluid in passage 66, which is a direct function of the pressure in supply passage 18. Toward this end additional ports 100 are provided in the valve sleeve 55 communicating with an annular recess 101 in continuous communication with control iluid passage 66 through passage 66a. A land 105 is provided on valve member 61 having a land edge 106 adjacent to the reduced stem portion 107 which controls communication between the leakage ports 100 and the drain ports 70. Thus, as the valve member 61 moves to the right'from the position shown the land edge 106 will progressively uncover ports 100 providing an increasing leakage path from the control pressure passage 66. Since the displacement of member 61 is a function of the magnitude of the supply pressure acting on the left end of the valve member, the leakage flow across land 105 to drain ports 70 is also a direct function of supply pressure in passage 18.

The land edge 106 is approximately aligned with the left edge of ports 100, although the exact position of this edge for optimum performance will depend upon the particularly hydraulic unit employed, the operating conditions and the selected parameters of the compensator valve 11. In some instances the valve land 105 may be overlapped (ie. closer to edge 108 of valve land 76 than the spacing between ports 62 and y100) 0r in some cases it may Ibe slightly underlapped (ie. the land edges being spaced further than the ports).

The leakage path thus modifies the operation of the compensator valve 11, When the supply tlow from the device 10, when operating as a pump, increases above the demands of the system, and a pressure rise occurs in passage 85, shifting valve member 61 to the right, ports 62 lwill open letting higher pressure fluid into the control passage 66 and control cylinder 37 decreasing the stroke of the hydraulic unit 10. During this rightward movement the leakage path through ports 100 increases as the supply pressure reaches its peak and prevents the stroke of the unit from decreasing too far as would occur under normal overshoot as described above. Thus, as seen at 110 and 110a in FIG. 2 the length of time of the oscillations of valve stem 61 and the control pressure are reduced due to this leakage tlow across port 100. The addition of ports 100 introduces a leakage term into the equation for damping ratio (which defines the extent of oscillations of the valve member 61). As noted above this increases the damping ratio without changing any of the other control parameters associated with the compensator valve 11 thereby permitting the optimum design of these parameters for a given hydraulic unit.

The damping function of the leakage ports -100 also has some effect when the hydraulic unit 10 is operating as a motor. To operate the hydraulic unit 10 as a motor, so that the drive shaft 14 and the load associated therewith are driven, the main high pressure passage 18 is connected to a suitable source (not shown) of hydraulic fluid under pressure. To effect the proper shifting of the cam member 23 to the opposite side of neutral against the increased force of biasing spring 42 this additional source of supply tiuid under pressure may be in excess of the regulated pressure when the device is operating as a pump. For example, if the regulated pressure during pumping is 4,000 p.s.i., the supply fluid pressure to the hydraulic unit during motoring (at least when the demand of the unit 10 is low) might be 4,300 p.s.i. Additionally, means may be provided for selectively increasing the effective area of the left side of valve 61 to amplify the force of the supply fluid signal and thereby effect shifting of the cam to the motoring side of neutral. Such means may take the form of a selectively operable biasing piston engaging the left end of valve 61 and having a greater area in communication with passage 18. The manner of achieving this shifting forms no part of the present invention and is therefore not described in detail, Of course, suitable controls (not shown) would be provided for disconnecting the conduit 18 from the associated system during this mode of operation.

The supply pressure signal acting through passage and across valve land 76 shifts (however achieved) the cam 23 toward its maximum displacement position on the other side of neutral under these conditions. If the hydraulic unit 10 were employed under these conditions as an engine starter, as the engine speed increased the demand of the hydrdaulic unit would increase and assuming it increased above the capacity of the pump (not shown) supplying fluid to the hydraulic unit the pressure in line 18 would begin to decrease. This decrease of pressure in line 18 as sensed by the valve member 61 would cause shifting of the valve to the left permitting the displacement of the hydraulic unit to decrease thereby matching the demand of the hydraulic unit 10 with the maximum capacity of the pump supplying fluid thereto at least during a portion of the starting cycle. The control by the compensator valve 11 during motoring including the function of leakage ports providing leakage as a function of supply pressure in line 85, serves to stabilize the valve member 61 in the same manner as described fully above with respect to the pumping operation.

I claim:

1. A hydraulic energy translating device, comprising: valve means having inlet and outlet ports therein, a cylinder block having plural cylinders therein serially communicable with said ports on relative rotation of said valve means and said cylinder block, pistons slidable in said cylinders, adjustable cam means for reciprocating said pistons and varying the displacement of the device, control means for moving the cam means in one direction in response to an increase in pressure in one of said ports and for moving the cam means in the other direction in response to a decrease in pressure in said one port including valve means connected to control said cam means, means deriving a control signal representing the pressure of uid in said one port and applying the same to bias the valve means in one direction toward a state causing movement of said cam means in said one direction, means applying a biasing Asignal to bias said valve means in the other direction toward a state causing movement of said cam means in the other direction, said Valve means producing an output signal according to the relative magnitudes of said biasing signal and said control signal to control said cam means, and means to reduce said output signal as a function of said control signal to minimize hunting of said valve means without modifying said biasing signal, said means to reduce said output signal including second valve means providing an increasing leakage path as the first valve means moves in said one direction from the control signal.

2. A hydraulic energy translating device as defined in claim 1, wherein said control means decreases the displacement of the device in response to an increase in' pressure in said one port and increases the displacement of the device in response to a decrease in pressure in said one port, said one port being the outlet port.

3. A hydraulic energy translating device, comprising: valve means having inlet and outlet ports therein, a cylinder block having plural cylinders therein serially communicable with said ports on relative rotation of said valve means and said cylinder block, pistons slidable in said cylinders, adjustable cam means for reciprocating said pistons and varying the displacement of the device, control means for positioning said cam means in response to pressure in one of said ports including a fluid operable control motor for positioning said cam means, a compensator valve for controlling the flow of fluid relative to said control motor, means continuously biasing said valve toward a position causing movement of said control motor in one direction, means for applying a fluid pressure signal to said valve proportional to the fluid pressure in said one of said ports in opposition to said biasing means, said fluid pressure signal biasing said valve toward a position causing movement of said control motor in the other direction, said valve being movable to an equilibrium position substantially blocking flow relative to said control motor at the desired pressure level in said one port, and means for modifying the fluid flow relative to said control motor as a function of pressure in said one port to minimize hunting of said control motor upon an increase in pressure in said one of said ports including means providing a leakage path for fluid from said compensator valve, that increases in size as the compensator valve moves in direction causing movement of the control motor in the other direction.

4. A hydraulic energy translating device as defined in claim 3, wherein said modifying means includes a second valve responsive to pressure in said one port for providing said leakage path for the flow of fluid relative to said control motor caused oy said compensator valve thereby reducing the pressure oscillations in said fluid.

5. A hydraulic energy translating device as defined in claim 3, wherein said control means maintains a substantially constant pressure in said one port.

6. A hydraulic energy translating device comprising: valve means having inlet and outlet ports therein, a cylinder block having plural cylinders therein serially communicable with said ports on relative rotation of said valve means and said cylinder block, pistons slidable in said cylinders, adjustable cam means for reciprocating said pistons and varying the displacement of the device, control means for positioning said cam means in response to pressure in one of said ports including a fluid operable control motor for positioning said cam means, a compensator Ivalve for :controlling the flow of fluid relative to said control motor, means continuously biasing said valve toward a position causing movement of said control motor in one direction, means for applying a fluid pressure signal to said valve proportional to the fluid pressure in said one of said ports in opposition to said biasing means, said fluid pressure signal biasing said valve toward a position causing movement of said control motor in the other direction, said valve being movable to an equilibrium position substantially blocking llow relative to said control motor at the desired pressure level in said one port, means for modifying the fluid flow relative to said control motor as a function of pressure in said one port to minimize hunting of said control motor upon an increase in pressure in said one of said ports, said modifying means including a second valve responsive to pressure in said one port for modifying the flow of fluid relative to said control motor caused by said compensator valve by reducing the pressure oscillations in said fluid, said compensator valve and said second valve including a valve member, said biasing means including spring means acting on one end of said Valve member, said fluid pressure signal biasing the other end of said valve lmember, first port means communicating with said control motor, first land means on said valve member movable relative to said first port means from an equilibrium position substantially blocking flow relative to said control motor in one direction to port fluid in said one lport to said control motor and in the other direction to connect said control motor to a tank, second port means communieating with said control motor, and second land means on said valve member movable relative to said second port means to progressively increase leakage from said second port means as said valve member moves in said one direction.

7. A hydraulic energy translating device as defined in claim 6, wherein said second land means substantially blocks said leakage from said second port means when the valve member is in said equilibrium position.

8. A hydraulic energy translating device as defined in claim 3 wherein said calm means is biased toward a maximum displacement position, said fluid operable control motor when pressurized moving said cam means toward a minimum displacement position, said fluid pressure signal tending to lmove said valve to a position porting fluid to said control motor to reduce the displacement of the device, said biasing means tending to move said valve to a position porting fluid from the control motor to increase the displacement of the device, whereby the device operates as a pump delivering substantially constant pressure.

9. A hydraulic energy translating device as defined in `claim 3, wherein said cam means is biased toward a minimum displacement position when operating as a motor, said fluid operable control motor when pressurized moving said cam means toward maximum displacement, said fluid pressure signal tending to move said valve to a position causing an increase in displacement of the device, said biasing means tending to move said valve to a position reducing displacement of the device.

10. A hydraulic energy translating pump, comprising: valve means having inlet and outlet ports therein, a relatively rotatable cylinder block having axially disposed cylinders therein serially communicable with said ports, pistons slidable in said cylinders, a pivotally adjustable cam member for reciprocating said pistons, a fluid operable control motor for positioning said cam member to vary the displacement of the pump, and a pressure compensator valve for maintaining substantially constant pressure in said outlet port including a valve sleeve having first port means communicating with said fluid operable means, second port means communicating with low pressure and third port means communicating with said fluid operable means, a valve member slidable in said sleeve and having first land means thereon, said valve member being movable from a first position where the first land means substantially blocks flow relative to said first port means to a second position where the first land means 4communicates the fluid in said outlet port with said fluid operable motor and to a third position where the first land means communicates the fluid operable motor with said second port means, second land means on said valve means, said second land means being positioned so that as said valve member moves from the first position toward said second position it will provide increasing communication between said third port means and said second port means thereby to proportionately reduce the pressure increase in said fluid operable means to minimize hunting of said valve member, means for applying a signal References Cited UNITED STATES PATENTS 2,678,607 5/1954 Hufferd et al 103-120 2,915,985 12/1959 Budzich 103-162 2,921,560 1/1960 Budzich 91-433 3,186,353 1/1965 Taplin 103--162 3,272,135 9/1966 Bloomquist 103--38 WILLIAM L. FREEH, Primary Examiner U.S. Cl. XR.

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