Hydraulic piston pump with throttle control

A pump system has a piston pump. The piston pump has a cylinder block with an inlet port, an outlet port, and a plurality of cylinders. Each cylinder in the plurality of cylinders is connected to the inlet port by an inlet passage and to the outlet port by an outlet passage. The piston pump has a plurality of pistons disposed in the plurality of cylinders. A drive shaft drives the pistons within the cylinders. A throttle member independently throttles flow in each inlet passage. The pump system has an electrohydraulic actuator governing movement of the throttle member.

FIELD

The present disclosure relates to hydraulic pumps, and more specifically to mechanisms for controlling hydraulic pump systems.

BACKGROUND

U.S. Patent Application Publication No. 2012/0111185, which is hereby incorporated by reference in entirety, discloses a high efficiency diametrically compact, radial oriented piston hydraulic machine. The machine includes a cylinder block with a plurality of cylinders coupled to a first port by first valve and to a second port by a second valve. A drive shaft with an eccentric cam is rotatably received in the cylinder block and a cam bearing extends around the eccentric cam. A separate piston is slideably received in each cylinder. A piston rod is coupled at one end to the piston and a curved shoe at the other end abuts the cam bearing. The curved shoe distributes force from the piston rod onto a relatively large area of the cam bearing and a retaining ring holds each shoe against the cam bearing. The cylinder block has opposing ends with a side surface there between through which every cylinder opens. A band engages the side surface closing the openings of the cylinders.

U.S. patent application Ser. No. 13/343,436, which is hereby incorporated by reference in entirety, discloses a radial piston pump having a plurality of cylinders within which pistons reciprocally move. Each cylinder is connected to a first port by an inlet passage that has an inlet check valve, and is connected to a second port by an outlet passage that has an outlet check valve. A throttling plate extends across the inlet passages and has a separate aperture associated with each inlet passage. Rotation of the throttling plate varies the degree of alignment of each aperture with the associated inlet passage, thereby forming variable orifices for altering displacement of the pump. Uniquely shaped apertures specifically affect the rate at which the variable orifices close with throttle member movement, so that the closure rate decreases with increased closure of the variable orifices.

SUMMARY

Pump systems are disclosed. In some examples, the pump system has a piston pump comprising a cylinder block having an inlet port, an outlet port, and a plurality of cylinders disposed therein, each cylinder in the plurality of cylinders being connected to the inlet port by a respective inlet passage in a plurality of inlet passages and to the outlet port by a respective outlet passage in a plurality of outlet passages. The piston pump has a plurality of pistons, each piston in the plurality of pistons being disposed in a respective cylinder in the plurality of cylinders. A drive shaft drives the plurality of pistons within their respective cylinders. A throttle member independently throttles flow in each inlet passage in the plurality of inlet passages. The pump system can further comprise an electrohydraulic actuator governing movement of the throttle member.

In further embodiments, the pump system has a piston pump comprising a cylinder block having an inlet port, an outlet port, and a plurality of cylinders disposed therein, each cylinder in the plurality of cylinders being connected to the inlet port by a respective inlet passage in a plurality of inlet passages and to the outlet port by a respective outlet passage in a plurality of outlet passages. The piston pump can have a plurality of pistons, each piston in the plurality of pistons being disposed in a respective cylinder in the plurality of cylinders. A drive shaft drives the plurality of pistons within the respective cylinders. A throttle member independently throttles flow in each inlet passage in the plurality of inlet passages. The pump system can further comprise a load sense apparatus governing movement of the throttle member based upon a load sense signal and an electrohydraulic actuator governing movement of the throttle member based upon an electronic signal.

In further embodiments, the pump system has a piston pump comprising a cylinder block having an inlet port, an outlet port, and a plurality of cylinders disposed therein, each cylinder in the plurality of cylinders being connected to the inlet port by a respective inlet passage in a plurality of inlet passages and to an outlet port by a respective outlet passage in a plurality of outlet passages. The piston pump can have a plurality of pistons, each piston in the plurality of pistons being disposed in a respective cylinder in the plurality of cylinders. A drive shaft drives the plurality of pistons within the respective cylinders. A throttle member independently throttles flow in each inlet passage in the plurality of inlet passages. The pump system can further comprise a load sense apparatus governing movement of the throttle member based upon a load sense signal and an electrically operated actuator governing movement of the throttle member based upon an electronic signal.

DETAILED DESCRIPTION OF THE DRAWINGS

With reference toFIGS. 1 and 2, a hydraulic pump10has a cylinder block30with exterior first and second end surfaces21and22between which a cylindrical exterior side surface38extends. Although a radial piston pump is shown herein, the following structures and systems could also be incorporated with and/or incorporate a wobble plate pump, or any non-variable displacement pump or the like. The cylinder block30has an inlet port28and an outlet port29through which hydraulic fluid is received and expelled from a hydraulic system. The inlet and outlet ports28and29open into inlet and outlet galleries31and32, respectively, that extend in circles through the cylinder block30around a central shaft bore41in the cylinder block30. Three cylinders36extend radially outward from and are oriented at 120 degree increments around the central shaft bore41. Although the exemplary pump10is illustrated with three cylinders to simplify the drawings, in practice the pump may have a greater number of cylinders (e.g., 6 or 8 cylinders) to reduce torque, flow and pressure ripples at the outlet. Each cylinder36includes a tubular sleeve39that is inserted into a bore in the cylinder block30. Although the tubular sleeve39is beneficial in reducing the diameter of the pump10as will be described, the sleeve can be eliminated by using a material for the cylinder block that can be machined to form the cylinder bores. Each cylinder36has an opening through the cylindrical side surface38of the cylinder block30. A sealing cup24with an O-ring is placed inside each opening and a continuous band-shaped closing ring35extends around the side surface38tightly closing each of the cylinder openings. The closing ring35eliminates the relatively long plugs that projected outward from the cylinders in conventional pump designs and thereby reduces the overall diameter of the pump10.

With particular reference toFIG. 2, a plurality of inlet passages26are formed by first bores that extend into the first end surface21of the cylinder block30and each inlet passage opens into both the inlet gallery31and a respective one of the cylinders36. In other words, each inlet passage26is directly connected to both the inlet gallery31and one of the cylinders36. A separate inlet check valve33is located in each of those inlet passages26. The inlet check valve33opens when the pressure within the inlet passage26is greater than the pressure within the associated cylinder chamber37, as occurs during the intake phase of the pumping cycle. A plurality of outlet passages27are formed by second bores that extend into the second end surface22of the cylinder block30with each outlet passage opening into both the outlet gallery32and a respective one of the cylinders36. Every outlet passage27is directly connected to both the outlet gallery32and one of the cylinders36. A separate outlet check valve34is located in each of those outlet passages27. The outlet check valve34opens when pressure within the associated cylinder chamber37is greater than the pressure within the outlet gallery32, as occurs during the exhaust phase of the pumping cycle. It should be understood that the inlet and outlet galleries31and32communicate with all the piston cylinders in the pump and an identical pair of check valves is provided for each cylinder. As depicted inFIG. 2, each of the inlet and outlet check valves33and34is passive, meaning that it operates in response to pressure exerted thereon and not by an actuator, such as an electric solenoid. However, the scope of the present disclosure also covers inlet and outlet values that are actuated by other than pressure exerted thereon.

The tubular sleeve39that partially forms the cylinder36enables the inlet and outlet check valves33and34to be placed closer to the longitudinal axis25of the drive shaft40. Note that the inlet and outlet check valves33and34are within the closed curved perimeter defined by the exterior side surface38of the cylinder block30. In prior configurations the valves had to be outward from the top dead center position of the piston in order to receive the fluid forced out of the cylinder chamber37. As shown inFIG. 2, the tubular sleeve39extends partially over the opening between the cylinder chamber37and the bores in which the inlet and outlet check valves33and34are located, thereby extending the cylinder bore farther into the cylinder chamber37.

Referring again to both toFIGS. 1 and 2, a drive shaft40extends through the central shaft bore41and is rotatable therein being supported by a pair of bearings42. The center section of the drive shaft40within the cylinder block30has an eccentric cam44. The cam44has a circular outer surface, the center of which is offset from longitudinal axis25of the drive shaft40. As a consequence, as the drive shaft40rotates within the cylinder block30, the eccentric cam44rotates in an eccentric manner about the axis25of the drive shaft. As specifically shown inFIG. 1, a cam bearing46has an inner race47that is pressed onto the outer circumferential surface of the eccentric cam44and an outer race48. A plurality of rollers49are located between the inner race47the outer race48of the cam bearing. With the proper heat treatment and surface finishing, the surface of the eccentric cam44can serve as the inner bearing race. The cam bearing46improves the efficiency of the pump10over previous pumps that used a sliding journal bearing for this function. The rollers may be cylindrical, spherical, or other shapes.

A separate piston assembly51is slideably received within each of the cylinders36. Every piston assembly51has a piston52and a piston rod54. The piston rod54extends between the piston52and the cam bearing46. The piston rod54has a curved shoe56which abuts the outer race48of the cam bearing46. The curved shoe56is wider than the shaft of the piston rod, creating a flange portion. A pair of annular retaining rings58extends around the eccentric cam44engaging the flange portion of each curved shoe56, thereby holding the piston rods54against the cam bearing46, which is particularly beneficial during the intake stroke portion of a pumping cycle. The annular retaining rings58eliminate the need for a spring to bias the piston assembly51against the cam bearing46. The curved shoe56evenly distributes the piston load over a wide area of the cam bearing46. As the drive shaft40and eccentric cam44rotate within the cylinder block30, the outer race48of the cam bearing46remains relatively stationary. The outer race48rotates at a very slow rate in comparison to the speed of the drive shaft40and the inner race47. Therefore, there is little relative motion between each curved shoe56and the cam bearing's outer race48.

The piston52is cup-shaped having an interior cavity53which opens toward the drive shaft40. An end of the piston rod54is received within the interior cavity53and has a partially spherical head60that fits into a mating partially spherical depression62in the piston52. The head of the piston52may have an aperture50there through to convey hydraulic fluid from the cylinder chamber37to lubricate the interface between the spherical head60and the piston52. The piston rod54is held against the piston52by an open single bushing or a split bushing55and a snap ring57that rests in an interior groove in the piston's interior cavity53. The piston rod54follows the eccentric motion of the eccentric cam44and the piston52in turn follows by sliding within the cylinder36. The bushing and snap ring arrangement allows the spherical head60of the piston rod to pivot with respect to the piston52when a rotational moment is imposed onto the piston rod54by rotation of the eccentric cam44. Because of that pivoting, the rotational moment is not transferred into the piston52, thereby minimizing the lateral force between the piston and the wall of the cylinder36.

With continuing reference toFIG. 2, the drive shaft40includes an internal lubrication passage64extending from one end of the drive shaft40to the outer surface of the eccentric cam44. The lubrication passage64has a single opening in the outer surface of the eccentric cam44at the center of the eccentric apex of the cam44to feed fluid into the cam bearing46. The other end of the lubrication passage64opens into a chamber66at the end of the drive shaft40and that chamber receives relatively low pressure fluid through a feeder passage68from the inlet gallery31. As the drive shaft40rotates, centrifugal force expels fluid from the lubrication passage64into the cam bearing46. This action draws additional fluid into the lubrication passage64from the chamber66, thereby providing a pumping function for fluid that lubricates the cam bearing46. If the cam bearing46has an inner race47, that inner race has apertures that convey the lubricating fluid to the rollers49. The outer race48also has through holes to lubricate the shoes56of the piston rods54, thereby providing splash lubrication and eliminating a need to have the central shaft bore41filled with fluid. Not having the crankcase filled with fluid reduces windage drag on the eccentric cam44and improves efficiency of the pump. Additional lubricating passages59are provided to convey fluid from the central shaft bore41to the bearings42for the drive shaft40. The fluid used for lubrication exits the central shaft bore41through a standard drain port69from which the fluid is conveyed to a tank for the hydraulic system.

Pumping Operation

Rotation of the eccentric cam44causes each piston52to move cyclically within the respective cylinder36, away from the sealing cup24during a fluid intake phase and then toward the sealing cup24during a fluid exhaust phase. Because of the radial arrangement of the cylinders36, at any point in time, some pistons52are in the intake phase while other pistons are in the exhaust phase.

The piston52illustrated inFIG. 2is at a top dead center position when the volume of its cylinder chamber37is the smallest, which occurs at a transition point from the exhaust phase to the intake phase during each piston cycle. From this point, the outlet check valve34closes and further rotation of the eccentric cam44moves the piston52into the intake phase. During the intake phase, the volume of the cylinder chamber37increases, thereby initially decompressing the fluid remaining therein which tends to drive or put energy back into the drive shaft40. Thereafter, further increase in the cylinder volume produces a lower pressure in cylinder chamber37than in the inlet gallery31, therefore forcing the inlet check valve33open. Thus, fluid flows from the inlet gallery31through the inlet passage26and the inlet check valve33into the expanding cylinder chamber37. At this time, when there is a low pressure in the cylinder chamber37, the pressure in the outlet gallery32is higher due to either the flow output of the other cylinder chambers passing through a restriction or a static or dynamic load on the output. That pressure differential forces the outlet check valve34closed against its valve seat.

Thereafter, further rotation of the eccentric cam44moves the piston52into the exhaust phase during which the piston moves outward, away from the center axis25. That motion initially compresses the fluid in the cylinder chamber37, thereby increasing the pressure of that fluid. Soon the pressure in the cylinder chamber37is approximately that same as the pressure in the inlet passage26, at which point the associated spring closes the inlet check valve33. Eventually, the cylinder chamber pressure exceeds the pressure in the outlet gallery32and forces the outlet check valve34open, releasing the fluid from the cylinder chamber37into the outlet gallery and to the outlet port29.

When continued rotation of the eccentric cam44moves the piston52to the top dead center position shown inFIG. 2, the exhaust phase is complete and thereafter the piston transitions into the intake phase of another pumping cycle.

Because the inlet and outlet check valves33and34open and close almost immediately at the top dead center and bottom dead center positions, essentially the entire piston cycle is use to draw fluid into the cylinder chamber and then expel that fluid. This is in contrast to prior pumps that had throttle plates, but relied on the position of the piston to open and close an inlet opening into the cylinder. Those prior pumps had a dead region, which in some cases was one third the piston cycle, during which fluid was neither being drawn into nor expelled from the cylinder chamber. Thus with the present pump configuration an equivalent fluid volume can be pumped by each piston cycle with less piston stroke distance. This feature contributes to the compact size of the present pump.

Throttle Member Operation

With reference toFIGS. 2 and 3, the pump10includes a throttle mechanism that varies the inlet opening area from the shared inlet gallery31into the inlet passage26and through the inlet check valve33for each cylinder36during the intake phase. The throttle mechanism can take many forms, including a single spool with multiple lands or a series of spools or poppets; a cam or other device that limits the maximum opening of the inlet check valves33such that the inlet check valves33are also metering members; a nozzle-type restriction with a plate that moves axially rather than radially; or one or more electrically operated or pilot-pressure-operated valves associated with the cylinders36. One embodiment of the throttle mechanism, as shown inFIGS. 2 and 3, has a throttle member90and an abutting transition plate91that are sandwiched between two sections of the cylinder block30so as to extend across each of the plurality of inlet passages26. The throttle member90and the transition plate91have central apertures92and93, respectively through which the drive shaft40extends. The transition plate91is held stationary within the cylinder block30and has a plurality of transmission apertures94, each fixedly aligned with one of the inlet passages26. The throttle member90is rotatable around the drive shaft40and has a plurality of control apertures95proximate to the transmission apertures94in the transition plate91. The control apertures95of the throttle member90and the transmission apertures94in the transition plate91are formed on nearly the same radius as that of the inlet passages26, thus assuring registration of those apertures with the inlet passages upon rotation of the throttle member90through a predefined arc. As will be described, rotation of the throttle member90aligns and misaligns the control apertures95with the transmission apertures94, thereby creating variable orifices that control the fluid flow between the inlet gallery31and the cylinders36.

The pump10further includes a hydraulic actuator100for rotating the throttle member90within the cylinder block30. For that purpose, a tab98projects outward from the outer edge of the throttle member90and into an actuator bore102in the cylinder block30. The actuator bore102has a control port104to which a hydraulic conduit from a control circuit connects. A control piston108is slideably received in the actuator bore102and engages the tab98of the throttle member90. Pressurized fluid applied to the control port104drives the control piston108to the right in the actuator bore102(seeFIG. 3), thereby causing the throttle member90to rotate into different positions such as those shown inFIG. 4. Alternatively the hydraulic actuator100could include a rack and pinion type of arrangement; a rotary piston; or a worm gear with a hydraulic motor, an electric stepper motor, a linear solenoid, a rotary solenoid, or another similar electromechanical actuator.

The angular position of the throttle member90within the cylinder block30determines the alignment of the control apertures95in the throttle member with the transmission apertures94in the transition plate91. Varying that alignment alters the degree to which those apertures overlap and thus alters the cross sectional area through which fluid is able to flow between the inlet gallery31and the cylinders36during the piston cycle intake phase. In other words, the adjustable alignment of the transmission and control apertures94and95forms a variable orifice in that flow path provided by the inlet passages26. Both the control apertures95and the transmission apertures94may have unique shapes so that fluid flow varies in a specific manner to regulate the displacement of the pump10and maintain the output pressure at a desired level.FIG. 3illustrates the control apertures95and the transmission apertures94in a fully aligned orientation that provides the maximum flow between the inlet gallery31and cylinders36. As the throttle member90rotates counter clockwise and the transmission and control apertures94and95become misaligned to greater degrees, the area of that variable orifice initially changes at a relatively high rate until reaching the position depicted inFIG. 4. As the orifice area thereafter becomes smaller, the rate that the area changes decreases, i.e., the area changes more slowly for identical increments of change in the angular position of the throttle member90.

In one embodiment, the variation in the rate of orifice area change is determined by the unique shape of the transverse cross section of the control apertures95in the throttle member90. Transverse cross section as used herein means a cross section across a control aperture95in a plane that is transverse to the direction that fluid flows through the control aperture95. As shown inFIG. 3, each control aperture95has a transverse cross sectional shape that has an oval primary region96from which a tapered region97projects, like a beak of a bird, and terminates at an apex. The primary region96has a relatively large cross sectional area as compared to the cross sectional area of the tapered region97. The control apertures95can have other shapes and still attain variation of the rate of change of the fluid flow, as described herein. In other embodiments, the control apertures95do not vary the rate of change of fluid flow, and such rate of change remains constant no matter the angle of rotation of the throttle member90. Each transmission aperture94in the transition plate91has a size and shape which ensures that the entire cross sectional area of the associated control aperture95communicates with the inlet passage26when the throttle member90in the fully aligned position. That full alignment of the transmission and control apertures94and95enables the entire area of the control aperture95to conduct fluid through the throttle member90and thus provides the maximum flow of fluid from the inlet gallery31into each cylinder36during the intake phase of the piston cycle. A spring114biases the control piston108into a position in which the throttle member90is in the fully aligned aperture position.

From the fully aligned position inFIG. 3, application of pressurized fluid to the control port104drives the control piston108which acts on the tab98rotating the throttle member90counter clockwise. Continued motion eventually moves the throttle member90into an intermediate position depicted inFIG. 4. As the throttle member90moved between those positions the larger primary regions96of the control apertures95move over the edge of the transmission apertures94in the transition plate91, thereby closing off some of the area of each transmission aperture94. Because of the large size of the oval primary regions96, the area through which fluid flows through the orifice, created by the control apertures95and the transmission apertures94, diminishes at a relatively fast rate. That is, for a given incremental distance that the control piston108moves and thus for a given incremental angular change in throttle member90position, a relatively large change in flow occurs.

Upon reaching the intermediate position inFIG. 4, only the tapered regions97of the control apertures95remain aligned to communicate with the transmission apertures94in the transition plate91. Thus fluid can only flow through the throttle member90via those tapered regions97. In this intermediate position, the control apertures95are only partially aligned with the transmission apertures94in the transition plate91. Depending upon the amount of overlap in this intermediate position, the amount of flow between the inlet gallery31and each of the inlet passages26is reduced from the fully aligned position.

The amount of this flow can be proportionally controlled by controlling the rotational position of the throttle member90and thus the amount of that aperture overlap. As the rotation of the throttle member90continues, the tapered regions97cause the flow area to change at a smaller rate than occurred during previous motion to reach that intermediate position from the fully aligned position of the transmission and control apertures94and95. Now for each given incremental distance that the control piston108moves and for each given incremental angle change of the throttle member90, a relatively smaller change in flow area occurs than happened previously. Therefore, the rate that the open area of the control apertures95changes decreases as that open area becomes smaller.

Continued activation of the hydraulic actuator100results in the throttle member90eventually reaching a position in which the control apertures95are entirely misaligned with the transmission apertures94in the transition plate91. That is, no part of the control apertures95overlaps or opens into the transmission apertures94and fluid flow between the inlet gallery31and the cylinders36is blocked.

The use of a throttle member90to control the amount of flow between the inlet gallery31and the inlet passages26enables the displacement of the pump10to be dynamically varied. When the control apertures95are only partially aligned with the transmission apertures94, the amount of fluid flowing into the cylinder chamber37during the intake phase of each piston cycle is reduced. As a result, the piston52reaches bottom dead center without the cylinder chamber37being completely filled with hydraulic fluid. Thus, a portion of the total effective piston displacement is lost. The amount of lost displacement does not vary significantly as a function of the speed of the pump10, since the average pressure drop across the throttle member90is constant for typical pump speeds of 800 to 2500 RPM.

The present pump configuration with the rotatable throttle member90provides variable throttle choking at the input of each inlet check valve33. This has a significant advantage over a pump that has throttle choking at a single place for all the cylinders36, such as between the inlet port28and the inlet gallery31. With the per inlet check valve throttling arrangement of the present pump10, the fluid volume between the throttle member90and the inlet check valve33is relatively small and results in improved consistency and dynamic response in both starting and stopping fluid flow.

Although the above example shows and describes decreased output flow when pressurized fluid is applied to the control port104, it is also contemplated that a decrease in the pressure in the hydraulic actuator100could decrease output flow at the outlet port29, depending on configuration of the throttle member90with respect to the transition plate91and with respect to the hydraulic actuator100.

Pump Systems

FIG. 6depicts a pump system118. The pump system118has a piston pump10. As described herein above with reference toFIGS. 1 and 2, the pump10has a cylinder block30having an inlet port28, an outlet port29, and a plurality of cylinders disposed therein, each cylinder36in the plurality of cylinders being connected to the inlet port28by a respective inlet passage26in a plurality of inlet passages and to the outlet port29by a respective outlet passage27in a plurality of outlet passages. The piston pump10has a plurality of pistons, each piston52in the plurality of pistons being disposed in a respective cylinder36in the plurality of cylinders. The piston pump10has a drive shaft40driving the plurality of pistons52within the respective cylinders36. The pump10also has a throttle member90independently throttling flow in each inlet passage26in the plurality of inlet passages. The throttle member90may be like that shown and described inFIGS. 3 and 4, or may take other forms as described hereinabove. The pump system118further has a hydraulic actuator100moving the throttle member90to throttle flow in each in inlet passage26in the plurality of inlet passages. The hydraulic actuator100may include a control piston108and the pressure in the hydraulic actuator100acts on the control piston108to move the throttle member90. The pump system118further has a load sense apparatus124that modulates a pressure in the hydraulic actuator100, thereby governing movement of the throttle member90. The load sense apparatus124may include a margin spool126, the margin spool126being biased in a first direction shown by the arrow128, being moveable in the first direction128by a load sense signal LS in line130, and being moveable in a second, different direction (shown by the arrow132) against the bias and the load sense signal LS in line130by a pressure at the outlet port29, thereby modulating the pressure in the hydraulic actuator100as described further herein below. The margin spool126is biased for example, by a spring134.

In one embodiment of the pump system118, a user operates a control valve122to vary the rate at which fluid flows from the pump10to a hydraulic actuator120on a machine. This operation results in a pressure drop across the control valve122. The margin spool126is set to a predetermined bias force provided by a pre-load of the spring134. Pressure from an outlet port29acts on the non-spring end127of the margin spool126, and a load sense signal LS in line130(which in this example is pressure downstream of the control valve122) acts on the spring end125of the margin spool126. The position of the margin spool126will adjust to balance the predetermined force of the spring134and the two applied pressures, thereby modulating flow into or out of the hydraulic actuator100, more specifically through the control port104and into the actuator bore102. The flow into and out of the hydraulic actuator100either increases or decreases pressure in the actuator bore102, which in turn adjusts the output flow of the pump10by moving the throttle member90.

If the output flow of the pump10is lower than the operator-set desired flow rate, the margin spool126will shift in the direction of arrow128to allow flow out of the hydraulic actuator100through a drain connection152to a tank150. When fluid flows out of the hydraulic actuator100, the spring114moves in a direction that moves the throttle member90to increase the output flow of the pump10. The throttle member90rotates such that the control apertures95and the transmission apertures94are more aligned than they previously had been. The output flow of the pump10will increase until balance with the predetermined force of the spring134has been achieved. If the output flow of the pump10is greater than the operator-set desired flow rate, the margin spool126will shift in the direction of arrow132to allow flow from the outlet port29into the hydraulic actuator100. This moves the control piston108against the spring114in a direction that moves the throttle member90to decrease the output flow of the pump10. The throttle member90rotates such that the control apertures95and the transmission apertures94are less aligned than they previously had been. The output flow of the pump10will decrease until balance with the predetermined force of the spring134has been achieved. Other embodiments of load sense apparatuses that function based on a load sense signal LS in line130created by other than adjusting a restriction of a control valve122are contemplated within the scope of the present disclosure. For example, a load sense signal can be generated by sensing the highest load of the pump system118with a system of logic values or can be generated by an electrohydraulic device.

With further reference toFIG. 6, in one embodiment, the pump system118further includes a position sensor136sensing a position of the throttle member90or the control piston108. In a further embodiment, the pump system118further includes at least one pressure sensor137sensing a pressure at one or both of the inlet port28and the outlet port29.

Now with reference toFIG. 7, a pump system118having a pressure compensator valve138will be described. Like reference numbers inFIGS. 6 and 7describe like parts and will not be further described. In the embodiment ofFIG. 7, a pressure compensator valve138references a pressure at the outlet port29of the pump10and overrides modulation of pressure in the hydraulic actuator100by the load sense apparatus124if pressure at the outlet port29exceeds a predetermined limit. A first end140of the pressure compensator valve138references the pressure at the outlet port29of the pump10. A second end142of the pressure compensator valve138has a spring144that biases the pressure compensator valve138in a direction opposite the effect of the pressure from the outlet port29. During normal operation, the pump system118is controlled by the load sense apparatus124, as described herein above with reference toFIG. 6. The spring144biases the pressure compensator valve138in the direction of arrow141into a fully open position in which the load sense apparatus124modulates pressure in the hydraulic actuator100to increase or decrease flow from the pump10according to normal functioning of the load sense apparatus124. Should an operator ever request output pressure from the pump10that exceeds a predetermined force set by the spring144, the pressure compensator valve138shifts in the direction of arrow140. In this instance, pressure from the outlet port29overcomes the bias of the spring144and the pressure compensator valve138shifts in the direction of arrow140to allow flow directly from the outlet port29, through the pressure compensator valve138, and into the hydraulic actuator100. This moves the control piston108against the spring114in a direction that decreases the output flow of the pump10.

Either or both of the load sense apparatus124and the pressure compensator valve138shown inFIGS. 6 and 7can be implemented with the pump systems118shown inFIGS. 8-14, although only the load sense apparatus124is shown therein.FIG. 8shows a pump system118incorporating an electrohydraulic actuator146, whileFIGS. 9-14show pump systems118incorporating both an electrohydraulic actuator146and a load sense apparatus124in various configurations for controlling output flow of a pump10with either or both of the electrohydraulic actuator146and the load sense apparatus124.

Pump System Control Method

Now with reference toFIG. 5, an exemplary method for controlling an output flow of the pump10will be described. At block2, an input electric current i is provided by a control circuit148to an electrically operated actuator. The input electric current i, can be provided to an electrically operated actuator, such as for example an electrohydraulic actuator146, as will be described further herein below. At block4, the electrically operated actuator changes position according to the input electric current i. In one example, the electrohydraulic actuator146modulates pressure in a hydraulic actuator100based on the input electric current i. At6, a throttle member90changes position according to movement of the electrically operated actuator. In one example, the throttle member90moves according to the pressure in the hydraulic actuator100. At block8, an output flow from the outlet port29of the pump10corresponds to the position of the throttle member90, which in turn corresponds to the pressure in the hydraulic actuator100, which in turn corresponds to the pressure produced by the electrohydraulic actuator146, which in turn corresponds to the input electric current i.

Non-limiting exemplary systems for carrying out the method ofFIG. 5are described herein below with reference toFIGS. 8-13.

With reference toFIG. 8, the pump system118has an electrohydraulic actuator146governing movement of the throttle member90. The electrohydraulic actuator146modulates a pressure in the hydraulic actuator100, thereby governing movement of the throttle member90, as further described herein below. The pump system118may have a control circuit148controlling the electrohydraulic actuator146to thereby govern movement of the throttle member90. In one example, the control circuit148is an electronic control unit (ECU). In one example, the electrohydraulic actuator146is an electrically operated pressure control valve, which can be, for example, an electric pressure reducing valve. An operator inputs a desired flow rate of the pump system118into the control circuit148, which outputs an electronic signal to achieve this desired flow rate. The electrohydraulic actuator146receives the electronic signal from the control circuit148, and responds by moving into a position that increases or decreases pressure in the hydraulic actuator100. The electrohydraulic actuator146does so by removing or refilling hydraulic fluid from the tank150. The electrohydraulic actuator146exhausts fluid from the hydraulic actuator100through a drain connection152. The electrohydraulic actuator146refills the hydraulic actuator100via a pilot pressure source153. The pilot pressure source153maybe a separate pump as shown or may be taken directly from the outlet port29of the pump10.

In one example, the electronic signal is an electric current i. The electric current i corresponds to an output pressure of the electrohydraulic actuator146, therefore to a position of the control piston108within the hydraulic actuator100, and in turn to a position of the throttle member90. The position of the control piston108thereby yields a predictable output flow at the outlet port29based on this given electric current i, regardless of the speed of the drive shaft40or the pressure at the outlet port29. In other words, the combination of per inlet check valve throttling with a non-variable displacement pump allows for efficient control of a pump system118wherein a given electric current i produces a predictable flow at the outlet port29. This control can be accomplished without need for complex and expensive compensation methods, as is required for electrohydraulic control of variable displacement pumps.

When combined in a pump system118with a load sense apparatus124and/or pressure compensator valve138, the position and therefore function of the electrohydraulic actuator146can be varied to produce different outcomes, as discussed with reference toFIGS. 9-13.

FIGS. 9-10depict two systems in which pressure from an electrohydraulic actuator146can be added to a pump system118having a load sense apparatus124to limit the output flow of the pump10. In the embodiment ofFIG. 9, an electrohydraulic actuator146is inserted in series with a drain connection152of the margin spool126and selectively controls pressure in the drain connection152. When the electrohydraulic actuator146is not activated by an electric current i, the spool of the electrohydraulic actuator146is biased by a spring into a position that provides a relatively unrestricted path from the drain connection152to the tank150. In this state, the load sense apparatus124functions in response to the pump output pressure and the load sense signal LS in line130, in the same manner as described herein above with respect toFIG. 6, and modulates the pressure in the hydraulic actuator100to maintain the desired pump output pressure at the outlet port29. Alternatively, when the electrohydraulic actuator146is energized by the electric current i, the spool of that actuator moves to a position in which a pressure level, derived from the pressure at the pump outlet port29, is applied to the drain connection152. That pressure level is defined by the amount that the hydraulic actuator spool is moved by the electric current i. In this state, the drain connection152is not tied to the relatively low tank pressure. The pressure applied to the drain connection152sets a minimum pressure that can be supplied to the hydraulic actuator100and thus sets a maximum area opening position of the pump throttle member90, i.e., sets a maximum allowed alignment of the control apertures95and the transmission apertures94. Now as the load sense apparatus124responds to the pump output pressure and the load sense signal LS in line130, the pressure supplied to the hydraulic actuator100is modulated between the pump output pressure at the outlet port29and the minimum pressure level in the drain connection152.

In the embodiment ofFIG. 10, an electrohydraulic actuator146is inserted in series with an outlet145of the load sense apparatus124and the hydraulic actuator100. The electrohydraulic actuator146modulates the pressure in the hydraulic actuator100to a pressure level derived from pump output pressure at the outlet port29and dependent on the pressure in the outlet145of the load sense apparatus124and an electric current i. When the electrohydraulic actuator146is not activated by the electric current i, the spool of the electrohydraulic actuator146is biased by a spring into a position that provides a relatively unrestricted path from the outlet145of the load sense apparatus124to the hydraulic actuator100. In this state, the load sense apparatus124functions in response to the pump output pressure and the load sense signal LS in line130in the same manner as described hereinabove with respect toFIG. 6, and modulates the pressure in the hydraulic actuator100to maintain pump output pressure at the outlet port29. Alternatively, when the electrohydraulic actuator146is energized by the electric current i, the spool of the electrohydraulic actuator146is biased to a position in which the pressure level in the hydraulic actuator100is biased, due to the electric current i, to a level higher than the pressure in the outlet145of the load sense apparatus124. The pressure bias created by the electric current i applied to the electrohydraulic actuator146sets a minimum pressure that can be supplied to the hydraulic actuator100and thus sets a maximum area opening position of the pump throttle member90, i.e., sets a maximum allowed alignment of the control apertures95and the transmission apertures94. Now as the load sense apparatus124responds to the pump output pressure and the load sense signal LS in line130, the pressure supplied to the hydraulic actuator100is modulated between the pump output pressure at the outlet port29and the bias pressure due to the electric current i applied to the electrohydraulic actuator146.

In other words, in the embodiments ofFIGS. 9 and 10, the electrohydraulic actuator146and margin spool126create a minimum pressure that can be supplied to the hydraulic actuator100so as to set a maximum area opening position of the throttle member90. In the embodiment ofFIG. 9, the electrohydraulic actuator146modulates a pressure in the margin spool126by restricting flow from the margin spool126to a drain connection152, while in the embodiment ofFIG. 10the pressure in the hydraulic actuator100is a level of the pressure modulated by the load sense apparatus124plus a bias pressure level produced by the electrohydraulic actuator146.

Now with reference toFIGS. 11 and 12, a pump system118that hydraulically selects the higher pressure from the electrohydraulic actuator146and the load sense apparatus124and uses that pressure to control the hydraulic actuator100and thus the flow of the pump system118will be described. In other words, the load sense apparatus124modulates the pressure in the hydraulic actuator100unless a pressure produced by a flow from the electrohydraulic actuator146is greater than a pressure produced by a flow from the load sense apparatus124. The electrohydraulic actuator146modulates the pressure in the hydraulic actuator100if the pressure produced by the flow from the electrohydraulic actuator146is greater than the pressure produced by the flow from the load sense apparatus124.

An algorithm in the control circuit148may limit the maximum flow of the pump10such that the flow will not exceed a certain limit for a certain period of time. To achieve this maximum flow limit, the control circuit148outputs an electric current i that corresponds to a pressure output of the electrohydraulic actuator146, therefore to a position of the control piston108within the hydraulic actuator100, and therefore to a position of the throttle member90. The position of the control piston108thereby may yield a predictable maximum flow at the outlet port29, regardless of drive shaft40speed or pressure at the outlet port29.

If an operator-desired flow does not exceed the maximum flow limit set by the control circuit148, the pressure produced by the load sense apparatus124is therefore higher than the pressure produced by the electrohydraulic actuator146and the system operates under control of the load sense apparatus124. If the operator-desired flow exceeds the maximum flow limit set by the control circuit148, the load sense apparatus124attempts to gain additional flow from pump10by reducing the pressure in the hydraulic actuator100. At the point when the pressure produced by the load sense apparatus124falls below the pressure produced by the electrohydraulic actuator146, a valve will hydraulically change positions and the pressure in the hydraulic actuator100and thus flow at the outlet port29will be controlled by the electrohydraulic actuator146rather than by the load sense apparatus124. The algorithm of the control circuit148is therefore able to limit an operator's command for too much flow at the pump outlet port29, i.e., for flow that exceeds the maximum flow limit set by the control circuit148.

On the other hand, when the operator-desired flow once again falls below the maximum flow limit set by the control circuit148, the valve once again hydraulically changes positions, and the load sense apparatus124once more assumes control over flow at the pump outlet29.

The above-mentioned valve may be a check valve or a shuttle valve, although other valves could be used to achieve the same objective of hydraulically selecting the higher pressure of the electrohydraulic actuator146and the load sense apparatus124.

The pump system118ofFIG. 11includes a check valve154that selectively allows flow from the electrohydraulic actuator146to the hydraulic actuator100when the pressure produced by the flow from the electrohydraulic actuator146is greater than the pressure produced by the flow from the load sense apparatus124. When the system incorporates a check valve154, the flow produced by the electrohydraulic actuator146saturates the margin spool126to control the pressure in the hydraulic actuator100.

The pump system118ofFIG. 12includes a shuttle valve156that selectively allows flow from one of the electrohydraulic actuator146and the load sense apparatus124to the hydraulic actuator100. When the pressure produced by the flow from the electrohydraulic actuator146is greater than the pressure produced by the flow from the load sense apparatus124, the shuttle valve156shuts off the flow from the load sense apparatus124to the hydraulic actuator100. When the pressure produced by the flow from the electrohydraulic actuator146is less that the pressure produced by the flow from the load sense apparatus124, the shuttle valve156shuts off the flow from the electrohydraulic actuator146to the hydraulic actuator100.

Now with reference toFIG. 13, an alternative example of the pump system118will be described. In this example, the throttle member comprises first and second throttle members89,90. The load sense apparatus124governs movement of the first throttle member89based upon a load sense signal LS in line130, as described herein above with reference toFIG. 6. The electrohydraulic actuator146governs movement of the second throttle member90based upon an electronic signal, such as an electric current i, as described herein above with reference toFIG. 8. The hydraulic actuator in this embodiment comprises first and second hydraulic actuators100,101. The load sense apparatus124governs movement of the first throttle member89by modulating a pressure in the first hydraulic actuator100and the electrohydraulic actuator146governs movement of the second throttle member90by modulating a pressure in the second hydraulic actuator101. In the embodiment shown, the first throttle member89is located in series with the second throttle member90. The order of the two throttle members89,90can be reversed from that shown inFIG. 13.

During normal operation of the load sense apparatus124, the electrohydraulic actuator146will be de-energized and the second throttle member90will be fully open so as to provide a negligible amount of restriction into the cylinder chambers37. Only the first throttle member89restricts the flow into the cylinder chambers37based on the pressure generated by the load sense apparatus124. An algorithm in the control circuit148may limit the maximum flow of the pump10such that the flow will not exceed a certain limit for a certain period of time. When the algorithm determines that an operator-desired flow exceeds the maximum flow limit, the control circuit148energizes the electrohydraulic actuator146with an electronic signal, such as an electric current i. The electrohydraulic actuator146produces a pressure that rotates the second throttle member90to a position that corresponds to the electronic signal. The flow at the outlet port29then is controlled by the second throttle member90, until the operator-desired flow drops below the maximum flow limit. This causes the load sense apparatus124to produce a pressure in the first hydraulic actuator100that causes the position of the first throttle member89to be more restrictive than the position of the second throttle member90(which corresponds to the maximum flow limit set by the algorithm of the control circuit148).

By using both a load sense apparatus124and an electrohydraulic actuator146(and, in some embodiments, a pressure compensator valve138) within one pump system118, both the load sense apparatus124and the electrohydraulic actuator146can govern movement of the throttle member90by modulating a pressure in the hydraulic actuator100. Because per inlet check valve throttling with electrohydraulic control provides predictable output flow for a given electric current i, decoupled from pump outlet pressure and drive shaft speed as described above, it also allows for electrohydraulic control to override a load sense apparatus124without using specialized compensation methods and/or hardware to gain stability of the pump system118.

Now with reference toFIG. 14, a further example of the pump system118will be described. The pump system118of this example has a first hydraulic actuator100moving a throttle member90to throttle flow in each inlet passage26in the plurality of inlet passages. The load sense apparatus124governs movement of the throttle member90by modulating a pressure in the first hydraulic actuator100. An electrohydraulic actuator146governs movement of the throttle member90by limiting movement of the throttle member90, as will be described further herein below. The system118has a mechanical stop limiting movement of the throttle member90and a second hydraulic actuator101moving the mechanical stop, wherein the electrohydraulic actuator146moves the mechanical stop by modulating a pressure in the second hydraulic actuator101. In the embodiment ofFIG. 14, the mechanical stop is pusher pin158. The first and second hydraulic actuators100,101are located adjacent one another such that the second hydraulic actuator101is configured to move the pusher pin158into contact with a control piston108in the first hydraulic actuator100to thereby limit movement of the throttle member90.

FIG. 14therefore discloses an alternative to directly overriding control by the load sense apparatus124with a higher pressure produced by the electrohydraulic actuator146, as was described with reference toFIGS. 9-13. Instead, pressure produced by the load sense apparatus124and pressure produced by the electrohydraulic actuator146are isolated from one another in individual chambers (for example, hydraulic actuators100,101). Control by the load sense apparatus124is overridden by a pusher piston160having a pusher pin158controlled by pressure produced by the electrohydraulic actuator146. In this arrangement, the pressure produced by the electrohydraulic actuator146is fed to a second hydraulic actuator101with a large area ratio. The small end of the hydraulic actuator101is routed with a seal162into the actuator bore102of the first hydraulic actuator100and acts as a hard mechanical stop, which hard mechanical stop may be a pusher pin158. The pusher pin158in turn limits the flow of the pump10by acting as a mechanical stop past which the control piston108cannot go, thereby limiting the position of the throttle member90and thereby limiting flow. An operator may use the control circuit148to set a given pressure in the second hydraulic actuator101(corresponding to a maximum flow limit of the pump system118), which pressure may be produced by the electrohydraulic actuator146, to ensure that the control piston108can travel only a limited distance before it will hit the pusher pin158. If the operator commands more flow than the maximum flow limit set by the control circuit148, the pressure produced by the load sense apparatus124will decrease until the control piston108travel is eventually limited by the pusher pin158.

It should be understood that the pump systems118described herein above are not limited to control by pressure produced from a load sense apparatus124and an electrohydraulic actuator146, but rather can be controlled by an electrically operated actuator in place of the electrohydraulic actuator146. In one embodiment, the electrically operated actuator is a stepper motor. In other embodiments, the electrically operated actuator is a linear solenoid, a rotary solenoid, or any other electro-mechanical actuator.

In the foregoing description, certain terms have been used for brevity, clearness, and understanding. No unnecessary limitations are to be inferred therefrom beyond the requirement of the prior art because such terms are used for descriptive purposes and are intended to be broadly construed. The different configurations and systems described herein may be used alone or in combination with other configurations and systems. It is to be expected that various equivalents, alternatives and modifications are possible within the scope of the appended claims. Each limitation in the appended claims is intended to invoke interpretation under 35 U.S.C. §112, sixth paragraph, only if the terms “means for” or “step for” are explicitly recited in the respective limitation.